![116
Osama Shaukat
alternative thanks to its maturity, its environmentally
friendly characteristics, low maintenance and the fact
that the sun is everywhere. The residential and
building sector is in many countries the first energy
consumer, about 40% of the global energy
consumption[1], therefore efficient utilization and
control of energy consumption at household level is
crucial[2]. The challenge resides in the matching of
the intermittent energy production with the dynamic
power demand[3]. This needs to design an energy
management strategy in order to optimize the use of
PV source and storage and to match the local
production with local consumption. In the literature,
several works are particularly focused on optimizing
the energy. Consequently, studies have been directed
to "smart homes"[4-7]. In this paper the case of a home
powered by on grid photovoltaic system with storage
is investigated. An energy flow management
algorithm was developed. It manages the energy flow
available in the home through a combination of
switches to meet the energy demand optimally using
the energy produced by the photovoltaic generator
and/or the energy stored in the batteries and/or the
energy drawn from the grid. The purpose is to
minimize this last and ultimately reach the self-
consumption mode. Two particular weeks have been
chosen, one in summer and one in winter, in order to
tests the algorithm. The paper is organized as follows.
In Section 2, the chosen photovoltaic system which
feeds the home to satisfy the energy demand profile
is presented. The models used to simulate the PV
system are given in section 3. The proposed energy
flow management algorithm and the operating mode
of the system are presented in section 4, while in
Section 5, the obtained simulation results, using the
MATLAB-SIMULINK are given and interpreted.
2. PV System description and energy demand of
the home
2.1 Description of the proposed PV system
In order to study how to meet the energy needs
of a family by using mainly the energy produced by
a photovoltaic installation, the sizing of the PV system
has been done by using PVSST 1.0 locally developed
software[8]. The optimal PV system configuration
obtained is an array of 3.2kWp and a battery bank of
12kWh. To feed the loads of the house, 4kW DC/AC
converter has been chosen. In order to prevent the
batteries from a total discharge, the lowest level of
the State Of Charge (SOC) was set to 25%. This SOC
corresponds to a capacity of 3kWh. The SOCmax was
set to 85% which corresponds to 10 kWh. As shown
on figure 1, the PV system operation depends on the
combination status of four switches (K1, K2, K3, K4)
which is related to the energy demand, the energy
Fig. 1. Grid-connected PV system with storage for the solar home](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-6-320.jpg)
![The energy transfer/flow management algorithm for a photovoltaic solar home
117
available from the PV generator, the SOC of the
battery bank and the energy taken off from the grid.
In order to meet optimally the energy needs of the
home, an energy flow management algorithm has
been developed.
2.2 Solar home energy demand
The case treated in this paper is a house of 57 m2
area situated in the coastal region of Bou-Ismaîl
(Latitude: 36° 38' 33'' North and Longitude: 36° 38'
33'' North) occupied by a family of three members.
The energy demand of a home depends on many
factors such as the number of appliances, the electrical
power used by each appliance and the amount of use
of appliances determined by the behavior of the
occupants in the home.
Solar cell : A solar cell (also called photovoltaic
cell or photoelectric cell) is a solid state electrical
device that converts the energy of light directly into
electricity by the photovoltaic effect. The following
are the different types of solar cells.
Amorphous Silicon solar cell (a-Si)
Bio hybrid solar cell
Buried contact solar cell
Cadmium telluride solar cell (CdTe)
Concentrated PV cell (CVP and HCVP)
2.3 Home energy demand
To design the most efficient PV power system for
supplying the dwelling, the solar energy resources
of a site and the energy consumption of the household
should be known. Unfortunately, the electricity
demand is rarely available, so the daily load profile
must be generated. A home energy management
(HEM) system is an integral part of a smart grid that
can potentially enable demand response applications
for residential customers. This paper presents an
intelligent HEM algorithm for managing high power
consumption household appliances with simulation
for demand response (DR) analysis. The proposed
algorithm manages household loads according to
their preset priority and guarantees the total
household power consumption below certain levels.
A simulation tool is developed to showcase the
applicability of the proposed algorithm in performing
DR at an appliance level. This paper demonstrates
that the tool can be used to analyze DR potentials for
residential customers. Given the lack of under-
standing about DR potentials in this market, this work
serves as an essential stepping-stone toward
providing an insight into how much DR can be
performed for residential customers. In the studied
case, the home load profile has been determined by
considering the daily load profile of each appliance,
their power ratings and their estimated operating
time per day, then their summation leads to the daily
consumption home profile. For the summer, the daily
energy demand varies between 13.87 kWh and 16.27
kWh. For the winter, the energy demand varies
between 12.03 kWh and 18.30 kWh. For these two
seasons, the high energy demand is explained by the
use of energy intensive equipment for heating and
air conditioning.
3. PV system modeling
In this section, the modeling of the PV generator,
the batteries and the power converter will be
presented. For PV system modeling, a one diode
model for the PV generator has been used[10]. For the
batteries, the Copetti's model has been chosen as it is
well suited for lead acid batteries allowing the
simulation of the whole operating process charge-
discharge-overcharge while considering the
temperature change of the battery[11, 12]. For the DC/
DC converter, a boost converter controlled by a MPPT
Fuzzy logic Controller has been used[13, 14] and the
DC/AC converter has been modeled by its yield
curve.
Solar energy : Solar is scaling up rapidly, with
capacity more than trebling over the past four years.
New installations totaling more than 50 GW in 2015
took global solar PV power generating capacity to
231 GW by year end, a 28.1% increase versus the end
of 2014. Capacity has more than trebled in the past
four years. The largest increments in 2015 were
recorded in the Asia Pacific region; China added 15.2
GW, just ahead of Japan with 12 GW, together
accounting for more than half of the growth in global
solar capacity. The US provided the third largest
addition (7.3 GW). China's cumulative installed
capacity (43.5 GW) moved it ahead of Germany (39.7
GW), with Japan (35.4 GW) in third place, and the
US (25.6 GW) overtaking Italy (18.9 GW) to take
fourth place. Solar power generation enjoyed another
year of very rapid growth in 2015, with a 33%](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-7-320.jpg)
![The energy transfer/flow management algorithm for a photovoltaic solar home
119
"energizes" our computers, lights, refrigerators,
washing machines, and air conditioners, to name only
a few uses[1]. We use energy to run our cars and
trucks. Both the gasoline used in our cars, and the
diesel fuel used in our trucks are made from oil. The
propane that fuels our outdoor grills and makes hot
air balloons soar is made from oil and natural gas.
4.4 Renewable Energy and Trends in Solar
Photovoltaic Energy Production
4.4.1 Energy Scenario
The present energy scenario is discussed under
categorical division of World, Bangladesh.
4.4.2 World Energy Scenario
Global economic recession drove energy
consumption lower in 2009 - the first decline since
1982. World primary energy consumption - including
oil, natural gas, coal, nuclear and hydro power - fell
by 1.1% in 2009. The issues at hand are numerous
and include increasing atmospheric carbon dioxide
concentrations from greenhouse gas emissions,
environmental safety of energy production
techniques, volatile energy prices, and depleting
carbon based fuel reserves to name a few (Nguyen
and Pearce 2010; Choi et al. 2011)[5]. As a result,
countries are facing an increasing challenge to
diversify energy sources and bringing renewable
generation to the forefront of policy discussion.
4.3 Renewable and Non-Renewable Energy
All forms of energy are stored in different ways,
in the energy sources that we use every day. These
sources are divided into two groups --renewable (an
energy source that we can use over and over again)
and nonrenewable (an energy source that we are
using up and cannot recreate in a short period of
time). [2] Figure
Renewable and nonrenewable energy sources can
be used to produce secondary energy sources
including electricity and hydrogen. Renewable
energy sources include solar energy, which comes
from the sun and can be turned into electricity and
heat. Wind, geothermal energy from inside the earth,
biomass from plants, and hydropower and ocean
energy from water are also renewable energy sources.
However, we get most of our energy from non-
renewable energy sources, which include the fossil
fuels --oil, natural gas, and coal[2]. They're called fossil
fuels because they were formed over millions and
millions of years by the action of heat from the Earth's
core and pressure from rock and soil on the remains
(or "fossils") of dead plants and animals. Another
nonrenewable energy source is the element.
Uranium, whose atoms we split (through a
process called nuclear fission) to create heat and
ultimately electricity. We use all these energy sources
to generate the electricity we need for our homes,
businesses, schools, and factories. Electricity
Fig. 1.1. Renewable Energy Sources and Non-Renewable Energy Sources](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-9-320.jpg)
![120
Osama Shaukat
In the United States, a rise in renewable energy
generation has been supported by the availability of
federal tax credits and programs in individual states
(U.S. Energy Information Administration 2013a)[3].
Many states are implementing renewable portfolio
standards, or renewable energy standards, that
outline goals to increase electricity generation from
renewable resources (U.S. Energy Information
Administration 2013a). These policies seek to remove
barriers to install renewable generation and can
include grant programs, loan programs, and state
renewable electricity tax credits. The Database of State
Incentives for Renewables & Efficiency (DSIRE)
provides an outline of state renewable portfolio
standards available throughout the nation (NorthGraph 1.1 : Power generation capacity in
world by source, 2008](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-10-320.jpg)
![The energy transfer/flow management algorithm for a photovoltaic solar home
121
Carolina State University 2013)[8]. In 2012, about 12
percent of U.S. electricity was generated from
renewable sources (U.S. Energy Information
Administration 2013b). The United States Energy
Information Administration states that the five
renewable sources most often utilized include
biomass, water, geothermal, wind and solar (U.S.
Energy Information &!!
4.5 Energy transfer/flow management algorithm
The figure 3 depicts the detailed flowchart of the
flow energy management algorithm. The main idea
is that the photovoltaic production should be utilized
as much as possible to reach the self-consumption.
The priority is given to the supply of the loads, then
the charge of the batteries and at last, the excess
energy is fed into the grid. The algorithm operation
is based mainly on the PV production, the SOC of
the battery and the Ibat which indicates if the battery
is in charge or discharge and if it is able to feed the
loads or not. Depending on the state of theses
parameters one of the modes indicated in the
previous paragraph is used.
5. Simulation Results and Discussion
In order to test the efficiency of the flow energy
management algorithm the satisfaction of the energy
demand of the solar home by the photovoltaic system,
has been evaluated for two chosen cases. The first
one is 'a favorable week' for the PV production
relatively to weather conditions in summer and the
second one is 'an unfavorable week' in the winter.
For each case, depending on the PV produced and
the load energy profile a combination of the switches,
directs the energy flow to the given target. The
simulations were done using the irradiation and the
temperature data acquired by the meteorological
station of the site of Bou-Ismaïl. The PV electrical
production was calculated using the PV system model
presented briefly in paragraph 3. The algorithm was
performed using the appropriate load profile of each
season. The figures 4 and 5 show the states of the K1,
K2, K3 and K4 switches and the variation of the
photovoltaic production, the energy demand (load
profile), the energy battery storage and the grid
energy for the summer and winter week respectively.
From the results presented in figure 4 and 5, the
excess energy produced in the favorable week in
summer is 8.5%, this percentage corresponds to an
energy of 9kWh which is injected to the grid. On the
other hand, in winter, 67% of the energy demand of
the unfavorable week is satisfied by the grid (65kWh)
this is due to the low photovoltaic energy production
during this season. To better explain the energy flow
in the studied PV system, two particular cases will
be discussed. The first one corresponds to a favorable
day in summer (25/08/2015), and the second
corresponds to an unfavorable day in winner (03/
02/2015).
As it can be seen on figure 6, between midnight
and 5am there is no PV production and the energy
demand is very low, the battery is at its low level
(25%). Between 5 am and 5 pm, the PV array produces
2kWh. After 5 pm, as the PV production is low, the
batteries supply a part of the demand. Between 7 pm
and 9 pm all the energy needed is taken from the
batteries. After 9 pm the batteries continues to
discharge in the loads but the demand is so high that
the rest of the energy needed was withdrawn from
the grid. Indicates that, for the chosen day it can be
noticed that the irradiation is at low level all the day
and this conducts to a low PV production. In this case,
the energy demand is exclusively satisfied by the grid.
The battery bank is all day long near its lowest level
(25%). The highest energy demand is situated
between 5 pm and 9 pm.
6. Conclusion
Thanks to the decreasing of the cost of PV
installations, the grid parity comes a reality.
Nowadays, it is not profitable as before to sale the
produced photovoltaic electricity. It is more
interesting for the PV owners to use it for their own
needs and try to reach their self-consumption. In this
article, as a first step an energy flow management
algorithm which directs the energy flow according
to priorities set was presented. It appears that even
if the energy flow algorithm is necessary it is not
sufficient. To go further and to use the PV electricity
produced as much as possible, a Home Energy
Management (HEM) is needed. This last will allow
the balancing between the demand and the
generation, by controlling the deferrable loads,
reduces the energy consumption for example by using](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-11-320.jpg)
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Osama Shaukat
natural light as long as possible, exchanging excess
energy produced between neighbors instead of
drawing off energy from the grid.
7. Acknowledgements
We would like to thank Dr Shahzad Ahmad
Siddiqui, for his valuable advices on the writing of
this research paper.
References
[1] OECD/IEA, 2013. Transition to sustainable
buildings. Strategies and opportunities to 2050.
[2] Asare-Bediako B., Ramirez Elizondo L.M,
Ribeiro P.F. and Kling W.L., Consideration of
Electricity and Heat Load Profiles for
Intelligent Energy Management Systems.
UPEC 2011. 46th International Universities' Power
Engineering Conference. Soest. Germany; 5-8th
September 2011.
[3] Riffonneau Y., Bacha S., Barruel F. and Ploix
S., Optimal Power Flow Management for Grid
Connected PV Systems with Batteries. IEEE
Trans. Sustain. Energy. 2(3), (2011).
[4] Gudi N., Wang L. and Devabhaktuni V., A
demand side management based simulation
platform incorporating heuristic optimization
for management of household appliances. Int J
Elec Power, 43(1), (2012) 185-193.
[5] Di Giorgio A. and Pimpinella L., An event
driven smart home controller enabling
consumer economic saving and automated
demand side management. Appl Energ, 96,
(2012) 92-103.
[6] Chen X., Wei T. and Hu S., Uncertainty-aware
household appliance scheduling considering
dynamic electricity pricing in smart home.IEEE
Trans. Smart Grid, 2, (2013) 932-941.
[7] Tascikaraoglu A., Boynuegri A.R. and
Uzunoglu M., A demand side management
strategy based on forecasting of residential
renewable sources: A smart home system in
Turkey. Energ Buildings, 80, (2014) 309-320.
[8] Chikh M., Mahrane A.and Bouachri F., PVSST
1.0 sizing and simulation tool for PV systems.
Energy Procedia, 6, (2011) 75-84.
[9] Al-Alawi A. and Islam S.M., Demand side
management for remote area power supply
systems incorporating solar irradiance model,
Renew. Energ., 29, (2004) 2027-2036.
[10] Sera D., Teodorescu R. and Rodriguez P., PV
panel model based on datasheet values. Proc.
IEEE Int. Symp. Ind. Electron. http://
dx.doi.org/10.1109/ISIE.2007.4374981 (2007)
2392-2396.
[11] Copetti J.B. and Chenlo F., Internal resistance
characterization of lead acid battery for PV
rates. Proceedings of the 11th European PV Solar
Energy. Conference, Monteux, (1992) 1116-1119.
[12] Copetti J.B. and Chenlo F., A general battery
model for PV system simulation. Prog
Photovoltaic Res. Appl. (1993) 283-292.
[13] Chekired F., Mellit A., Kalogirou S.A. and
Larbes C., Intelligent maximum power point
trackers for photovoltaic applications using
FPGA chip: A comparative study. Elsevier, Sol.
Energy, 101, (2014) 83-99.
[14] Chekired F., Larbes C., Rekioua D. and
Haddad F., Implementation of a MPPT fuzzy
controller for photovoltaic systems on FPGA
circuit. Elsevier, Energy Procedia. 6, (2011) 541-
549.](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-12-320.jpg)
![Numerical analysis of friction factors in smooth and rough microchannels
123
Numerical analysis of friction factors in
smooth and rough microchannels
MOHD OWAIS QIDWAI1, M.M. HASAN1 and MOHD ARIZ2
1Department of Mechanical Engineering, F/o Engineering and Technology, Jamia Millia Islamia, New Delhi, India
2Global Evolutionary Energy Design, New Delhi, India
*E-mail: owais00738@gmail.com
Abstract
Based on the available literature, the researchers are divided that transition in smooth channel
occurs in low Reynolds region or not, and various experimental results contradicts each other.
Also, the effect of surface roughness has any considerable effects on early transition behavior or
not. The numerical procedures are used these days and various researches are ongoing with the
help of numerical methods on studying the flow characteristics. The author aims in understanding
the flow behavior in low Reynolds regime between 100 - 1400 for smooth and rough channel
with roughness to hydraulic diameter ratio of 0.264. The flow remains in laminar regime for
smooth channel and rough channel, and no significant effect of roughness obtains in both the
cases. The early transition effects are not seen in the analysis. The friction factor decreases with
increase in Reynolds number.
Key words : CFD, Friction Factor, Laminar, Reynolds Number, Single phase, Transition,
Turbulent, Velocity profile.
1. Introduction
For successful operation of electronic devices, the
microchannel heat sink is best available option for
decades. As per Garimella et al.[1] liquid cooling has
lot of potential as discussed in the TCO decision
criterion for high density and high computing
electronic devices. In the development of liquid
cooling microchannel heat sink various researches
regarding friction factor[2,3], pressure drop[4] and
thermal resistance[5] has been published in recent
years.
Most of the investigation performed in earlier
phase were experimental, but recent development in
mathematical models open ways for numerical
studies and consistent results were obtained by
various researchers[6,7,8], proving the validity of the
numerical simulations.
Mala and Li[9] performed experimental studies
on circular cross sectional microchannel of stainless
steel, with one channel with hydraulic diameter in
Invertis Journal of Renewable Energy, Vol. 7, No. 3, 2017 ; pp. 123-130
the range of 50-254. The ratio of surface roughness to
the hydraulic diameter in the channel varies from
0.69-3.5. They found that for Reynolds number 80 -
2500, the critical Reynolds number obtained was in
the range of 300-900. As per their observation early
transition is obtained with increase in friction factor.
Brackbill and Kandlikar[10] also obtained early
transition from laminar to turbulent regime in the
range of 572-2367, for introducing saw tooth profile
roughness in the microchannel.
Many authors published results with higher
range of critical Reynolds number. Tam et al.[11,12]
obtained critical Reynolds number in the range of
2100-2700. Baviere et al.[13] found critical Reynolds
number in the range of 2417- 2822.
Based on the literature review, and applicability
of numerical procedures to the study of microchannel
flow characteristics, the authors aims at finding out
the region of transition in the microchannel with the
help of friction factor obtained theoretically and
DOI No. : 10.5958/2454-7611.2017.00017.0](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-13-320.jpg)
![124
Mohd Owais Qidwai, M.M. Hasan and Mohd Ariz
through numerical procedures for Reynolds number
varying from 100-1400, for single phase water flow
in the microchannel with constant thermos-physical
properties.
Classical conventional theory from open
literature consists of relations of friction factor (Darcy
and Fanning) with the pressure drop. The fanning
friction is defined as
21
2
= w
Ff
u
(1)
Darcy friction factor and fanning friction factor
are related as
4=D Ff f (2)
And Darcy friction factor in terms of pressure
drop is given by
2
2
= H
D
P D
f
L u
(3)
Where P is the pressure drop across the length
of microchannel.
The Poiseuille number (Po) is defined as the
product of Darcy friction factor and Reynolds number
(Re), where Reynolds number is defined as,
Re = HuD
(4)
The friction factor obtained in the above relation
is evaluated based on assumption that Steady,
laminar, incompressible flow takes place. The
Poiseuille number on the other hand is obtained for
the same assumption along with fully developed flow
between two parallel infinite plates. For a finite
dimension rectangular cross sectional microchannel,
Poiseuille number is defined based on the aspect ratio
by Shah and London[14].
2 3
4 5
= 96(1 – 1.3553 + 1.9647 – 1.7012
+ 0.09564 – 0.2537 )
Po
(5)
Various experimental investigations have
confirmed that microscale friction factor can be
obtained by macroscale theories. Therefore Navier-
Stokes equation can be used to study incompressible,
Newtonian liquids in laminar regime[15]. Several
numerical studies have been performed earlier which
confirm the applicability of numerical schemes
available for thorough analysis Xu et al.[16], Lee et
al.[17], Jones et al.[18].
The present paper aims at investigating the effect
of laminar to turbulent transition with the increase
in Reynolds number for smooth and rough channel
with roughness to hydraulic diameter ratio of 0.264.
The numerical scheme is validated with the
experimental work of Hao et al.[19].
2. Mathematical Model
Figure 1 Shows the schematic diagram of smooth
channel with the dimensions as used in Hao et al.[19].
The average surface roughness in the experimentation
was obtained 0.032 micron, on the inside surface of
the channel.
To study the entire domain is computationally
expensive process and hence single channel with
exact dimensions are considered for the analysis.
For the rough channel as shown in Figure 2, the
ribs are introduced only on one side of the channel to
analyse its effect on the friction factor and transition
regime. The first roughness element is introduced at
x= L1=6.8 mm from the inlet boundary and thereafter
a pitch value of 2.95 mm is chosen to place the ribs,
so that L2= L3= L4= L5= L6= 2.95 mm. Length L1 is
selected with a higher value so that fully developed
flow is obtained in the region. The ribs have cross-
sectional dimension of hr = wr = 50 ?m are introduced
Fig. 1. Schematic Diagram of Smooth Microchannel
Fig. 2. Ribs of square cross section with hr = wr are
introduces in the smooth channel as roughness element](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-14-320.jpg)
![126
Mohd Owais Qidwai, M.M. Hasan and Mohd Ariz
3. Numerical Solution
Ansys CFX is used to solve the continuity and
momentum equation using Finite volume method
(FVM). A mesh of fluid domain is generated for the
channel with dimension of length L = 23 mm; height
h = 0.146 mm andwidth w = 0.027 mm. The
dimensions considered are same as that of Hao et
al.[19] for the smooth channel with no ribs. The
numerical solution also involves surface roughness
effect of 0.032 micron as mentioned by Hao et al.[19].
For a flow of Re = 100 three mesh sizes were used as
shown in Table 1. The number of nodes are increased
for each solution and the resultant pressure drop from
inlet to outlet boundary is mentioned along with the
central velocity in fully developed region about 18
mm down the inlet boundary. The convergence
criterion adopted for the solution of momentum and
continuity equations is to be less than 10-6.
4. Result and Discussion
From Fig. 6, it is observed that frictional factor
decreases with increase in Reynolds number for both
smooth and rough microchannel. There is no
significant deviation from Darcy friction factor and
numerical frictional factor for both smooth and rough
microchannels. However, the slight deviation starts
taking place at Re=600 for friction factor in rough
microchannel, which is not so significant for which
early transition effect in rough channel can be
concluded. Though a higher value of friction factor
is obtained for Re > 900 in case of rough channel. But,
it can be said that there is no effect of surface
roughness on transition for flow with low Reynolds
number.
From Fig. 7 it appears that the velocity profile is
parabolic for low Reynolds number, and as the
Reynolds number is increased the parabolic velocity
Table 1. Grid independence study
Mesh Number of Pressure Drop Fully developed
Configuration Nodes P (Pa) central velocity (m s–1
)
1 429229 8482.00 0.955027
2 576000 8478.65 0.950493
3 792000 8474.30 0.948219
Fig. 6. Comparison of Friction Factor obtained using pressure drop for smooth and rough microchannel from
numerical solution and C/Re., where C is obtained from equation 5 for smooth channel](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-16-320.jpg)
![Numerical analysis of friction factors in smooth and rough microchannels
129
Fig. 11. Flow Recirculation in rough
channel for Re=600
Fig. 12. Flow Recirculation in rough
channel for Re=1300
5. Conclusion
Based on the above discussion, following points may
be concluded:
1. Roughness to hydraulic diameter ratio of
0.264 has no significant effect in early
transition from laminar to turbulent regime
in the range of Re = 100 to 1400.
2. The friction factor has higher value for
Re>900 for both the channels, but based on
streamwise velocity, it cannot be asserted that
profile obtained in Fig. 9 & Fig. 10, is not
parabolic in nature.
3. Conventional theories are in good agreement
with experimental results mostly. Only in the
studies where entrance loss, exit loss and
developing flow are not considered, reports
deviation from it [2].
Nomenclature Subscript
Shear Stress w wall
Fluid Density F Fanning Factor
¯u Average flow velocity D Darcy Factor
f Friction Factor H Hydraulic
Difference
P Pressure
D Diameter
L Length
Viscosity
aspect ratio (h/w)
h Height of Channel
w Width of channel
References
[1] Garimella S.V., Persoons T., Weibel J.and Yeh
L.T., "Technological drivers in data centers and
telecom systems: Multiscale thermal, electrical,
and energy management," Appl. Energy, 107,
(2013) 66-80.](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-19-320.jpg)
![130
Mohd Owais Qidwai, M.M. Hasan and Mohd Ariz
[2] Dai B., Li M. and Ma Y., "Effect of surface
roughness on liquid friction and transition
characteristics in micro- and mini-channels,"
Appl. Therm. Eng., 67(1-2), (2014) 283-293.
[3] Steinke M.E. and Kandlikar S.G., "Single-
phase liquid friction factors in microchannels,"
Int. J. Therm. Sci., 45(11), (2006) 1073-1083.
[4] Gunnasegaran P., Mohammed H. and Shuaib
N.H., "Pressure Drop and Friction Factor for
Different Shapes of Microchannels," 7-8, (2009).
[5] Qu W. and Mudawar I., "Experimental and
numerical study of pressure drop and heat
transfer in a single-phase micro-channel heat
sink," Int. J. Heat Mass Transf., 45(12), (2002)
2549-2565.
[6] Lee P.S. and Garimella S.V., "Thermally
developing flow and heat transfer in
rectangular microchannels of different aspect
ratios," Int. J. Heat Mass Transf., 49(17-18),
(2006) 3060-3067.
[7] Li J. and Peterson G.P., "3-Dimensional
numerical optimization of silicon-based high
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with liquid flow,"Int. J. Heat Mass Transf.,50(15-
16), (2007) 2895-2904.
[8] Sahar A.M., Özdemir M.R., Fayyadh E.M.,
Wissink J., Mahmoud M.M. and Karayiannis
T.G., "Single phase flow pressure drop and heat
transfer in rectangular metallic microchannels,"
Appl. Therm. Eng., 93, (2015) 1324-1336.
[9] Mohiuddin Mala G. and Li D., "Flow
characteristics of water in microtubes," Int. J.
Heat Fluid Flow, 20(2), (1999) 142-148.
[10] Brackbill T.P. and Kandlikar S.G., "Effect of
Sawtooth Roughness on Pressure Drop and
Turbulent Transition in Microchannels," Heat
Transf. Eng., 28(8-9), (2007) 662-669.
[11] Tam L.M. et al., "The Effect of Inner Surface
Roughness and Heating on Friction Factor in
Horizontal Micro-Tubes," 44403, (2011) 2971-
2978.
[12] Tam L.M., Tam H.K. and Ghajar A.J.,
"Simultaneous Heat Transfer and Pressure
Drop Measurements for a Horizontal Micro-
Tube," 38921, (2011) T10122-T10122-8.
[13] Conference I., June M. and York N.,
"ICMM2004-2338," 0, (2016) 1-8.
[14] Shah R.K. and London A.L., Laminar Flow
Forced Convection in Ducts: A Source Book for
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[15] Xu B., Ooti K.T., Wong N.T. and Choi W.K.,
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[16] Lee P.S., Garimella S.V. and Liu D.,
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(2005) 1688-1704.
[17] He Y., Shao B.D., Cheng H.M. and Tang Y.J.,
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J. Micromech. Microeng, 16, (2006) 1397-1402.](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-20-320.jpg)
![Effect of operating conditions on the performance of vapor absorption refrigeration system
131
Effect of operating conditions on the performance
of vapor absorption refrigeration system
MD. MERAJ1, RASHID IMAM2 and MD. ASHFAQUE ALAM3
1
Engineering & Technology, Department of Mechanical Engineering, JMI, New Delhi-110025, India
2
Department of Mechanical Engineering, Vishveshwarya Group of Institution, Noida, UP-203207, India
3
Department of Mechanical Engineering, NIT, Jamshedpur, Jharkhand-831014, India
*E-mail: md.meraj1221@gmail.com
Abstract
In this paper, a thermodynamic analysis of single effect vapor absorption refrigeration system
(VARS) is carried out. LiBr-Water is used as working fluid during the analysis. This analysis is
carried out to study the effect of operating conditions on the performance of vapour absorption
refrigeration system. The generator temperature, the condenser temperature and the evaporator
temperature are taken as operating conditions for the analysis. From the result and discussion,
it is found that the coefficient of performance (COP) of the system increases with increasing the
generator and evaporator temperature up to a certain limit but decreases with increasing the
condenser temperature.
Key words : VARS, LiBr-Water, COP, generator temperature, condenser temperature, and
evaporator temperature.
1. Introduction
Present day, the rate at which the non-renewable
energy sources are being utilized has become a major
concern for the world. The future prospect of non-
renewable sources of energy like coal, natural gases
and petroleum products are going to be scarce. To
preserve these energy sources for our future
generation, we have to reduce its consumption by
using alternative sources of energy like solar energy,
geothermal energy, waste heat, hydro-energy etc.
There are many ideas has been developed to utilized
these alternative sources of energy and researches has
also been under the processes to improve the systems
which utilize these sources. In this regard vapor
absorption refrigeration systems are the best
alternative system to reduce the rate of consumption
of non-renewable energy sources.
Vapor absorption refrigeration system (VARS)
invented much earlier than vapor compression
refrigeration system (VCRS) but due to law coefficient
of performance and inven-tion of high performance
Invertis Journal of Renewable Energy, Vol. 7, No. 3, 2017 ; pp. 131-136
refrigerant for VCRS during Second World War, the
popularity of VARS decreases. But, in present
scenario of energy crisis has forced the world to use
vapor absorption refrigeration system in place of
vapor compression system to reduce the rate of
consumption of non-renewable energy sources.
VARS is one of the best replacement for the VCRS,
from the energy point of view as well as environ-
mental,because this system works on law grade of
energy such as solar energy, geothermal-energy,
waste-heat from industries, cheap available energy
etc., and VARS uses such type of refrigerant-absorbent
combination which do not contributes in global
warming and ozone depletion like LiBr2-H2O,
H2O -NH3, LiNO3-NH3, NaSCN-NH3 etc. From
construction point of view absorption system and
compression system are similar but only the
differenceis that the compressor in case of
compression system is replaced by the generator-
solution heatexchanger-pump-absorber assembly
which is used to circulate the working fluid entire
the system by creating pressure differ-ences[8-9].
DOI No. : 10.5958/2454-7611.2017.00018.2](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-21-320.jpg)
![132
Md. Meraj, Rashid Imam and Md. Ashfaque Alam
Aphornratna. S and Sriveerakul. T[1] performed
an experimental investigation of single-effect
absorption. They have usedaqueous lithium-bromide
as working fluid. They have done their analysis on
2 kW cooling capacity experimental refrigerators
which was tested with various operating
temperatures. From results, it was found that the
system performance is strongly dependent on
thesolution circulation ratio (SCR)and also on
solution heat exchanger. They have resulted that
energy input to the generator is reduced to 60% by
usingsolution heat exchanger. Darwish A. N. et. al.[2]
presented the performance analysis and evaluation
of the roburabsorption-refrigeration water-ammonia
(ARWA) system.The system is analysed by using
aspen plus flowsheet simulator.They analysed COP,
heat duties of the evaporator, absorber, and the
condenser and the flow rate of refrigerant passing
through the evaporator.The results obtained by them
is compared with some manufacturer data and
experimental data reported in the journal literature.
Saghiruddin and Siddiqui[3] perfor-medthe economic
analysis of ordinary and evacuated tubular type flat-
plate collectors for operating absorption cycles with
and without heat recovery absorber. Water-
Ammonia, NaSCN3 and LiNO3-NH3 mixtures has
been selected as the working fluids in the cycles. Use
of heat recovery absorber, in addition to the primary
absorber in the conventional absorption cycles lead
to the improvement in COP by 20-30 % in Water-
Ammonia and 33-36% in NaSCN-NH3 and LiNO3-
NH3 mixtures. Kaushik and Kumar[4] performed
thermodynamic analysisof two-stage vapour
absorption refrigeration system using ammonia-
water (NH3-H2O) and ammonia-lithium nitrate
(NH3-LiNO3) as working fluid. They havefound that
minimum evaporator temperature is achieved in
NH3-LiNO3 combination, and COP of the NH3-
LiNO3 is higher than the NH3-H2O. Marcos et al.[5]
performed optimization of COP in single and double
effect LiBr-H2O absorption chillers. They analyzed
water cooling system and air cooling system and
found that solution concentration is an important
parameter to improve COP. With this, they also
found the crystallization limit. Their study showed
that COP of single effect becomes 0.85 to 0.74 in water
cooled and 0.72 to 0.65 in air cooled system. Similarly,
in double effect cycle it was 1.18-1.2 in the water
cooled and 1.15-1.07 in the air cooled system. Talbi
M. M. and AgnewB.[6] presented an exergy analysis
on a single effect absorption refrigeration cycle with
lithium-bromide-water as the working fluid pair.
They calculated the loads on each component of
absorption system, exergy of each components and
total exergy of system. However, some of the research
has been done on the energy and exergy analysis of
the double effect absorption refrigeration system[10,
11,12].
From literature survey it is absorbed that there
is no work has been done on the effect of operating
condition on the performance of vapour absorption
refrigeration system. Therefore this paper present a
thermodynamic analysis which shows variation of
performance of VARS with the operating conditions
like the generator temperature, the condenser
temperature and the evaporator temperature in the
presented analysis.
2. System Description
A schematic diagram of vapour absorption
refrigeration system is shown in Fig. 1. on which
analysis has beendone. This system consists of an
Evaporator (E), Absorber (A), Generator (G),
Condenser(C), Throttle valve (TV), Pre-cooler (PC),
Fig. 1. Schematic diagram of single effect vapour
absorption refrigeration system.](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-22-320.jpg)
![Effect of operating conditions on the performance of vapor absorption refrigeration system
133
Preheater (PH) and Pump.The operating temperature
of generator, condenser, evaporator and absorber is
shown byTg, Tc, Te and Ta respectively in figure 1.
There are two pressure levels in this system; the
generator and condenser operate at high pressure
level while evaporator and absorber operate at low
pressure level. LiBr-H2O mixture is used as working
fluid in the system in which LiBr works as an
absorbent and H2O works as refrigerant. LiBr-H2O
mixture circulate in the generator-solution heat
exchanger-pump-throttle valve-absorber assembly,
generating the water vapor in the generator which
flows through condenser, throttle valve, solution heat
exchanger and evaporator. The working fluid having
the different state condition while circulating inthe
system is shown in figure by different state point.
3. Thermodynamic Analysis
The thermodynamic analysis of single effect
vapour absorption refrigeration system is based on
the application of mass, concentration and energy
conservations. For performing these applications,
there are some assumptions are made to simplify the
analysis, these are as follows :
The systems are in steady state.
The refrigerant (water) at the outlet of the
condenser is saturated liquid.
The refrigerant (water) at the outlet of the
evaporator is saturated vapour.
Temperature of absorber and condenser are
same.
Expansion process in throttle valve is
isenthalpic.
Heat loss to environment and pressure drop
in the systems are negligible.
No leakage of air in the system.
Mass, concentration and energy conservation
equation of generator are respectively expressed by:
5 6 10 m m m (1)
5 5 6 6 10 10 m X m X m X (2)
6 6 10 10 5 5= –EQ m h m h m h (3)
Mass, concentration and energy conservation
equation of absorber are respectively expressed by:
3 2 12m m m (4)
3 3 2 2 12 12m X m X m X (5)
2 2 12 12 3 3= –aQ m h m h m h (6)
Energy consumption of solution pump can be
defined as:
4 4 3 3= –pW m h m h (7)
Energy conservation of condenser and
evaporator is given by :
6 6 7 7= –cQ m h m h (8)
1 1 9 9= –eQ m h m h (9)
The coefficient of performance of the system can
be expressed as follows :
/( )e E DCOP Q Q W (10)
A computer program has been developed
using FORTRAN language to carry out the system
simulation of the single effect vapour absorption
refrigeration system. The analysis has been done for
1 Ton of refrigeration. Properties of LiBr-water have
been evaluated by the correlation from[7]. The
calculation parameters for the simulation are
summarized in table 1.
Table 1
Calculation Parameters
Evaporator temperature (Te in °C) 5-(2.5)-12.5
Absorber temperature (Ta in °C) 30-(5)-40
Condenser Temperature (Tc in °C) 30-(5)-40
Effectiveness of pre-heater (%) 75
Effectiveness of pre-cooler (%) 75
Cooling load (kJ/h) 12600
4. Results and Discussion
Figure 2 shows the variation of coefficient of
performance (COP) of single effect vapour absorption
system with genera-tor temperature at which heat
is supplied at different values of evaporator
temperature (i.e. Te = 5.0, 7.5, 10.0 and 12.5 °C), the
condenser temperature and absorber temperature at
which heat is rejected to the surroundings are equal
and kept constant i.e. Ta = Tc = 30 °C. From this graph,
it is seen that the coefficient of performance of single
effect increases drastically from low values at low
generator temperature, reach to maximum value and](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-23-320.jpg)
![136
Md. Meraj, Rashid Imam and Md. Ashfaque Alam
References
[1] Aphornratana S. and Sriveerakul T.,
"Experimental studies of a single-effect
absorption refrigerator using aqueous lithium-
bromide: Effect of operating condition to
system performance", Experimental thermal and
fluid science, 32, (2007) 658-669.
[2] Darwish N.A., Al-Hashimi S.H. and Al-
Mansoori A.S., "Performance analysis and
evaluation of a commercial absorption-
refrigeration water-ammonia (ARWA) system",
International Journal of Refrigeration, 31, (2008)
1214-1223.
[3] Saghiruddin and Siddiqui M.A., "Economic
analysis and performance study of three
ammonia-absorption cycles using heat
recovery absorber", Energy Conversion and
Management, 37, (1996) 421-432.
[4] Kaushik S.C. and Kumar R., "Thermodynamic
study of two stage vapour absorption
refrigeration system using NH3 refrigerant
with liquid solid absorbents", Energy Conversion
and Management, 4, (1985) 427-431.
[5] Marcos J.D., Izquierdo M. and Palacios E.,
"New method for COP optimization in water-
and air-cooled single and double effect LiBr-
water absorption machines", International
Journal of Refrigeration, 34, (2011) 1348-1359.
[6] Talbi M.M. and Agnew B.,"Exergy analysis:
an absorption refrigerator using lithium
bromide and water as working fluid", Applied
Thermal Engineering, 20, (2011) 619-630.
[7] Kaita Y., "Thermodynamics properties of
lithium bromide-water solutions at high
temperatures", International Journal of
Refrigeration, 24, (2001) 374-390.
[8] Cengel Y.A. and Boles M.A., "Thermo-
dynamics an engineering approach", Tata
McGraw Hill Education Private Limited New
Delhi, (2008).
[9] Arora C.P., "Refrigeration and Air
Conditioning", McGraw Hill Education Private
Limited New Delhi, (2009).
[10] Kaushik S.C. and Arora A., "Energy and
exergy analysis of single effect and series
flow double effect water-lithium bromide
absorption refrigeration systems", International
Journal of Refrigeration, 32, (2009) 1247-
1258.
[11] Mairaj M., Siddiqui S.A. and Hafiz A.,
"Energetic and exergetic analysis of some
models of vapour absorption chillers using
lithium bromide and water", Journal of Basic and
Applied Engineering Research, 2(4), (2015) 326-
329.
[12] Meraj M., Hafiz A. and Ahmad M.J., "Second
Law Analysis of Series and Parallel Flow
Double Effect Vapor Absorption Chiller", IOSR
Journal of Me-chanical and Civil Engineering,
13(3), (2016) 104-110.](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-26-320.jpg)
![Study and analysis of plate type heat exchanger
137
Study and analysis of plate type heat exchanger
NUMAN ANSARI1, MOHD. ISLAM1 and SAYYED HAIDER2
1
Department of Mechanical Engg. Faculty of Engg.& Technology, Jamia Millia Islamia,New Delhi-110025,India
2
Department of Mechanical Engg., Al-Falah School of Engineering & Technology, Haryana-121004, India
*E-mail: numan.ansari5@gmail.com
Abstract
With the emerging trend in technology heat transfer dissipation from those device which required
certain temperature for their efficient operation in the major problem. In this regards high
technology is required to cool such devices. As far as heat exchanger is the major solution of such
problem. Plate heat exchanger are the exchangers having a number of plates combined together
having different corrugation for transfer the heat between the two fluid. An experimental
investigation are done on the plate type heat exchanger to estimate the value of heat transfer
characteristics by changing the mass flow rate and study the pressure drop. And from the
experiment estimate the value of heat transfer characteristics for hot and cold fluid, nusselt
number, pressure drop and effectiveness. It is shown that Nusselt number, pressure drop and
effectiveness are depends on the mass flow rate. This paper is concerned with the plate type heat
exchanger and investigation is carried out by considering the different works published in the
field of heat exchangers.
Key words : Plate heat exchanger (PHX), heat exchanger (HXCH), chevron angle, gaskets,
plate, flow arrangement, domestic hot water (DHW).
1. Introduction
Plate heat exchanger (PHE) are the exchangers
having a number of plates combined together having
different corrugation for transfer the heat between
the two fluid. Plate heat exchanger are widely used
in the industry and are commonly designed with the
corrugated channel surface resulting in the enhanced
heat transfer performance by increasing the area over
which heat transfer take place and generating a
vigorous mixing of effect within the working fluid.
Component of plate type heat exchanger are shown
in figure.
As shown in Figure 1, the plate heat exchanger is
basically a series of individual plates pressed between
two heavy end covers. These plates are gasketed,
welded or brazed together depending on the
application of the heat exchanger. The basic geometry
of plates used in plate heat exchanger is shown in
figure 2. Stainless steel is a commonly used metal
for the plates because of its ability to withstand
Invertis Journal of Renewable Energy, Vol. 7, No. 3, 2017 ; pp. 137-141
high temperatures, its strength, and its corrosion
resistance[1]. The entire assembly is held together by
the tie bolts. This has a major advantage over a
conventional heat exchanger in that the fluids are
exposed to a much larger surface area because the
fluids spread out over the plates. This facilitates the
transfer of heat, and greatly increases the speed of
the temperature change.
The concept behind a heat exchanger is the use
of pipes or other containment vessels to heat or cool
one fluid by transferring heat between it and another
fluid. In most cases, the exchanger consists of a coiled
pipe containing one fluid that passes through a
chamber containing another fluid. The walls of the
pipe are usually made of metal or another substance
with a high thermal conductivity to facilitate the
interchange, whereas the outer casing of the larger
chamber is made of a plastic or coated with thermal
insulation, to discourage heat from escaping from the
exchanger.
DOI No. : 10.5958/2454-7611.2017.00019.4](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-27-320.jpg)
![138
Numan Ansari, Mohd. Islam and Sayyed Haider
Chevron Angle , Typically varying from 20° to
65°, is the measure of softness (small , low thermal
efficiencyand pressure drop) and hardness (large ,
high thermal efficiency and pressure drop) of thermal
and hydraulic characteristics of plates. Some authors
define "/2- " as the chevron angle[1].
Pressure drop (p) is in direct relationship to the
size of the plate heat exchanger. If it is possible to
increase the allowable pressure drop, and incidentally
accept higher pumping costs, then the heat exchanger
will be smaller and less expensive. As a guide,
allowable pressure drops between 20 and 100 kPa are
accepted as normal for water/water duties.
2. PHE Parts and their function
Plate with gaskets Plate heat exchangers consist
of pressed plates; the separating gaskets between each
plate; the end plates used to clamp the plate stack
together and the frame to hold the plate stack in place.
The plate is pressed metal. A wide range of metals
and corrugation shapes can be used that suit the
chemical, flow and corrosive properties of the media
passing across the plates[4].
Plates for an exchanger have indentations and
corruga-tions to encourage more turbulent flow
across them and to make thinner films of media to
promote better heat transfer. The headers to feed and
remove the fluids pass through all plates. Unwanted
headers are blocked off on a plate.
Gaskets, each plate have a gasket that produces
a sealing and channel system through the entire plate
pack in which the two heats exchanging media flow
in a counter-current direction.
The circular portion of a gasket stops the fluid
from going across the heat transfer plate and sends it
to next open channel. The remaining portion or field
gaskets directs the opposing fluid across the heat
transfer surface. The gaskets separate the plates and
create the thin chamber through which the fluid film
swirls and flows. They also serve the purpose of
directing the media from the entry port to the exit
port.
The frame is made up of thick steel pressure
retaining parts, the fixed cover and the movable cover
that when pulled together with the tightening bolts
from the pres-sure retaining structure for the plates.
The carrying bar and guide bar act as a carrier and
guide both the plates and movable cover. The heat
exchangers plates, which make up the heat transfer
surface, are clamped between two plates of steel with
the use of the tightening bolts. The heat exchanger
construction allows a plate heat exchanger to be easily
opened for inspection and cleaning. In a plate heat
exchanger corrugated plate like this one are used
between the barrier of hot fluid and cold fluid. The
corrugated plate increases the area of heat transfer.
PHE plate usually made of advance material like
titanium which makes plate stronger and durable.
The thickness of plate is 0.5 mm to 0.6 mm.
Fig. 1. Various parts of PHE
Fig. 2. Geometry of plate [1]](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-28-320.jpg)
![Study and analysis of plate type heat exchanger
139
Flow arrangement, Plate heat exchangers are
very easy counter flow where hot and cold fluid flow
in opposite direction as shown in figure. The plate is
placed in such a way that the flow bond by each
successive plate alter-nates the hot fluid and cold
fluid.
The heat exchangers plates with gaskets are
arranged in an alternating pattern of left hand flow
and right hand flow to direct the fluids in an opposing
direction within the heat exchanger. The complete
assembly of all the plates and gaskets is called the
"Plate Pack".
3. Modelling of PHX and analysis of heat transfer
performance
The experimental set up presents the principal
geometry parameters and illustrates the experimental
setup established to investigate the heat transfer
characteristics in the corrugated channel for different
flow conditions. The basic components of the
experimental apparatus include a water loop, an air
loop, and a measurement system. The water loop
comprises a water tank containing a heater, a pump,
a flow meter, and a temperature controller.
Fig. 5. Flow arrangement of PHE
Fig. 3. Plate with gaskets [4]
port ring gasket
Flow ports
Fig. 4. Gaskets [4]
Fig. 6. Experimental set up of PHE
Importantly, all of the components in the water
system are thermally insulated such that wall
temperature of the corrugated channel can be
maintained at a nearly constant temperature. The air
loop consists of the test section containing the
corrugated channel, a blower, an air flow meter, and
a number of valves which enable the flow rate to be
adjusted. Additionally, a flow straightner can be
installed at the entrance of the test section to maintain
a uniform inlet flow. The test section is constructed
from Galvanized Iron (GI) sheets, each with
corrugated surface on both the sides. The two cases
are clipped between two corrugated plates, metal
plates to form the corrugated channel through which
the working fluid is passed. During experiments, hot
water was flowed through the two hollow cases to
maintain the channel surfaces at an approximately
constant temperature and thermocouples wrapped](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-29-320.jpg)
![Study and analysis of plate type heat exchanger
141
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with Web Data: A Survey," IEEE Trans.
Knowledge and Data Eng., preprint, 21 Dec. 2007,
doi:10.1109/TKDE.2007.190746.(PrePrint).](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-31-320.jpg)
![142
Hasan Shamim, Shadab Ahmad and Shah Alam
Study of CFD approach to discretise first derivative
of partial differential equation
HASAN SHAMIM, SHADAB AHMAD and SHAH ALAM
Jamia Millia Islamia, New Delhi
*E-mail: hasanshmm@gmail.com
Abstract
The process of converting governing partial differential equations into algebraic equation is
known as discretisation. The finite difference method is one of the most powerful and simple
method to convert partial differential equation (PDE) into algebraic equation. The accuracy of
finite difference method increases with refining grid size. In this present study the finite difference
method (FDM) is used to determine the shear stress over a flat plate. This shear stress is function
of velocity gradient as given by Newton's Law of viscosity. The first derivative of velocity gradient
is replaced by forward difference, Reward difference and the central difference approaches. The
exact solution of equation is compared with approximate solution and the errors are determined
for different values of grid points spacing. It is found that central approach is more accurate
than forward and reward approaches. The effect of order of accuracy such as O(y), O(y)2,
O(y)3 has been also studied. It shows as order of accuracy increases, computational errors
decreases.
Key words : Discretisation, PDE, Finite difference method, Forward difference, Reward
difference, Central difference, Order of accuracy, Mesh spacing,
1. Introduction
Computational Fluid Dynamics (CFD) provides
a qualitative (sometimes even quantitative)
prediction of fluid flows by means of mathematical
modelling, Numerical methods and software tools[1].
These advantages of CFD over experiment based
approach are given as :
Substantial reduction of lead time and cost
of new designs.
Ability to study systems where controlled
experiments are difficult or impossible to
perform.
Ability to study systems under hazardous
conditions.
The main problem with CFD is that code user
must have high skilled position. The assumptions,
made regarding type of flow e.g. two dimensional,
three dimensional, Compressible, Non compressible,
Invertis Journal of Renewable Energy, Vol. 7, No. 3, 2017 ; pp. 142-146
Viscous, Non viscous. Laminar, Turbulent should be
realistic and must also be able to handle convergence,
consistent, stability associated problems in the
solution[2]. However there are several applications
of CFD in different fields of daily life and engineering.
While analysing problems with CFD approach, we
get PDE[3]. These equations are first converted into
algebraic equations;only then solution of PDE is
possible. This process is known as Discretisation.
There are three different methods to discretise the
PDEs into algebraic equations. These methods are
finite difference method (FDM), Finite element
method (FEM) and Finite volume method (FVM)[4].
In this present analysis we are using FDM to discretise
PDE and observe the accuracy of solution.
2. Statement of Problem
In this we want to determine shear stress over a
flat plate, when air is flowing over it. The velocity
profile is vertical direction is assumed as :
DOI No. : 10.5958/2454-7611.2017.00020.0](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-32-320.jpg)
![Study of CFD approach to discretise first derivative of partial differential equation
143
= 1 –
y
L
u c e
Where c = 480
L is the length of plate (Let L = 3 cm)
We determine shear stress at a distance of
0.35 cm with step size of 0.10 cm, 0.15 cm, 0.20 cm,
0.25 cm, 0.30 cm, 0.35 cm, 0.40 cm.
Dynamic viscosity of air at 15°C, µ = 1.81×10–5
Pa-s
3. Methodology
i. On the basis of above parameters, We want
to find out shear stress near wall by using
Newton's law of viscosity, Given as :
wall =
du
dy wall
ii. The velocity gradient which is first order
differential equation is discretised by using
Finite difference approach.
iii FDM for 1st order accuracy i.e. forward,
reward and central difference approaches
are used to discretise the equation. The
expressions are as given below[1,5]:
1
, i, j–
+ O( y)
i ju uu
dy y
Forward
, i, j – 1–
+ O( y)
i ju uu
dy y
Reward
1
2
, i, j – 1–
+ O( y)
*
i ju uu
dy y
Central
iv FDM for 2nd order accuracy i.e. forward,
reward/backward & central difference
approaches are used to discretise the
equation[6].
The expressions are as given below:-
2 3
2
, i, j + 1 , 2
+ 4* – *
+ O( y )
*
i j i ju u uu
dy y
Forward
23
2
, i, j – 1 , 2
* – 4*
+ O( y )
*
i j i ju u uu
dy y Reward
2 1 24
4
, i, j + 1 , , – 2
– + 4* – *
+ O( y )
*
i j i j i ju u u uu
dy y
Central
v. FDM for 3rd order accuracy i.e. forward,
reward/backward & central difference
approaches are used to discretise the
equation[7, 8].
The expressions are as given below :
3 13 18 11
6
, i, j + 2 , , 3
* + 9* * – *
+ ( y )
*
i j i j i ju u u uu
dy y
Forward
2 311 9 2
6
, i, j –1 , – , – 3
* – 18* * – *
+ ( y )
*
i j i j i ju u u uu
dy y
Reward
3 1 1 2 32 18 18 9 2
12
, i, j +2 , , , , – 3
* – 9* * – * * – *
+ O( y )
*
i j i j i j i j i ju u u u u uu
dy y
Central
vi The approximate solution of equation is
compared with exact solution at (y=.35) for
different step sizes (y) i.e. 0.10, 0.15, 0.20,
0.25, 0.30, 0.35, 0.40.
vii The effect of order of accuracy on solution
has been also observed. The order of accuracy
is taken as O(y), O(y)2, O(y)3 for
FORWARD, BACKWARD/REWARD &
CENTRAL difference method.
viii The errors are also represented in graph for
different step sizes and order of accuracy.](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-33-320.jpg)
![146
Hasan Shamim, Shadab Ahmad and Shah Alam
Table 3. 3rd order accuracy for forward, reward and central difference approach
S.l. No. Step size Error (%) Error (%) Error (%)
forward reward central
1 0.1 0.001 -0.001 -3.70E-05
2 0.15 0.003 -0.003 -1.88E-04
3 0.2 0.007 -0.008 -5.94E-04
4 0.25 0.013 -0.016 -1.45E-03
5 0.3 0.022 -0.028 -3.01E-03
6 0.35 0.035 -0.046 -5.58E-03
7 0.4 0.051 -0.070 -9.54E-03
Fig. 3. Errors in forward (series 1), reward (series 2) and central (series3) approach using O( y)3
ERROR(%)
5. Conclusion
The effect of grid spacing and order of
accuracy has been observed by calculating
shear stress.
It is found the central difference approach has
least error for fix step size compare to
forward and backward difference approach.
Higher order of accuracy has least error in
each approach.
References
[1] Anderson D. John, "Computational Fluid
Dynamics: Basic with applications" Tata
McGraw Hill, (2012).
[2] Versteeg H. K. and Weeratunge Malalasekera,
"An Introduction to Computational Fluid
Dynamics" (2007).
[3] Ferziger J.H. and Peric M., Computational
Methods for Fluid Dynamics. Springer, (1996).
[4] Hirsch C., Numerical Computation of Internal
and External Flows. John Wiley & Sons,
Chichester, I & II, (1990).
[5] Wesseling P., Principles of Computational
Fluid Dynamics. Springer, (2001).
[6] Donea J. and Huerta A., Finite Element
Methods for Flow Problems. John Wiley & Sons,
(2003).
[7] Ohner R.L., Applied CFD Techniques: An
Introduction Based on Finite Element Methods.
John Wiley & Sons, (2001).
[8] Cuvelier C., Segal A. and Van Steenhoven
A.A., Finite Element Methods and Navier-
Stokes Equations. Kluwer, (1986).
[9] CFD-Wiki http://www.cfd-online.com/Wiki/
Main Page.](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-36-320.jpg)
![148
Tribhuwan Chandra Joshi and Yogendra Singh Kushwah
commonly known as Charge Air Cooler is being used
between turbo-charger and the intake manifold.
Wherein high temperature & high pressure air loses
its heat to cooling media (air or water). This basically
results in increased density of combustion air &
ultimately improved combustion/fuel efficiency of
vehicle. Further high boost pressure requirements of
intake air to meet latest emission norms (Euro V/
Euro VI) has emphasized the need of highly efficient
and durable design of heat exchangers. This will
ensure increased heat transfer rate along and less
charge air side pressure drop & overall compactness.
This requirement has enforced the technology
developers to explore liquid based charge air cooling
instead of traditional air to air charge air coolers.
To achieve the aforementioned objective, some
work has been done by researchers in the areas of
liquid cooled charge air coolers, Lamich et al. [1] has
disclosed a cross flow type liquid charge air cooler,
where liquid water is flowing in a U flow manner
through equally spaced tube banks, whereas charge
air flows through inter tube spacing to make a Charge
air flow path.
Maceroni et al.[2] has disclosed a Liquid cooled
charge air cooler directly integrated with Intake
manifold, in this arrangement charge air from inlet
tank flows through heat exchanger, where it loses its
heat with liquid water flowing through heat
exchanger tubes.
The majority of the work reported in existing
literature suggests either straight or U flow type
flow arrangement for liquid water, whereas
charge air flows through tube banks in the direction
normal to tube bank, to form a Cross flow type heat
exchanger. Thus better heat transfer characteristics
possible with counter flow type arrangements can't
be utilized.
Hence there is a scope of flow path optimization,
So as to achieve advantages of both cross as well as
counter flow type arrangements. Present study
explores an innovative charge air flow mechanism,
to form a cross- counter flow type arrangement[3]. So
as to ensure increased heat transfer rate along with
comparable charge air side pressure drop and overall
heat exchanger compactness.
A CFD methodology to evaluate the various
performance parameters of heat exchanger is used.
Fin Surfaces associated with the heat exchanger has
been modeled as a shell region[4], So as to take into
account conduction heat transfer though fins. The
CFD modeling strategy for heat exchangers has been
correlated with test results of baseline liquid to air
heat exchanger (radiator configuration) & it has been
found that CFD results are in close agreement with
test results. In a different study to develop correlation
between liquid water temperature and cooling air
Anders et al.[5] have also used CFD modeling strategy
for estimating the liquid water temperatures at
radiator inlet and found that estimated values were
within ± 4°C of the experimental data.
2. Theoretical Background
Overall heat transfer rate (Q) for heat exchangers
can be expressed by equation (1), (2) & (3).
Q = .Cmin. (Tca_in-Tc_in)/3600 (1)
Charge Air Side :
Qca = mca .Cp_ca . (Tca_in-Tca_out) /3600 (2)
Liquid Water Side :
Qc = mc .Cp_c . (Tc_in-Tc_out) /3600 (3)
In perfect isolated system, heat gained by the cold
fluid (liquid) will be equal to heat lost by the hot fluid
(charge air).
Qc = Qca (4)
Where Heat exchanger effectiveness () is defined
as the ratio of the actual heat transfer rate (Q) to the
maximum possible heat transfer rate (Qmax); and
Cmin is minimum of heat capacities of both the fluids:
= Q/Qmax (5)
Qmax = Cmin. (Tca_in-Tc_in) (6)
Cmin =Min. (mca .Cp_ca, mc .Cp_c) (7)
In current study a detailed investigation of
various flow configurations & their impact on various
design parameters has been discussed.
Where :
WHX : Weight of heat exchanger (kg)
VHX : Heat exchanger volume (m3)
W* H* D* fp : Width* Height* Depth* fin pitch (all in
mm)](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-38-320.jpg)
![Thermal design of liquid cooled charge air cooler: A computational approach
157
(SVP, Subros Technical Centre) for supporting this
research work of new technology development.
References
[1] U.S. Patent Number 7,793,710 B2, Intercooler
Apparatus and Method, Bernhard Lamich &
Denis Bazika dated Sep. 14, 2010
[2] U.S. Patent Number 2013/0220289 A1, Intake
System with An Integrated Charge Air Cooler,
Karen Elizabeth dated Aug., 2013
[3] Kays W.M. and London A.L., Compact Heat
Exchangers, Third Edition.
[4] Star CCM+ User Guide, Version 10.04.009
[5] Anders Jerhamre and Anders Jonson,
Development and Validation of Liquid water
Temperatures and Cooling Air Flow CFD
Simulation at Volvo Cars, SAE Technical paper
2004-01-0051](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-47-320.jpg)
![Utilisation of exhaust gases of automobile
159
the fuel consumption, further for small cars
compressor needs 3 to 4 bhp i.e. a significant ratio of
the power output. Keeping these problems in mind,
a car air conditioning system is proposed which is
using exhaust heat. The advantages of this system
over conventional air-conditioning system are that
it does not affect designed efficiency life and fuel
consumption of engine. G VICATOS, J
GRYZAGORIDIS and S WANG[1] proposed a new
system, Vapour Absorption Refrigeration of System
based on Exhaust gas from an Internal Combustion
Engine. Here energy from the Exhaust gas of an
Internal Combustion Engine is used to power
absorption Refrigeration system to Air condition an
Ordinary passenger. In this theoretical Design is
verified by the unit that is tested under both
laboratory and road-test conditions, by using
Refrigerant 134a and absorbent Dimethyl formamide
(DMF) the cooling effect for a passenger in a vehicle
is been very comfort.
Dimethylformamide is a clear liquid, miscible
with water and majority organic solvents and it is
used as a solvent in the production of fibers, films,
and surface coatings.
Recent trend about the best ways of using the
deployable sources of energy in to useful work in
order to reduce the rate of consumption of fossil fuel
as well as pollution. Out of all the available sources,
the internal combustion engines are the major
consumer of fossil fuel around the globe. Out of the
total heat supplied to the engine in the form of fuel,
approximately, 30 to 40% is converted into useful
mechanical work. The remaining heat is expelled to
the environment through exhaust gases and engine
cooling systems, resulting in to entropy rise and
serious environmental pollution, so it is required to
utilized waste heat into useful work. The recovery
and utilization of waste heat not only conserves fuel,
usually fossil fuel but also reduces the amount of
waste heat and greenhouse gases damped to
environment. It is imperative that serious and
concrete effort should be launched for conserving this
energy through exhaust heat recovery techniques[2].
Such a waste heat recovery would ultimately reduce
the overall energy requirement and also the impact
on global warming. The Internal Combustion Engine
has been a primary power source for automobiles and
automotives over the past century. Presently, high
fuel costs and concerns about foreign oil dependence
have resulted in increasingly complex engine designs
to decrease fuel consumption. For example, engine
manufacturers have implemented techniques such as
enhanced fuel-air mixing, turbo-charging, and
variable valve timing in order to increase thermal
efficiency. However, around 60-70% of the fuel
energy is still lost as waste heat through the coolant
or the exhaust. Moreover, increasingly stringent
emissions regulations are causing engine manu-
facturers to limit combustion temperatures and
pressures lowering potential efficiency gains. As the
most widely used source of primary power for
machinery critical to the transportation, construction
and agricultural sectors, engine has consumed more
than 60% of fossil oil. On the other hand, legislation
of exhaust emission levels has focused on carbon
monoxide (CO), hydrocarbons (HC), nitrogen oxides
(NOx), and particulate matter (PM). Energy
conservation on engine is one of best ways to deal
with these problems since it can improve the energy
utilization efficiency of engine and reduces emissions.
Given the importance of increasing energy conversion
efficiency for reducing both the fuel consumption and
emissions of engine, scientists and engineers have
done lots of successful research aimed to improve
engine thermal efficiency, including supercharge,
lean mixture combustion, etc. However, in all the
energy saving technologies studied. Engine exhaust
heat recovery is considered to be one of the most
effective. Many researchers recognize that Waste
Heat Recovery from engine exhaust has the potential
to decrease fuel consumption without increasing
emissions, and recent technological advancements
have made these systems viable and cost effective.
This paper gives a comprehensive review of the waste
heat from internal combustion engine, waste heat
recovery system and methods of waste heat recovery
system. Better fuel economy and higher power with
lower maintenance cost has increased the popularity
of diesel engine vehicles. Diesel engines are used for
bulk movement of goods, powering Stationary/
mobile equipment, and to generate electricity more
economically than any other device in this size range.
In most of the global car markets, record diesel car
sales have been observed in recent years. The
exhorting anticipation of additional improvements in](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-49-320.jpg)
![160
Jitendra Singh and Shahbaz Mozammil
diesel fuel and diesel vehicle sales in future have
forced diesel engine manufacturers to upgrade the
technology in terms of power, fuel economy
andemissions[3].
2. Vapour AbsorptionSystem
An absorption refrigerator is a refrigerator that
uses a heat source (e.g., solar, kerosene-fuelled flame,
waste heat from factories or district heating systems)
to provide the energy needed to drive the cooling
system[4]. In the early years of the twentieth century,
the vapour absorption cycle using water-ammonia
systems was popular and widely used. Ammonia-
water combination possesses most of the desirable
qualities which are listed below:
1 m3 of water absorbs 800 m3 of ammonia
(NH3).
Latent heat of ammonia at-15°C = 1314 kJ/
kg.
Critical temperature of NH3 =132.6°C.
Boiling point at atmospheric pressure =
–33.3°C
The NH3-H2O system requires generator
temperatures in the range of 125°C to 170°C with air
cooled absorber and condenser and 80°C to 120°C
when water-cooling is used. These temperatures
cannot be obtained with flat-plate collectors. The
coefficient of performance (COP), which is defined
as the ratio of the cooling effect to the heat input,
which vary between 0.6 to 0.7. Ammonia is highly
soluble in water.
3. Refrigerant used for the absorption refrigera-
tion systems
The properties of the refrigerant combination are
that, in liquid phase, they must have a margin of
miscibility within the operating temperature range
of the cycle. The mixture should also be chemically
stable, non-toxic, and non-explosive. In addition to
these requirements, the following aredesirable[5].
a. Refrigerant should have high heat of
vaporization and high concentration within
the absorbent in order to maintain low
circulation rate between the generator and
the absorber per unit of the coolingcapacity.
b. Transport properties that influence heat and
mass transfer, e.g., viscosity, thermal
conductivity, and diffusion coefficient should
befavorable.
c. Both refrigerant and absorbent should be
non- corrosive, environmental friendly, and
economical. There are some 40 refrigerant
compounds and 200 absorbent compounds
available. However, the most common
working fluids are water/ammonia and
LiBr/water. Since the invention of absorption
refrigeration systems, water/ammonia has
been widely used for both cooling and
heatingpurposes.
The main properties are :
Ammonia is the only refrigerant from inorganic
group which was used universally for many
applications and still used to great extent at the
present time. It possesses many properties required
for an ideal refrigerant. It has wide applications
because of its low volumetric displacement, low cost,
low weight of liquid refrigerant per ton of
refrigeration and high efficiency. Presently it is
widely used in cold storages, ice manufacturing
plants and skating rinks due to its low production
and maintenance cost[6].
Properties of Ammonia
1. Its toxic, flammable, irritating and
fooddestroying
2. Anhydrous ammonia has no effect on
lubricating oil but in presents of moisture,
ammonia forms an emulsion with oil thatFig. 1. Vapour Absorption System](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-50-320.jpg)
![Utilisation of exhaust gases of automobile
161
causes operating difficulties. The formed
emulsion becomes effective when the
percentage of water exceeds 0.01%. It is not
oil miscible therefore it will not dilute the oil
in the crank case of thecompressor.
3. It is highly volatile and becomes explosive
when mixed with air and compressed
therefore air leaks must be avoided in
ammonia refrigerationsystems.
4. Ammonia attacks on non-ferrous metals in
the presents of water therefore copper and
brass are never used with ammonia
refrigerationsystem.
5. Ammonia can be used economically for -70°C
evaporator temperature and its application
for further low temperature becomes highly
uneconomical and difficult to maintain low
vacuum required in the evaporator.
6. Ammonia (refrigerant) and water (absorbent)
are highly stable for a wide range of
operating temperature andpressure.
7. Ammonia has a high latent heat of
vaporization, which is necessary for efficient
performance of the system. Its latent heat of
vaporization at -15 ºC is 1315 kJ/Kg.
8. Its boiling point at atmospheric pressure is -
33.3 ºC & freezing point is -77ºC
9. It has highest refrigerating effect per Kg of
refrigerant. The leakage of this refrigerant
may be quickly & easily detected by the use
of burning sulphur candle which in the
presence of ammonia will form white fumes
of ammonium sulphite. It is environmental
friendly[7].
4. Exhaust gas recirculation system
When EGR system is applied, the engine intake
consists of fresh air and recycled exhaust gas. Exhaust
gases were tapped from exhaust pipe and connected
to inlet airflow passage. An EGR control valve was
provided in this pipe for EGR control. The exhaust
gases were regulated by this valve and directly send
to the inlet manifold, upstream of compressor or
downstream of compressor. Sufficient distance for
through mixing of fresh air and exhaust gases were
ensured. The above shown system is also called as
hot EGR because it not fitted with an EGR cooler
which is used to cool the intake mixture. A typical
EGR valve is shown in Fig 2.
4.1 Types of EGR systems
Low pressure EGR Systems : If portion of turbine
outlet exhaust gas is delivered to compressor inlet
through the flow control valve then it called low
Fig. 2. EGR System in a Turbocharged Engine
Fig. 3. Low Pressure EGR System](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-51-320.jpg)
![162
Jitendra Singh and Shahbaz Mozammil
of exhaust gas. This type of loop can be used by
directing exhaust from the turbine outlet to the inter-
cooler outlet directly bypassing thecompressor.
High pressure EGR systems : Another method of
EGR is high pressure EGR loop. In high pressure EGR
system, a flow passage is devised between the exhaust
of engine (up-stream of theturbine) and the intake
manifolds of engine (downstream of the super
charging compressor).
In this system the exhaust gas is recirculated from
upstream of the turbine to downstream of the
compressor or the downstream of the inter-cooler as
shown in Figure. The compressor and intercooler are
therefore not exposed to the exhaust gas. Such high
pressure loop EGR system is only applicable when
the turbine upstream pressure is sufficiently higher
than the boost pressure (Compressor downstream
pressure) i.e. if (P3-P2)>0.
5. Possibility of heat recovery and availability
from I.C. engine
Waste heat is heat, which is generated in a process
by way of fuel combustion or chemical reaction, and
then "dumped" into the environment even though it
could still be reused for some useful and economic
purpose. This heat depends in part on the tempera-
ture of the waste heat gases and mass flow rate of
exhaust gas. Waste heat losses arise both from
equipment inefficiencies and from thermodynamic
limitations on equipment and processes. For example,
consider internal combustion engine approximately
30 to 40% is converted into useful mechanical work.
The remaining heat is expelled to the environment
through exhaust gases and engine cooling systems.
It means approximately 60 to 70% energy losses as a
waste heat through exhaust (30% as engine cooling
System and 30 to 40% as environment through
exhaustgas.
Exhaust gases immediately leaving the engine
can have temperatures as high as 842-1112°F [450-
600°C]. Consequently, these gases have high heat
content, carrying away as exhaust emission. Efforts
can be made to design more energy efficient
reverberatory engine with better heat transfer and
lower exhaust temperatures; however, the laws of
thermodynamics place a lower limit on the
pressure EGR loop. In low pressure EGR system, a
flow passage is a devised between the exhaust of
super charger turbine and the intake manifolds
connected to the super charging compressor. The flow
of EGR regulated is with a throttling valve showing
in Figure above. The pressure differences generally
are sufficient to drive the EGR flow of a desired
amount except during idling. If the exhaust gas is
recycled to the intake directly, the operation is called
hot EGR. If the exhaust gas is recycled through EGR
cooler, the operation is called cooledEGR[8].
For turbocharged I.C. engine modification is done
in EGR system because a positive differential
pressure between the turbine outlet and compressor
inlet is generally available (p4-p1)>0. Furthermore,
tailpipe pressure p1 can be elevated by partial
throttling that ensures sufficient driving pressure for
the EGR flow. The low pressure EGR loop is not
applicable as the conventional compressor and inter-
coolers are not designed to ensure the temperature
Fig. 4. High Pressure EGR System](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-52-320.jpg)
![Utilisation of exhaust gases of automobile
163
temperature of exhaust gases. Fig. 5 show total energy
distributions from internal combustionengine.
6. Result and Discussion
6.1 Benefits of 'waste heat recovery' canbe broadly
classified in twocategories:
Direct Benefits :Recovery of waste heat has a direct
effect on the Combustion process efficiency. Thisis
reflected by reduction in the utility consumption and
process cost.
Indirect Benefits: Reduction in pollution: A
number of toxic combustible wastes such as carbon
monoxide (CO), hydrocarbons (HC), nitrogen oxides
(NOx), and particulate matter (PM) etc, releasing to
atmosphere. Recovering of heat reduces the
environmental pollutionlevels.
Reduction in equipment sizes: Waste heat recovery
reduces the fuel consumption, which leads to
reduction in the flue gas produced. This results in
reduction in equipment sizes.
Reduction in auxiliary energy consumption:
Reduction in equipment sizes gives additional
benefits in the form of reduction in auxiliary
energyconsumption.(9)
In automobile engines significant amount of
heat is released to the environment. For example, As
much as 35% of the thermal energy generated from
combustion in an automotive engine is lost to the
environment through exhaust gas and other losses.
The amount of such loss, recoverable at least partly
or greatly depends on the engine load. Among
various advanced concepts, Exhaust Energy Recovery
for Internal Combustion (IC) engines has been proved
to not just bring measurable advantages for
improving fuel consumption but also increase engine
power output (power density) or downsizing, further
reducing CO2 and other harmful exhaust emissions
correspondingly. Which was predicted that if 6% of
the heat contained in the exhaust gases were
converted to electric power, 10% reduction of fuel
consumption can beachieved.
7. Conclusions
The possibility to design a refrigeration unit inside
an automobile using the waste heat from the
engine of the vehicle based on Vapor Absorption
Refrigeration System is realistic. Also keeping in mind
the Environmental safety view, this system is Eco-
friendly as it involves the use of Ammonia (a natural
gas) as a refrigerant and is not responsible for Green
House effect and OZONE layer depletion. In this way
we can conclude, that out of the total heat supplied
to the engine in the form of fuel combustion,
approximately, 35% to 40% is converted into useful
mechanical work; the remaining heat is categorized
under the waste heat and expelled out of the system,
resulting in the rise of entropy, so it is required to
utilize this waste heat into useful work. Possible
methods to recover the waste heat from internal
combustion engine through the study on the
performance and emissions of the internal
combustion engine are discussed upon and can be
designed. Waste heat recovery system is the best way
to recover waste heat and saving thefuel[10].
References
[1] Endo T., Kawajiri S., Kojima Y., Takahashi
K., Baba T., Ibaraki S. and Takahashi T.,
"Study on Maximizing Exergy in Automotive
Engines," SAE Int. Publication, 2007-01-0257,
2007.
[2] Nantha Gopal K., Rayapati Subbarao,
Pandiyarajan V. and Velraj R., "Thermo-
dynamic analysis of a diesel engine integrated
with a PCM based energy storage system,"
International Journal of Thermodynamics 13(1),
(2010) 15-21.
Fig. 5. Fuel Energy Content in I. C. Engine](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-53-320.jpg)
![164
Jitendra Singh and Shahbaz Mozammil
[3] Hakan Özcan and Söylemez M.S., "Thermal
balance of a LPG fuelled, four stroke SI engine
with water addition," Energy Conversion and
Management, 47(5), (2006) 570-581.
[4] Sathiamurthi P., "Design and Development of
Waste Heat Recovery System for air
Conditioning," Unit European Journal of Scientific
Research, 54(1), (2011) 102-110.
[5] YadavR.J. and Verma R.S., "Effective
Utilization of Waste Heat from Diesel Genset
to Run Air Conditioning Plant", Advances in
Energy Research, (2006).
[6] Shah Alam, A proposed model for utilizing
exhaust heat to run automobile air conditioner,
joint international conference on "sustainable
energy and environment (SEE 2006)" 21-23
Nov(2006), E-011(P).
[7] Florides G.A., Kalogirou S.A., Tassou S.A.and
Wrobel L.C., Design and construction of a LiBr-
water absorption machine, Energy Conversion
and Management, 44, (2003) 2483-2508.
[8] Verband der Automobilindustrie, VDA
Jahresbericht (2009).
[9] Dietsche K.-H. and Klingebiel M., Bosch
Automotive Handbook, 7th edition, (2007).
[10] Balaji K. and Senthil Kumar R., Study of vapor
absorption system using waste heat in sugar
industry, IOSR Journal of Engineering, (2012),
2250-3021.](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-54-320.jpg)
![166
Md. Khurshid Alam and Shah Alam
Apart from this there are several disadvantages
of this system given as :
There is limitation on the distance between
the indoor and outdoor unit i.e. refrigerant
piping can't exceed the limits stipulated by
the manufacturer (usually 100 to 150 ft)
otherwise the performance will suffer;
Maintenance (cleaning/change of filters) is
within the occupied space;
Limited air throw, which can lead to possible
hot/cold spots;
Impact on building aesthetics of large
building because too many outdoor units will
spoil the appearance of the building.
VRF System : The term VRF refers to the ability
of the system to control the amount of refrigerant
flowing to each of the evaporators, enabling the use
of many evaporators of differing capacities and
configurations, individualized comfort control,
simultaneous heating and cooling in different zones,
and heat recovery from one zone to another. VRF
systems operate on the direct expansion (DX)
principle meaning that heat is transferred to or from
the space directly by circulating refrigerant to
evaporators located near or within the conditioned
space. Refrigerant flow control is the key to many
advantages as well as the major technical challenge
of VRF systems.
Note the term VRF systems should not be
confused with the centralized VAV (variable air
volume) systems, which work by varying the air flow
to the conditioned space on variation in room loads.
2. Literature Review
Roth et al. (2002)[1] had presented the paper on
VRF installed costs are highly dependent on the
application, construction and lay out of the building
and whether the installation is new or retrofit. Lack
of familiarity with the technology in the U.S will add
to VRF costs. Total costs of VRF systems are likely to
be about 5% to 20% higher than chilled water systems
of similar capacity. Hai, Xiaohong (2006 )[2] et al. had
discussed an effective method to combine the ice-
storage technology with VRV system in commercial
central air conditioning system and the system has
the merits of flexible controls, simple operation and
compact construction, which need not the big space
and difficult construction like the big central ice-
storage system.
William Goetzler (2007)[3] et al. had presented the
paper on VRF systems are not suitable for all
commercial building applications. However, they are
an excellent option for certain projects, and one more
tool for engineers to consider. Johnson (2007)[4] et al.
had presented the paper on Ozone depletion issues
became an increasing concern at that time issues of a
high refrigerant charge of multi split system was
likely a strong negative for the system. Since that time,
refrigerant developments, advances in charge
management, control and inverter technology
Increasing market acceptance of VRF technology.
Morton Blatt, (2008)[5] et al. had presented It observed
to VRF systems are enhanced versions of ductless
multi-split systems. VRF systems are very popular
in Asia and support from major U.S. Main impact of
this system on the electric utility, application
recommendations and technology attributes. John
Rogers, (2008)[6] et al. had presented this paper
reviews The India Low Carbon Growth study is
developing a bottom-up model that covers the
electricity supply, residential, nonresidential
buildings, transport, industry, and agricultural
sectors. In heating/cooling includes like electric water
heater, air cooler, fans, air conditioning. It can be seen
that by 2031 the difference amounts to 135,240 GWh/
year, equivalent to the total output of approximately
thirty eight 500 MW power stations. Qiu Tu (2010)[7]
et al. had presented the paper on the heating control
strategy is key technology for stable and reliable
heating operation of variable refrigerant flow air
conditioning system with multi-module outdoor
units. Li and Wu (2010)[8] et al. had presented the
paper on the heat recovery VRF system can save up
to 17% of the energy consumption compared to a heat
pump VRF system set up. This heat recovery VRF
system offered an additional advantage that it can
provide cooling and heating for different zones at the
same time. Tolga N. Aynur (2010)[9] et al. had
presented the paper on VRF system to control
refrigerant mass flow rate according to cooling and
heating load. This detailed review indicates that the
researchers focus on three main subjects: (a) Control
strategies of the A Review of a HVAC With VRF
System (IJIRST/ Volume 1 / Issue 10 / 003) All rights](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-56-320.jpg)
![Life cycle analysis of multi-split Variable Refrigerant Flow (VRF) system : case study
167
reserved by www.ijirst.org.Jignesh M.B. [10] has
designed HVAC System foa a commercial building.
3. Methodology
In this analysis we have taken first floor of a
building located in A. M. U. Aligarh (U.P.). The area
of building is 18327 square feet. First we got the plan
of building. With the help of plan we have calculated
all dimensions like length width of each wall,
partition wall, window, floor area. Then according
to the climate of AMU U.P., we select recommended
outdoor condition with the help of Weather Data
Handbook published by ISHRAE carrier hand book.
We get all the specification such as height of walls,
glasses and ceiling height from architecture drawing
and calculated the area of all walls, partition walls
and windows, of conference room which we can get
fresh air, Dehumidified air and TR. On the basis of
cooling load we have selected air conditioning system.
Basis of Design
Ambient Summer
Conditions : DBT 110°F (43.3°C)
WBT 75°F (23.9°C)
RH 60%
Ambient Monsoon
Conditions : DBT 95°F (35.0°C)
WBT 83°F (28.3°C)
RH 88%
Indoor Temperature : 73 ± 2°F in Office and 74 ± 2°F
in Lift Lobby & Entrance Hall
etc.
RH : RH may exceed 65% in peak
monsoon. Monsoon Reheat is
not considered hence humidity
may increase up to 65-70%
during peak monsoon.
Occupancy : 80 Sq.ft per person in office at
chargeable area, chargeable
area is 22% more than air
conditioned area, 125 Sq.ft per
person in entrance lobby,
Equipment Load : 3.5 W/sq.ft in Office area
Lighting Load : 1.25 W/ft² in Office, 1.25 W/ft²
in Lobby
Fresh Air : As per ASHRAE standard 62.1-
2007 (5 CFM per person+0.06
cfm/sqft) + 30 % higher to meet
LEED requirement for Office
area, lobbies
Design Requirements
It is often a practice to design an Air Conditioning
system for peak load conditions.
However, the average usage of an Air
Conditioning system during a year is between 65%
to 75%
Hence, the proposed system to be adopted for
this project will be planned & selected for
Lowest capital cost.
Flexibility to shut down in the areas that are not
in use
Energy efficiency
Minimum Outdoor unit
The Heat Load of Conference Room is given
below :
Table 1. Heat load calculation](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-57-320.jpg)
![170
Md. Khurshid Alam and Shah Alam
Life Cycle Analysis using Annual Cost Analysis method :
C= M+E+I / (P/S, i %, n) – S/ (F/S, i %, n).
F/S = (1+i)n–1, P/S= (1+i)n–1/ i(1+i)n,
Option Installation Maintanance Electric Intrest Year Annual
(Lac) (Lac) consumption (Lac) rate (Lac) expenses (Lac)
SPLIT 52.8 2.45 33.8 0.085 10 42.61193
VRF 79.2 3.5 26.4 0.085 12 0.638348
4. Conclusions
Based on comparative analysis of alternative with
operating cost and life cycle cost, a VRF (variable
refrigerant flow) system is selected for a particular
project of 149.5 TR.
Codes and Standards
NBC - National Building Code
ASHRAE - American Society of Heating,
Refrigeration and Air-Conditioning
Engineers.
ARI - Air-conditioning and Refrigeration
Institute
NFPA - National Fire Protection Association
UL - Underwriters' Laboratories
AMCA - Air Movement and Control
Association
References
[1] Roth Kurt, et al. "Energy consumption
characteristics of commercial building HVAC
system volume III: energy saving potential"
TIAX LLC FOR DOE, 2002 www.eere.gov/
b u i l d i n g / i n f o / d o c u m e n t s / p d f s /
hvacvolume2finalreport.pdf.
[2] Hai Xiaohong, Tao Zhang, Yun Fanhua and
Jun Shen, "Design and Research of the
Commercial Digital VRV Multi-Connected
Units With Sub-Cooled Ice Storage System"
(2006). International Refrigeration and Air
Conditioning Conference. Paper 759. http://
docs.lib.purdue.edu/iracc/759.
[3] William Goetzler, Member ASHRAE,
ASHRAE Journal, April 2007.
[4] Johnson Spellman" ASHRAE headquarters
building renovation, mechanicals narrative"
June26, 2007. http;//images.ashrae.biz/
renovation/documents/o6js22_mech_design_
narrative_permit%20set_2.pdf.
[5] Ammi Amarnath, "Variable Refrigerant Flow:
An Emerging Air Conditioner and Heat Pump
Technology" Electric Power Research Institute.
Morton Blatt, Energy Utilization Consultant
2008, ACEEE Summer Study on Energy
Efficiency in Buildings.
[6] John rogers, "Residential Consumption of
Electricity in India” Background Paper India:
Strategies for Low Carbon Growth, July 2008
The World Bank.
[7] Qiu Tua b, Ziping Fenga b, Shoubo Maoc,
Kaijun Donga b, Rui Xiaoa b and Wenji
Songa, "Heating control strategy for variable
refrigerant flow air conditioning system with
multi-module outdoor units" Guangzhou
Institute of Energy Conversion, Chinese
Academy of Science, Guangzhou 510640, China
b Key Laboratory of Renewable Energy and
Gas Hydrate, Chinese Academy of Science,
Guangzhou 510640, China c Haier Air-
Conditioning Electronic Co. LTD., Qingdao
266510, China
[8] Li Y.M., Wu J.Y. and Shiochi S., "Experimental
validation of the simulation module of the
water-cooled variable refrigerant flow system
under cooling operation", Applied Energy, 87,
(2010) 1513-1521.
[9] Tolga N. Aynur, "Variable refrigerant flow
systems: A review", Center for Environmental
Energy Engineering, Department of
Mechanical Engineering, University of
Maryland, 3157 Glenn Martin Hall Building,
College Park, MD 20742, USA.](https://image.slidesharecdn.com/ijre-732017-170927101635/85/Ijre-published-paper-7-3-2017-60-320.jpg)
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Ijre published paper 7(3), 2017
- 1. InvertisJournalofRenewableEnergyVol.7,No.3(115-172)July-September2017 ISSN 2231-3419 ISSN 2454-7611(Online) (Printed) Volume 7, No. 3, July-September 2017 Available at : www.indianjournals.com Abstracted/Indexed by Index Copernicus International (ICV-77.28), MIAR (ICDS-3.7), Google Scholar, CNKI Scholar, ISA(CSIR), INSPEC, EBSCO Discovery, Summon(ProQuest), Primo & Primo Central, J-Gate, Indian Science, OAJI, Cite Factor, DRJI, ISRA-JIF, ICMJE, DOAJ{Under Process with - Indian Citation Index, Thomson Reuters, SCOPUS, EMBASE, CABI, SJIF, Jour-Informatics, GIF, EZB, ScholarSteer,IIJIF}
- 2. INTERNATIONALADVISORY BOARD Shyam SinghChauhan Ex. Director R.R. Institute of Modern Technology Bakshi Ka Talab, Sitapur Road Lucknow, U.P., India R.K. Sharma Infra red Division SSPL Lucknow Road, Timarpur Delhi - 110 054, India P.J. George Kurukshetra Institute of Technology and Management (KITM) Pehwa Road, Bhorasida, Kurukshetra - 136 119, Haryana, India Nawal Kishore Department of Applied Physics Guru Jambheshwar University Hisar - 125 001, Haryana, India I.P. Jain Director Centre for Non-conventional Energy Sources 14, Vigyan Bhavan University of Rajasthan, Jaipur - 302 004 Rajasthan, India Naresh Padha Department of Physics & Electronics University of Jammu, Jammu - 180 006 India Bhaskar Bhattacharya Director School of Engineering and Technology Sharda University, Knowledge Park 3 Greater Noida - 201 306, U.P., India Kishan Pal Singh Principal University Polytechnic Mangalayatan University Aligarh Mathura Highway Aligarh - 202145, India Avinashi Kapoor Department of Electronic Science University of Delhi, South Campus New Delhi - 110 021 India Amitava Majumdar Sr. General Manager-Technical & Corp. R&D Moser Baer India (MBI) 66, Udyog Vihar, G B Nagar Greater Noida - 201 306, U.P, India C A N Fernando Department of Electronics Wayamba University of Sri Lanka, Kuliyapitiya Sri Lanka R.C. Maheshwari Advisor, Hindustan College of Science & Technology Agra - 211 011, U.P. India G.D. Sharma Physics Department, JNV University Jodhpur - 342 005 Rajasthan, India Z.H. Zaidi Chief Editor Invertis Journal of Science & Technology New Delhi - 110 060 India Abdol Rahim Baharvand Researcher and University Teacher Lorestan University, Khorramabad Iran
- 3. INVERTIS JOURNAL OFRENEWABLE ENERGY Volume 7 July-September 2017 No. 3 Owned, Published and Printed by Sanjeev Gautam, 60/10, Old Rajinder Nagar, New Delhi - 110 060 Printed at Alpha Printers, WZ-35/C, Naraina Ring Road, New Delhi - 110 028. Ph : 9810804196 Chief Editor : Prof. Z.H. Zaidi, Ph. : 09213888999, E-mail : ijre2011@gmail.com CONTENTS The energy transfer/flow management algorithm 115 for a photovoltaic solar home Osama Shaukat Numerical analysis of friction factors in smooth and rough 123 microchannels Mohd Owais Qidwai, M.M. Hasan and Mohd Ariz Effect of operating conditions on the performance of vapor 131 absorption refrigeration system Md. Meraj, Rashid Imam and Md. Ashfaque Alam Study and analysis of plate type heat exchanger 137 Numan Ansari, Mohd. Islam and Sayyed Haider Study of CFD approach to discretise first derivative 142 of partial differential equation Hasan Shamim, Shadab Ahmad and Shah Alam Thermal design of liquid cooled charge air cooler: 147 A computational approach Tribhuwan Chandra Joshi and Yogendra Singh Kushwah Utilisation of exhaust gases of automobile 158 Jitendra Singh and Shahbaz Mozammil Life cycle analysis of multi-split Variable Refrigerant 165 Flow (VRF) system : case study Md. Khurshid Alam and Shah Alam Patron Umesh Gautam R.M. Mehra Sharda University Mohd Parvez Al-Falah University Assistant Editor Sumit Kumar Gautam Editors Chief Editor Z.H. Zaidi
- 4. Ranjana Jha N SI T, Delhi University
- 5. The energy transfer/flowmanagement algorithm for a photovoltaic solar home 115 The energy transfer/flow management algorithm for a photovoltaic solar home OSAMA SHAUKAT Departmentof Mechanical Engineering Al-Falah University, Faridabad, Haryana *E-mail: osama_shaukat@yahoo.com Abstract In this research paper, an energy flow management algorithm for a grid-connected-photovoltaic system with battery storage devoted to supply a home is presented. Modeling of photovoltaic (PV) and wind farms (WF) stations to take into account these renewable energies into the power flow formulation are summarized. This algorithm favors the fulfillment of the home energy demand by the energy produced by the photovoltaic generator or stored in the batteries than that got from the grid. This algorithm manages the flow of energy in the house through the combination of four switches. It was applied to the case of a home installed in a coastal region of Bou-Ismaïl (Algeria).The efficiency of the algorithm was tested for two weeks, a favorable week in summer and an unfavorable week in winter regarding the weather conditions. The simulation of the on grid PV system was done using real data of the irradiation and the temperature acquired by the meteorological station of the Bou-Ismaïl site and a home load profile for each season. The results obtained revealed that the energy demand satisfaction for the house is high in the favorable summer week and is only 33% in the winter unfavorable week. To meet a high rate of self- consumption a Home Energy Management is required. Key words : Photovoltaic (PV); solar home; energy management; consumption profiles. 1. Introduction Solar energy is radiant light and heat from the Sun that is harnessed using a range of ever-evolving technologies such as solar heating, photovoltaic, solar thermal energy, solar architecture, molten salt power plants and artificial photosynthesis. It is an important source of renewable energy and its technologies are broadly characterized as either passive solar or active solar depending on how they capture and distribute solar energy or convert it into solar power. In 2011, the International Energy Agency said that "the development of affordable, inexhaustible and clean solar energy technologies will have huge longer-term benefits. It will increase countries' energy security through reliance on an indigenous, inexhaustible and mostly importindependent resource, enhance sustainability, reduce pollution, Invertis Journal of Renewable Energy, Vol. 7, No. 3, 2017 ; pp. 115-122 lower the costs of mitigating global warming, and keep fossil fuel prices lower than otherwise. In 2000, the United Nations Development Program me, UN Department of Economic and Social Affairs, and World Energy Council published an estimate of the potential solar energy that could be used by humans each year that took into account factors such as insolation, cloud cover, and the land that is usable by humans. These renewable energy sources are a very good solution in the global energy problem. The energy generated by the photovoltaic systems constitutes a large part of the total amount of energy produced by renewable energy sources. Photo voltaic cells belong to the group of distributed generations. These cells produce power by utilizing the sunlight. There have been many research efforts to improve the efficiency of photovoltaic systems. In this context, photovoltaic (PV) generation appears as the most promising DOI No. : 10.5958/2454-7611.2017.00016.9
- 6. 116 Osama Shaukat alternative thanksto its maturity, its environmentally friendly characteristics, low maintenance and the fact that the sun is everywhere. The residential and building sector is in many countries the first energy consumer, about 40% of the global energy consumption[1], therefore efficient utilization and control of energy consumption at household level is crucial[2]. The challenge resides in the matching of the intermittent energy production with the dynamic power demand[3]. This needs to design an energy management strategy in order to optimize the use of PV source and storage and to match the local production with local consumption. In the literature, several works are particularly focused on optimizing the energy. Consequently, studies have been directed to "smart homes"[4-7]. In this paper the case of a home powered by on grid photovoltaic system with storage is investigated. An energy flow management algorithm was developed. It manages the energy flow available in the home through a combination of switches to meet the energy demand optimally using the energy produced by the photovoltaic generator and/or the energy stored in the batteries and/or the energy drawn from the grid. The purpose is to minimize this last and ultimately reach the self- consumption mode. Two particular weeks have been chosen, one in summer and one in winter, in order to tests the algorithm. The paper is organized as follows. In Section 2, the chosen photovoltaic system which feeds the home to satisfy the energy demand profile is presented. The models used to simulate the PV system are given in section 3. The proposed energy flow management algorithm and the operating mode of the system are presented in section 4, while in Section 5, the obtained simulation results, using the MATLAB-SIMULINK are given and interpreted. 2. PV System description and energy demand of the home 2.1 Description of the proposed PV system In order to study how to meet the energy needs of a family by using mainly the energy produced by a photovoltaic installation, the sizing of the PV system has been done by using PVSST 1.0 locally developed software[8]. The optimal PV system configuration obtained is an array of 3.2kWp and a battery bank of 12kWh. To feed the loads of the house, 4kW DC/AC converter has been chosen. In order to prevent the batteries from a total discharge, the lowest level of the State Of Charge (SOC) was set to 25%. This SOC corresponds to a capacity of 3kWh. The SOCmax was set to 85% which corresponds to 10 kWh. As shown on figure 1, the PV system operation depends on the combination status of four switches (K1, K2, K3, K4) which is related to the energy demand, the energy Fig. 1. Grid-connected PV system with storage for the solar home
- 7. The energy transfer/flowmanagement algorithm for a photovoltaic solar home 117 available from the PV generator, the SOC of the battery bank and the energy taken off from the grid. In order to meet optimally the energy needs of the home, an energy flow management algorithm has been developed. 2.2 Solar home energy demand The case treated in this paper is a house of 57 m2 area situated in the coastal region of Bou-Ismaîl (Latitude: 36° 38' 33'' North and Longitude: 36° 38' 33'' North) occupied by a family of three members. The energy demand of a home depends on many factors such as the number of appliances, the electrical power used by each appliance and the amount of use of appliances determined by the behavior of the occupants in the home. Solar cell : A solar cell (also called photovoltaic cell or photoelectric cell) is a solid state electrical device that converts the energy of light directly into electricity by the photovoltaic effect. The following are the different types of solar cells. Amorphous Silicon solar cell (a-Si) Bio hybrid solar cell Buried contact solar cell Cadmium telluride solar cell (CdTe) Concentrated PV cell (CVP and HCVP) 2.3 Home energy demand To design the most efficient PV power system for supplying the dwelling, the solar energy resources of a site and the energy consumption of the household should be known. Unfortunately, the electricity demand is rarely available, so the daily load profile must be generated. A home energy management (HEM) system is an integral part of a smart grid that can potentially enable demand response applications for residential customers. This paper presents an intelligent HEM algorithm for managing high power consumption household appliances with simulation for demand response (DR) analysis. The proposed algorithm manages household loads according to their preset priority and guarantees the total household power consumption below certain levels. A simulation tool is developed to showcase the applicability of the proposed algorithm in performing DR at an appliance level. This paper demonstrates that the tool can be used to analyze DR potentials for residential customers. Given the lack of under- standing about DR potentials in this market, this work serves as an essential stepping-stone toward providing an insight into how much DR can be performed for residential customers. In the studied case, the home load profile has been determined by considering the daily load profile of each appliance, their power ratings and their estimated operating time per day, then their summation leads to the daily consumption home profile. For the summer, the daily energy demand varies between 13.87 kWh and 16.27 kWh. For the winter, the energy demand varies between 12.03 kWh and 18.30 kWh. For these two seasons, the high energy demand is explained by the use of energy intensive equipment for heating and air conditioning. 3. PV system modeling In this section, the modeling of the PV generator, the batteries and the power converter will be presented. For PV system modeling, a one diode model for the PV generator has been used[10]. For the batteries, the Copetti's model has been chosen as it is well suited for lead acid batteries allowing the simulation of the whole operating process charge- discharge-overcharge while considering the temperature change of the battery[11, 12]. For the DC/ DC converter, a boost converter controlled by a MPPT Fuzzy logic Controller has been used[13, 14] and the DC/AC converter has been modeled by its yield curve. Solar energy : Solar is scaling up rapidly, with capacity more than trebling over the past four years. New installations totaling more than 50 GW in 2015 took global solar PV power generating capacity to 231 GW by year end, a 28.1% increase versus the end of 2014. Capacity has more than trebled in the past four years. The largest increments in 2015 were recorded in the Asia Pacific region; China added 15.2 GW, just ahead of Japan with 12 GW, together accounting for more than half of the growth in global solar capacity. The US provided the third largest addition (7.3 GW). China's cumulative installed capacity (43.5 GW) moved it ahead of Germany (39.7 GW), with Japan (35.4 GW) in third place, and the US (25.6 GW) overtaking Italy (18.9 GW) to take fourth place. Solar power generation enjoyed another year of very rapid growth in 2015, with a 33%
- 8. 118 Osama Shaukat increase. Itsoverall share of global power generation remains low (1.1%), but that share has almost doubled in just two years. Solar is starting to have a noticeable impact in terms of sources of power generation growth, contributing more than 30% of the growth of global power in 2015. Energy plays a pivotal role in our daily activities. The degree of development and civilization of a country is measured by the amount of utilization of energy by human beings. Energy demand is increasing day by day due to increase in population, urbanization and industrialization. The world's fossil fuel supply viz. coal, petroleum and natural gas will thus be depleted in a few hundred years. The rate of energy consumption increasing, supply is depleting resulting in inflation and energy shortage. This is called energy crisis. Hence alternative or renewable sources of energy have to be developed to meet future energy requirement. 4. Energy Classification Energy can be classified into several types: 4.1 Primary and Secondary Energy Primary energy sources are those that are either found or stored in nature. Common primary energy sources are coal, oil, natural gas, and biomass (such as wood). Other primary energy sources available include nuclear energy from radioactive substances, thermal energy stored in earth's interior, and potential energy due to earth's gravity. The major primary and secondary energy sources are Coal, hydro power, natural gas, petroleum etc. Primary energy sources are mostly converted in industrial utilities into secondary energy sources; for example coal, oil or gas converted into steam and electricity. Primary energy can also be used directly. 4.2 Commercial Energy and Non Commercial Energy The energy sources that are available in the market for a definite price are known as commercial energy. By far the most important forms of commercial energy are electricity, coal and refined petroleum products. Commercial energy forms the basis of industrial, agricultural, transport and commercial development in the modern world. The energy sources that are not available in the commercial market for a price are classified as non- commercial energy. Non-commercial energy sources include fuels such as firewood, cattle dung and agricultural wastes, which are traditionally gathered, and not bought at a price used especially in rural households. These are also called traditional fuels. Non-commercial energy is often ignored in energy accounting. Fig. 2. Solar PV generation capacity
- 9. The energy transfer/flowmanagement algorithm for a photovoltaic solar home 119 "energizes" our computers, lights, refrigerators, washing machines, and air conditioners, to name only a few uses[1]. We use energy to run our cars and trucks. Both the gasoline used in our cars, and the diesel fuel used in our trucks are made from oil. The propane that fuels our outdoor grills and makes hot air balloons soar is made from oil and natural gas. 4.4 Renewable Energy and Trends in Solar Photovoltaic Energy Production 4.4.1 Energy Scenario The present energy scenario is discussed under categorical division of World, Bangladesh. 4.4.2 World Energy Scenario Global economic recession drove energy consumption lower in 2009 - the first decline since 1982. World primary energy consumption - including oil, natural gas, coal, nuclear and hydro power - fell by 1.1% in 2009. The issues at hand are numerous and include increasing atmospheric carbon dioxide concentrations from greenhouse gas emissions, environmental safety of energy production techniques, volatile energy prices, and depleting carbon based fuel reserves to name a few (Nguyen and Pearce 2010; Choi et al. 2011)[5]. As a result, countries are facing an increasing challenge to diversify energy sources and bringing renewable generation to the forefront of policy discussion. 4.3 Renewable and Non-Renewable Energy All forms of energy are stored in different ways, in the energy sources that we use every day. These sources are divided into two groups --renewable (an energy source that we can use over and over again) and nonrenewable (an energy source that we are using up and cannot recreate in a short period of time). [2] Figure Renewable and nonrenewable energy sources can be used to produce secondary energy sources including electricity and hydrogen. Renewable energy sources include solar energy, which comes from the sun and can be turned into electricity and heat. Wind, geothermal energy from inside the earth, biomass from plants, and hydropower and ocean energy from water are also renewable energy sources. However, we get most of our energy from non- renewable energy sources, which include the fossil fuels --oil, natural gas, and coal[2]. They're called fossil fuels because they were formed over millions and millions of years by the action of heat from the Earth's core and pressure from rock and soil on the remains (or "fossils") of dead plants and animals. Another nonrenewable energy source is the element. Uranium, whose atoms we split (through a process called nuclear fission) to create heat and ultimately electricity. We use all these energy sources to generate the electricity we need for our homes, businesses, schools, and factories. Electricity Fig. 1.1. Renewable Energy Sources and Non-Renewable Energy Sources
- 10. 120 Osama Shaukat In theUnited States, a rise in renewable energy generation has been supported by the availability of federal tax credits and programs in individual states (U.S. Energy Information Administration 2013a)[3]. Many states are implementing renewable portfolio standards, or renewable energy standards, that outline goals to increase electricity generation from renewable resources (U.S. Energy Information Administration 2013a). These policies seek to remove barriers to install renewable generation and can include grant programs, loan programs, and state renewable electricity tax credits. The Database of State Incentives for Renewables & Efficiency (DSIRE) provides an outline of state renewable portfolio standards available throughout the nation (NorthGraph 1.1 : Power generation capacity in world by source, 2008
- 11. The energy transfer/flowmanagement algorithm for a photovoltaic solar home 121 Carolina State University 2013)[8]. In 2012, about 12 percent of U.S. electricity was generated from renewable sources (U.S. Energy Information Administration 2013b). The United States Energy Information Administration states that the five renewable sources most often utilized include biomass, water, geothermal, wind and solar (U.S. Energy Information &!! 4.5 Energy transfer/flow management algorithm The figure 3 depicts the detailed flowchart of the flow energy management algorithm. The main idea is that the photovoltaic production should be utilized as much as possible to reach the self-consumption. The priority is given to the supply of the loads, then the charge of the batteries and at last, the excess energy is fed into the grid. The algorithm operation is based mainly on the PV production, the SOC of the battery and the Ibat which indicates if the battery is in charge or discharge and if it is able to feed the loads or not. Depending on the state of theses parameters one of the modes indicated in the previous paragraph is used. 5. Simulation Results and Discussion In order to test the efficiency of the flow energy management algorithm the satisfaction of the energy demand of the solar home by the photovoltaic system, has been evaluated for two chosen cases. The first one is 'a favorable week' for the PV production relatively to weather conditions in summer and the second one is 'an unfavorable week' in the winter. For each case, depending on the PV produced and the load energy profile a combination of the switches, directs the energy flow to the given target. The simulations were done using the irradiation and the temperature data acquired by the meteorological station of the site of Bou-Ismaïl. The PV electrical production was calculated using the PV system model presented briefly in paragraph 3. The algorithm was performed using the appropriate load profile of each season. The figures 4 and 5 show the states of the K1, K2, K3 and K4 switches and the variation of the photovoltaic production, the energy demand (load profile), the energy battery storage and the grid energy for the summer and winter week respectively. From the results presented in figure 4 and 5, the excess energy produced in the favorable week in summer is 8.5%, this percentage corresponds to an energy of 9kWh which is injected to the grid. On the other hand, in winter, 67% of the energy demand of the unfavorable week is satisfied by the grid (65kWh) this is due to the low photovoltaic energy production during this season. To better explain the energy flow in the studied PV system, two particular cases will be discussed. The first one corresponds to a favorable day in summer (25/08/2015), and the second corresponds to an unfavorable day in winner (03/ 02/2015). As it can be seen on figure 6, between midnight and 5am there is no PV production and the energy demand is very low, the battery is at its low level (25%). Between 5 am and 5 pm, the PV array produces 2kWh. After 5 pm, as the PV production is low, the batteries supply a part of the demand. Between 7 pm and 9 pm all the energy needed is taken from the batteries. After 9 pm the batteries continues to discharge in the loads but the demand is so high that the rest of the energy needed was withdrawn from the grid. Indicates that, for the chosen day it can be noticed that the irradiation is at low level all the day and this conducts to a low PV production. In this case, the energy demand is exclusively satisfied by the grid. The battery bank is all day long near its lowest level (25%). The highest energy demand is situated between 5 pm and 9 pm. 6. Conclusion Thanks to the decreasing of the cost of PV installations, the grid parity comes a reality. Nowadays, it is not profitable as before to sale the produced photovoltaic electricity. It is more interesting for the PV owners to use it for their own needs and try to reach their self-consumption. In this article, as a first step an energy flow management algorithm which directs the energy flow according to priorities set was presented. It appears that even if the energy flow algorithm is necessary it is not sufficient. To go further and to use the PV electricity produced as much as possible, a Home Energy Management (HEM) is needed. This last will allow the balancing between the demand and the generation, by controlling the deferrable loads, reduces the energy consumption for example by using
- 12. 122 Osama Shaukat natural lightas long as possible, exchanging excess energy produced between neighbors instead of drawing off energy from the grid. 7. Acknowledgements We would like to thank Dr Shahzad Ahmad Siddiqui, for his valuable advices on the writing of this research paper. References [1] OECD/IEA, 2013. Transition to sustainable buildings. Strategies and opportunities to 2050. [2] Asare-Bediako B., Ramirez Elizondo L.M, Ribeiro P.F. and Kling W.L., Consideration of Electricity and Heat Load Profiles for Intelligent Energy Management Systems. UPEC 2011. 46th International Universities' Power Engineering Conference. Soest. Germany; 5-8th September 2011. [3] Riffonneau Y., Bacha S., Barruel F. and Ploix S., Optimal Power Flow Management for Grid Connected PV Systems with Batteries. IEEE Trans. Sustain. Energy. 2(3), (2011). [4] Gudi N., Wang L. and Devabhaktuni V., A demand side management based simulation platform incorporating heuristic optimization for management of household appliances. Int J Elec Power, 43(1), (2012) 185-193. [5] Di Giorgio A. and Pimpinella L., An event driven smart home controller enabling consumer economic saving and automated demand side management. Appl Energ, 96, (2012) 92-103. [6] Chen X., Wei T. and Hu S., Uncertainty-aware household appliance scheduling considering dynamic electricity pricing in smart home.IEEE Trans. Smart Grid, 2, (2013) 932-941. [7] Tascikaraoglu A., Boynuegri A.R. and Uzunoglu M., A demand side management strategy based on forecasting of residential renewable sources: A smart home system in Turkey. Energ Buildings, 80, (2014) 309-320. [8] Chikh M., Mahrane A.and Bouachri F., PVSST 1.0 sizing and simulation tool for PV systems. Energy Procedia, 6, (2011) 75-84. [9] Al-Alawi A. and Islam S.M., Demand side management for remote area power supply systems incorporating solar irradiance model, Renew. Energ., 29, (2004) 2027-2036. [10] Sera D., Teodorescu R. and Rodriguez P., PV panel model based on datasheet values. Proc. IEEE Int. Symp. Ind. Electron. http:// dx.doi.org/10.1109/ISIE.2007.4374981 (2007) 2392-2396. [11] Copetti J.B. and Chenlo F., Internal resistance characterization of lead acid battery for PV rates. Proceedings of the 11th European PV Solar Energy. Conference, Monteux, (1992) 1116-1119. [12] Copetti J.B. and Chenlo F., A general battery model for PV system simulation. Prog Photovoltaic Res. Appl. (1993) 283-292. [13] Chekired F., Mellit A., Kalogirou S.A. and Larbes C., Intelligent maximum power point trackers for photovoltaic applications using FPGA chip: A comparative study. Elsevier, Sol. Energy, 101, (2014) 83-99. [14] Chekired F., Larbes C., Rekioua D. and Haddad F., Implementation of a MPPT fuzzy controller for photovoltaic systems on FPGA circuit. Elsevier, Energy Procedia. 6, (2011) 541- 549.
- 13. Numerical analysis offriction factors in smooth and rough microchannels 123 Numerical analysis of friction factors in smooth and rough microchannels MOHD OWAIS QIDWAI1, M.M. HASAN1 and MOHD ARIZ2 1Department of Mechanical Engineering, F/o Engineering and Technology, Jamia Millia Islamia, New Delhi, India 2Global Evolutionary Energy Design, New Delhi, India *E-mail: owais00738@gmail.com Abstract Based on the available literature, the researchers are divided that transition in smooth channel occurs in low Reynolds region or not, and various experimental results contradicts each other. Also, the effect of surface roughness has any considerable effects on early transition behavior or not. The numerical procedures are used these days and various researches are ongoing with the help of numerical methods on studying the flow characteristics. The author aims in understanding the flow behavior in low Reynolds regime between 100 - 1400 for smooth and rough channel with roughness to hydraulic diameter ratio of 0.264. The flow remains in laminar regime for smooth channel and rough channel, and no significant effect of roughness obtains in both the cases. The early transition effects are not seen in the analysis. The friction factor decreases with increase in Reynolds number. Key words : CFD, Friction Factor, Laminar, Reynolds Number, Single phase, Transition, Turbulent, Velocity profile. 1. Introduction For successful operation of electronic devices, the microchannel heat sink is best available option for decades. As per Garimella et al.[1] liquid cooling has lot of potential as discussed in the TCO decision criterion for high density and high computing electronic devices. In the development of liquid cooling microchannel heat sink various researches regarding friction factor[2,3], pressure drop[4] and thermal resistance[5] has been published in recent years. Most of the investigation performed in earlier phase were experimental, but recent development in mathematical models open ways for numerical studies and consistent results were obtained by various researchers[6,7,8], proving the validity of the numerical simulations. Mala and Li[9] performed experimental studies on circular cross sectional microchannel of stainless steel, with one channel with hydraulic diameter in Invertis Journal of Renewable Energy, Vol. 7, No. 3, 2017 ; pp. 123-130 the range of 50-254. The ratio of surface roughness to the hydraulic diameter in the channel varies from 0.69-3.5. They found that for Reynolds number 80 - 2500, the critical Reynolds number obtained was in the range of 300-900. As per their observation early transition is obtained with increase in friction factor. Brackbill and Kandlikar[10] also obtained early transition from laminar to turbulent regime in the range of 572-2367, for introducing saw tooth profile roughness in the microchannel. Many authors published results with higher range of critical Reynolds number. Tam et al.[11,12] obtained critical Reynolds number in the range of 2100-2700. Baviere et al.[13] found critical Reynolds number in the range of 2417- 2822. Based on the literature review, and applicability of numerical procedures to the study of microchannel flow characteristics, the authors aims at finding out the region of transition in the microchannel with the help of friction factor obtained theoretically and DOI No. : 10.5958/2454-7611.2017.00017.0
- 14. 124 Mohd Owais Qidwai,M.M. Hasan and Mohd Ariz through numerical procedures for Reynolds number varying from 100-1400, for single phase water flow in the microchannel with constant thermos-physical properties. Classical conventional theory from open literature consists of relations of friction factor (Darcy and Fanning) with the pressure drop. The fanning friction is defined as 21 2 = w Ff u (1) Darcy friction factor and fanning friction factor are related as 4=D Ff f (2) And Darcy friction factor in terms of pressure drop is given by 2 2 = H D P D f L u (3) Where P is the pressure drop across the length of microchannel. The Poiseuille number (Po) is defined as the product of Darcy friction factor and Reynolds number (Re), where Reynolds number is defined as, Re = HuD (4) The friction factor obtained in the above relation is evaluated based on assumption that Steady, laminar, incompressible flow takes place. The Poiseuille number on the other hand is obtained for the same assumption along with fully developed flow between two parallel infinite plates. For a finite dimension rectangular cross sectional microchannel, Poiseuille number is defined based on the aspect ratio by Shah and London[14]. 2 3 4 5 = 96(1 – 1.3553 + 1.9647 – 1.7012 + 0.09564 – 0.2537 ) Po (5) Various experimental investigations have confirmed that microscale friction factor can be obtained by macroscale theories. Therefore Navier- Stokes equation can be used to study incompressible, Newtonian liquids in laminar regime[15]. Several numerical studies have been performed earlier which confirm the applicability of numerical schemes available for thorough analysis Xu et al.[16], Lee et al.[17], Jones et al.[18]. The present paper aims at investigating the effect of laminar to turbulent transition with the increase in Reynolds number for smooth and rough channel with roughness to hydraulic diameter ratio of 0.264. The numerical scheme is validated with the experimental work of Hao et al.[19]. 2. Mathematical Model Figure 1 Shows the schematic diagram of smooth channel with the dimensions as used in Hao et al.[19]. The average surface roughness in the experimentation was obtained 0.032 micron, on the inside surface of the channel. To study the entire domain is computationally expensive process and hence single channel with exact dimensions are considered for the analysis. For the rough channel as shown in Figure 2, the ribs are introduced only on one side of the channel to analyse its effect on the friction factor and transition regime. The first roughness element is introduced at x= L1=6.8 mm from the inlet boundary and thereafter a pitch value of 2.95 mm is chosen to place the ribs, so that L2= L3= L4= L5= L6= 2.95 mm. Length L1 is selected with a higher value so that fully developed flow is obtained in the region. The ribs have cross- sectional dimension of hr = wr = 50 ?m are introduced Fig. 1. Schematic Diagram of Smooth Microchannel Fig. 2. Ribs of square cross section with hr = wr are introduces in the smooth channel as roughness element
- 15. Numerical analysis offriction factors in smooth and rough microchannels 125 in the smooth channel as roughness element as shown as shown in Figure 3. Three-dimensional geometry of the model for smooth and rough channelis shown in Figure 4&Figure 5 respectively. Assumption 1. The flow takes place under steady state conditions. 2. No phase change during the flow. 3. No slip boundary condition. 4. Viscous dissipation, flow mal distribution and external heat transfer are neglected. 5. Thermophysical properties of the fluid remain constants. 6. Developing flow at inlet boundary. The governing equation is along with the boundary conditions for 3D, laminar incompressible flow in cartesian coordinates are: Continuity equation + + = 0 u w x y z (6) The boundary conditions are 2 2 2 2 2 2 1 + + w = – + + + u u u P u x y z x u u u x y z (7) 2 2 2 2 2 2 1 + + w = – + + + u P u x y z z x y z (8) 2 2 2 2 2 2 1 + + w = – + + + w w w P u x y z z w w w x y z (9) The boundary conditions are At z = 0 At z = L w = Uin, u=v=0 = u = v = 0 w x At x = 0 At x = w u=v=w=0 u=v=w=0 At y = 0 At y = h u=v=w=0 u=v=w=0 The above equation is solved using CFD software CFX., the pressure and velocity distribution are identified in the fluid domain for smooth and rough channels. Fig. 3. Schematic diagram of channel with roughness element Fig. 4. Smooth channel Fig. 5. Rough channel
- 16. 126 Mohd Owais Qidwai,M.M. Hasan and Mohd Ariz 3. Numerical Solution Ansys CFX is used to solve the continuity and momentum equation using Finite volume method (FVM). A mesh of fluid domain is generated for the channel with dimension of length L = 23 mm; height h = 0.146 mm andwidth w = 0.027 mm. The dimensions considered are same as that of Hao et al.[19] for the smooth channel with no ribs. The numerical solution also involves surface roughness effect of 0.032 micron as mentioned by Hao et al.[19]. For a flow of Re = 100 three mesh sizes were used as shown in Table 1. The number of nodes are increased for each solution and the resultant pressure drop from inlet to outlet boundary is mentioned along with the central velocity in fully developed region about 18 mm down the inlet boundary. The convergence criterion adopted for the solution of momentum and continuity equations is to be less than 10-6. 4. Result and Discussion From Fig. 6, it is observed that frictional factor decreases with increase in Reynolds number for both smooth and rough microchannel. There is no significant deviation from Darcy friction factor and numerical frictional factor for both smooth and rough microchannels. However, the slight deviation starts taking place at Re=600 for friction factor in rough microchannel, which is not so significant for which early transition effect in rough channel can be concluded. Though a higher value of friction factor is obtained for Re > 900 in case of rough channel. But, it can be said that there is no effect of surface roughness on transition for flow with low Reynolds number. From Fig. 7 it appears that the velocity profile is parabolic for low Reynolds number, and as the Reynolds number is increased the parabolic velocity Table 1. Grid independence study Mesh Number of Pressure Drop Fully developed Configuration Nodes P (Pa) central velocity (m s–1 ) 1 429229 8482.00 0.955027 2 576000 8478.65 0.950493 3 792000 8474.30 0.948219 Fig. 6. Comparison of Friction Factor obtained using pressure drop for smooth and rough microchannel from numerical solution and C/Re., where C is obtained from equation 5 for smooth channel
- 17. Numerical analysis offriction factors in smooth and rough microchannels 127 Fig. 7. Fully developed velocity profile across the channel at 18 mm down the inlet boundary in smooth channel distribution tends to become blunter, which suggest early transition from laminar to turbulent regime. Fig. 6 shows the shift from Darcy friction profile is started early around Re=600. and considerable shift takes between the Darcy friction factor and numerical friction factor around Re=1000. The flow is still in the laminar regime, but this could be the understanding of earlier authors that some of them may concluded this as the early shift in transition. The velocity profile at Re=900 and Re=1000 are presented in Fig. 9 for comparison. Based on the profile obtained it may be said that it marks the beginning of transition but does Fig. 8. Fully developed velocity profile across the channel at 18 mm down the inlet boundary in rough channel not reach critical Reynolds value in smooth channel. Fig. 8 shows almost parabolic velocity profile except the slight deviation is due to the presence of roughness element in the microchannel. The deviation from parabolic velocity profile increases with the increase in Reynolds number. As shown in Fig. 10, the velocity profile for rough channel at Re = 500 have a sharp parabolic peak, while the profile for Re = 600 is considerably blunter. Both the profiles are deviated from central mean position because of the presence of roughness element. As
- 18. 128 Mohd Owais Qidwai,M.M. Hasan and Mohd Ariz Fig. 9. Comparison of fully developed velocity profile in smooth channel (a) Fig. 10. Comparison of fully developed Velocity profile in rough channel (b) (a) (b) shown in Fig. 11 and Fig. 12, the flow recirculation increases with increase in Reynolds number. Since the flow is not shifted into transition or turbulent phase either, otherwise due to stronger transverse momentum exchange in turbulent regime, the case would have obtained wise versa, with length of flow recirculation zone smaller and blunter profile at Re=1300.
- 19. Numerical analysis offriction factors in smooth and rough microchannels 129 Fig. 11. Flow Recirculation in rough channel for Re=600 Fig. 12. Flow Recirculation in rough channel for Re=1300 5. Conclusion Based on the above discussion, following points may be concluded: 1. Roughness to hydraulic diameter ratio of 0.264 has no significant effect in early transition from laminar to turbulent regime in the range of Re = 100 to 1400. 2. The friction factor has higher value for Re>900 for both the channels, but based on streamwise velocity, it cannot be asserted that profile obtained in Fig. 9 & Fig. 10, is not parabolic in nature. 3. Conventional theories are in good agreement with experimental results mostly. Only in the studies where entrance loss, exit loss and developing flow are not considered, reports deviation from it [2]. Nomenclature Subscript Shear Stress w wall Fluid Density F Fanning Factor ¯u Average flow velocity D Darcy Factor f Friction Factor H Hydraulic Difference P Pressure D Diameter L Length Viscosity aspect ratio (h/w) h Height of Channel w Width of channel References [1] Garimella S.V., Persoons T., Weibel J.and Yeh L.T., "Technological drivers in data centers and telecom systems: Multiscale thermal, electrical, and energy management," Appl. Energy, 107, (2013) 66-80.
- 20. 130 Mohd Owais Qidwai,M.M. Hasan and Mohd Ariz [2] Dai B., Li M. and Ma Y., "Effect of surface roughness on liquid friction and transition characteristics in micro- and mini-channels," Appl. Therm. Eng., 67(1-2), (2014) 283-293. [3] Steinke M.E. and Kandlikar S.G., "Single- phase liquid friction factors in microchannels," Int. J. Therm. Sci., 45(11), (2006) 1073-1083. [4] Gunnasegaran P., Mohammed H. and Shuaib N.H., "Pressure Drop and Friction Factor for Different Shapes of Microchannels," 7-8, (2009). [5] Qu W. and Mudawar I., "Experimental and numerical study of pressure drop and heat transfer in a single-phase micro-channel heat sink," Int. J. Heat Mass Transf., 45(12), (2002) 2549-2565. [6] Lee P.S. and Garimella S.V., "Thermally developing flow and heat transfer in rectangular microchannels of different aspect ratios," Int. J. Heat Mass Transf., 49(17-18), (2006) 3060-3067. [7] Li J. and Peterson G.P., "3-Dimensional numerical optimization of silicon-based high performance parallel microchannel heat sink with liquid flow,"Int. J. Heat Mass Transf.,50(15- 16), (2007) 2895-2904. [8] Sahar A.M., Özdemir M.R., Fayyadh E.M., Wissink J., Mahmoud M.M. and Karayiannis T.G., "Single phase flow pressure drop and heat transfer in rectangular metallic microchannels," Appl. Therm. Eng., 93, (2015) 1324-1336. [9] Mohiuddin Mala G. and Li D., "Flow characteristics of water in microtubes," Int. J. Heat Fluid Flow, 20(2), (1999) 142-148. [10] Brackbill T.P. and Kandlikar S.G., "Effect of Sawtooth Roughness on Pressure Drop and Turbulent Transition in Microchannels," Heat Transf. Eng., 28(8-9), (2007) 662-669. [11] Tam L.M. et al., "The Effect of Inner Surface Roughness and Heating on Friction Factor in Horizontal Micro-Tubes," 44403, (2011) 2971- 2978. [12] Tam L.M., Tam H.K. and Ghajar A.J., "Simultaneous Heat Transfer and Pressure Drop Measurements for a Horizontal Micro- Tube," 38921, (2011) T10122-T10122-8. [13] Conference I., June M. and York N., "ICMM2004-2338," 0, (2016) 1-8. [14] Shah R.K. and London A.L., Laminar Flow Forced Convection in Ducts: A Source Book for Compact Heat Exchanger Analytical Data. (1978). [15] Xu B., Ooti K.T., Wong N.T. and Choi W.K., "Experimental investigation of flow friction for liquid flow in microchannels," Int. Commun. Heat Mass Transf., 27(8), (2000) 1165-1176. [16] Lee P.S., Garimella S.V. and Liu D., "Investigation of heat transfer in rectangular microchannels," Int. J. Heat Mass Transf., 48(9), (2005) 1688-1704. [17] He Y., Shao B.D., Cheng H.M. and Tang Y.J., "Numerical Simulation and Size Optimization of the Flow Channel of Rectangular Micro- Channel Heat Sink," Appl. Mech. Mater., 444- 445, (2013) 568-573. [18] Jones B.J., Lee P.-S. and Garimella S.V., "Infrared micro-particle image velocimetry measurements and predictions of flow distribution in a microchannel heat sink," Int. J. Heat Mass Transf., 51, (2008) 1877-1887. [19] Hao P.-F., Yao Z.-H., He F. and Zhu K.-Q., "Experimental investigation of water flow in smooth and rough silicon microchannels," J. Micromech. Microeng, 16, (2006) 1397-1402.
- 21. Effect of operatingconditions on the performance of vapor absorption refrigeration system 131 Effect of operating conditions on the performance of vapor absorption refrigeration system MD. MERAJ1, RASHID IMAM2 and MD. ASHFAQUE ALAM3 1 Engineering & Technology, Department of Mechanical Engineering, JMI, New Delhi-110025, India 2 Department of Mechanical Engineering, Vishveshwarya Group of Institution, Noida, UP-203207, India 3 Department of Mechanical Engineering, NIT, Jamshedpur, Jharkhand-831014, India *E-mail: md.meraj1221@gmail.com Abstract In this paper, a thermodynamic analysis of single effect vapor absorption refrigeration system (VARS) is carried out. LiBr-Water is used as working fluid during the analysis. This analysis is carried out to study the effect of operating conditions on the performance of vapour absorption refrigeration system. The generator temperature, the condenser temperature and the evaporator temperature are taken as operating conditions for the analysis. From the result and discussion, it is found that the coefficient of performance (COP) of the system increases with increasing the generator and evaporator temperature up to a certain limit but decreases with increasing the condenser temperature. Key words : VARS, LiBr-Water, COP, generator temperature, condenser temperature, and evaporator temperature. 1. Introduction Present day, the rate at which the non-renewable energy sources are being utilized has become a major concern for the world. The future prospect of non- renewable sources of energy like coal, natural gases and petroleum products are going to be scarce. To preserve these energy sources for our future generation, we have to reduce its consumption by using alternative sources of energy like solar energy, geothermal energy, waste heat, hydro-energy etc. There are many ideas has been developed to utilized these alternative sources of energy and researches has also been under the processes to improve the systems which utilize these sources. In this regard vapor absorption refrigeration systems are the best alternative system to reduce the rate of consumption of non-renewable energy sources. Vapor absorption refrigeration system (VARS) invented much earlier than vapor compression refrigeration system (VCRS) but due to law coefficient of performance and inven-tion of high performance Invertis Journal of Renewable Energy, Vol. 7, No. 3, 2017 ; pp. 131-136 refrigerant for VCRS during Second World War, the popularity of VARS decreases. But, in present scenario of energy crisis has forced the world to use vapor absorption refrigeration system in place of vapor compression system to reduce the rate of consumption of non-renewable energy sources. VARS is one of the best replacement for the VCRS, from the energy point of view as well as environ- mental,because this system works on law grade of energy such as solar energy, geothermal-energy, waste-heat from industries, cheap available energy etc., and VARS uses such type of refrigerant-absorbent combination which do not contributes in global warming and ozone depletion like LiBr2-H2O, H2O -NH3, LiNO3-NH3, NaSCN-NH3 etc. From construction point of view absorption system and compression system are similar but only the differenceis that the compressor in case of compression system is replaced by the generator- solution heatexchanger-pump-absorber assembly which is used to circulate the working fluid entire the system by creating pressure differ-ences[8-9]. DOI No. : 10.5958/2454-7611.2017.00018.2
- 22. 132 Md. Meraj, RashidImam and Md. Ashfaque Alam Aphornratna. S and Sriveerakul. T[1] performed an experimental investigation of single-effect absorption. They have usedaqueous lithium-bromide as working fluid. They have done their analysis on 2 kW cooling capacity experimental refrigerators which was tested with various operating temperatures. From results, it was found that the system performance is strongly dependent on thesolution circulation ratio (SCR)and also on solution heat exchanger. They have resulted that energy input to the generator is reduced to 60% by usingsolution heat exchanger. Darwish A. N. et. al.[2] presented the performance analysis and evaluation of the roburabsorption-refrigeration water-ammonia (ARWA) system.The system is analysed by using aspen plus flowsheet simulator.They analysed COP, heat duties of the evaporator, absorber, and the condenser and the flow rate of refrigerant passing through the evaporator.The results obtained by them is compared with some manufacturer data and experimental data reported in the journal literature. Saghiruddin and Siddiqui[3] perfor-medthe economic analysis of ordinary and evacuated tubular type flat- plate collectors for operating absorption cycles with and without heat recovery absorber. Water- Ammonia, NaSCN3 and LiNO3-NH3 mixtures has been selected as the working fluids in the cycles. Use of heat recovery absorber, in addition to the primary absorber in the conventional absorption cycles lead to the improvement in COP by 20-30 % in Water- Ammonia and 33-36% in NaSCN-NH3 and LiNO3- NH3 mixtures. Kaushik and Kumar[4] performed thermodynamic analysisof two-stage vapour absorption refrigeration system using ammonia- water (NH3-H2O) and ammonia-lithium nitrate (NH3-LiNO3) as working fluid. They havefound that minimum evaporator temperature is achieved in NH3-LiNO3 combination, and COP of the NH3- LiNO3 is higher than the NH3-H2O. Marcos et al.[5] performed optimization of COP in single and double effect LiBr-H2O absorption chillers. They analyzed water cooling system and air cooling system and found that solution concentration is an important parameter to improve COP. With this, they also found the crystallization limit. Their study showed that COP of single effect becomes 0.85 to 0.74 in water cooled and 0.72 to 0.65 in air cooled system. Similarly, in double effect cycle it was 1.18-1.2 in the water cooled and 1.15-1.07 in the air cooled system. Talbi M. M. and AgnewB.[6] presented an exergy analysis on a single effect absorption refrigeration cycle with lithium-bromide-water as the working fluid pair. They calculated the loads on each component of absorption system, exergy of each components and total exergy of system. However, some of the research has been done on the energy and exergy analysis of the double effect absorption refrigeration system[10, 11,12]. From literature survey it is absorbed that there is no work has been done on the effect of operating condition on the performance of vapour absorption refrigeration system. Therefore this paper present a thermodynamic analysis which shows variation of performance of VARS with the operating conditions like the generator temperature, the condenser temperature and the evaporator temperature in the presented analysis. 2. System Description A schematic diagram of vapour absorption refrigeration system is shown in Fig. 1. on which analysis has beendone. This system consists of an Evaporator (E), Absorber (A), Generator (G), Condenser(C), Throttle valve (TV), Pre-cooler (PC), Fig. 1. Schematic diagram of single effect vapour absorption refrigeration system.
- 23. Effect of operatingconditions on the performance of vapor absorption refrigeration system 133 Preheater (PH) and Pump.The operating temperature of generator, condenser, evaporator and absorber is shown byTg, Tc, Te and Ta respectively in figure 1. There are two pressure levels in this system; the generator and condenser operate at high pressure level while evaporator and absorber operate at low pressure level. LiBr-H2O mixture is used as working fluid in the system in which LiBr works as an absorbent and H2O works as refrigerant. LiBr-H2O mixture circulate in the generator-solution heat exchanger-pump-throttle valve-absorber assembly, generating the water vapor in the generator which flows through condenser, throttle valve, solution heat exchanger and evaporator. The working fluid having the different state condition while circulating inthe system is shown in figure by different state point. 3. Thermodynamic Analysis The thermodynamic analysis of single effect vapour absorption refrigeration system is based on the application of mass, concentration and energy conservations. For performing these applications, there are some assumptions are made to simplify the analysis, these are as follows : The systems are in steady state. The refrigerant (water) at the outlet of the condenser is saturated liquid. The refrigerant (water) at the outlet of the evaporator is saturated vapour. Temperature of absorber and condenser are same. Expansion process in throttle valve is isenthalpic. Heat loss to environment and pressure drop in the systems are negligible. No leakage of air in the system. Mass, concentration and energy conservation equation of generator are respectively expressed by: 5 6 10 m m m (1) 5 5 6 6 10 10 m X m X m X (2) 6 6 10 10 5 5= –EQ m h m h m h (3) Mass, concentration and energy conservation equation of absorber are respectively expressed by: 3 2 12m m m (4) 3 3 2 2 12 12m X m X m X (5) 2 2 12 12 3 3= –aQ m h m h m h (6) Energy consumption of solution pump can be defined as: 4 4 3 3= –pW m h m h (7) Energy conservation of condenser and evaporator is given by : 6 6 7 7= –cQ m h m h (8) 1 1 9 9= –eQ m h m h (9) The coefficient of performance of the system can be expressed as follows : /( )e E DCOP Q Q W (10) A computer program has been developed using FORTRAN language to carry out the system simulation of the single effect vapour absorption refrigeration system. The analysis has been done for 1 Ton of refrigeration. Properties of LiBr-water have been evaluated by the correlation from[7]. The calculation parameters for the simulation are summarized in table 1. Table 1 Calculation Parameters Evaporator temperature (Te in °C) 5-(2.5)-12.5 Absorber temperature (Ta in °C) 30-(5)-40 Condenser Temperature (Tc in °C) 30-(5)-40 Effectiveness of pre-heater (%) 75 Effectiveness of pre-cooler (%) 75 Cooling load (kJ/h) 12600 4. Results and Discussion Figure 2 shows the variation of coefficient of performance (COP) of single effect vapour absorption system with genera-tor temperature at which heat is supplied at different values of evaporator temperature (i.e. Te = 5.0, 7.5, 10.0 and 12.5 °C), the condenser temperature and absorber temperature at which heat is rejected to the surroundings are equal and kept constant i.e. Ta = Tc = 30 °C. From this graph, it is seen that the coefficient of performance of single effect increases drastically from low values at low generator temperature, reach to maximum value and
- 24. 134 Md. Meraj, RashidImam and Md. Ashfaque Alam Fig. 3. Variation of coefficient of performance (COP) with condenser temperature (Tc) for different evaporator temperature (Te). Fig. 2. Variation of coefficient of performance (COP) with generator temperature (Tg) for different evaporator temperature (Te).
- 25. Effect of operatingconditions on the performance of vapor absorption refrigeration system 135 then with further increase in the increase in the generator temperature slightly decreases i.e. the results show an optima. From this result it is also observed that with increasing the evaporator temperature for a particular generator coefficient of performance increases. Figure 3 shows the variation of coefficient of performance (COP) of single effect vapour absorption system with con-denser temperature at which heat is rejected to surrounding at different values of evaporator temperature (i.e. Te = 5.0, 7.5, 10.0 and 12.5 °C). From this graph, it is seen that the coefficient of performance decreases gradually with increases in condenser temperature for a particular evaporator temperature. From this result it is also observed that with decreasing the evaporator temperature for a particular condenser temperature coefficient of performance decreases. Figure 4 shows the variation of coefficient of performance of single effect absorption system with evaporator temperature at which the refrigerated space should be maintained at different values of condenser temperature (Tc = 30, 35, 40 & 45 °C). From the graph, it is observed that the coefficient of performance increases gradually with increase in evaporator temperature for a particular condenser temperature. From this result it is also observed that with increasing condenser temperature for a particular evaporator temperature coefficient of performance increases. 5. Conclusions From the above result and discussion it is concluded that coefficient of performance system has strong effect on the operation condition. The performance of system is dependent on generator temperature, evaporator temperature and condenser temperature simultaneously. It is found that the coefficient of performance of presented single effect vapour absorption refrigeration system increases with increasing the generator and evaporator temperature up to a certain limit but decreases with increasing the condenser temperature. Therefore, to obtain efficient performance of single effect vapour absorption refrigeration system, it should be operated at moderate generator, evaporator and condenser temperature of vapour absorption refrigeration. Fig. 4. Variation of coefficient of performance (COP) with evaporator temperature (Te) for different condenser temperature (Tc).
- 26. 136 Md. Meraj, RashidImam and Md. Ashfaque Alam References [1] Aphornratana S. and Sriveerakul T., "Experimental studies of a single-effect absorption refrigerator using aqueous lithium- bromide: Effect of operating condition to system performance", Experimental thermal and fluid science, 32, (2007) 658-669. [2] Darwish N.A., Al-Hashimi S.H. and Al- Mansoori A.S., "Performance analysis and evaluation of a commercial absorption- refrigeration water-ammonia (ARWA) system", International Journal of Refrigeration, 31, (2008) 1214-1223. [3] Saghiruddin and Siddiqui M.A., "Economic analysis and performance study of three ammonia-absorption cycles using heat recovery absorber", Energy Conversion and Management, 37, (1996) 421-432. [4] Kaushik S.C. and Kumar R., "Thermodynamic study of two stage vapour absorption refrigeration system using NH3 refrigerant with liquid solid absorbents", Energy Conversion and Management, 4, (1985) 427-431. [5] Marcos J.D., Izquierdo M. and Palacios E., "New method for COP optimization in water- and air-cooled single and double effect LiBr- water absorption machines", International Journal of Refrigeration, 34, (2011) 1348-1359. [6] Talbi M.M. and Agnew B.,"Exergy analysis: an absorption refrigerator using lithium bromide and water as working fluid", Applied Thermal Engineering, 20, (2011) 619-630. [7] Kaita Y., "Thermodynamics properties of lithium bromide-water solutions at high temperatures", International Journal of Refrigeration, 24, (2001) 374-390. [8] Cengel Y.A. and Boles M.A., "Thermo- dynamics an engineering approach", Tata McGraw Hill Education Private Limited New Delhi, (2008). [9] Arora C.P., "Refrigeration and Air Conditioning", McGraw Hill Education Private Limited New Delhi, (2009). [10] Kaushik S.C. and Arora A., "Energy and exergy analysis of single effect and series flow double effect water-lithium bromide absorption refrigeration systems", International Journal of Refrigeration, 32, (2009) 1247- 1258. [11] Mairaj M., Siddiqui S.A. and Hafiz A., "Energetic and exergetic analysis of some models of vapour absorption chillers using lithium bromide and water", Journal of Basic and Applied Engineering Research, 2(4), (2015) 326- 329. [12] Meraj M., Hafiz A. and Ahmad M.J., "Second Law Analysis of Series and Parallel Flow Double Effect Vapor Absorption Chiller", IOSR Journal of Me-chanical and Civil Engineering, 13(3), (2016) 104-110.
- 27. Study and analysisof plate type heat exchanger 137 Study and analysis of plate type heat exchanger NUMAN ANSARI1, MOHD. ISLAM1 and SAYYED HAIDER2 1 Department of Mechanical Engg. Faculty of Engg.& Technology, Jamia Millia Islamia,New Delhi-110025,India 2 Department of Mechanical Engg., Al-Falah School of Engineering & Technology, Haryana-121004, India *E-mail: numan.ansari5@gmail.com Abstract With the emerging trend in technology heat transfer dissipation from those device which required certain temperature for their efficient operation in the major problem. In this regards high technology is required to cool such devices. As far as heat exchanger is the major solution of such problem. Plate heat exchanger are the exchangers having a number of plates combined together having different corrugation for transfer the heat between the two fluid. An experimental investigation are done on the plate type heat exchanger to estimate the value of heat transfer characteristics by changing the mass flow rate and study the pressure drop. And from the experiment estimate the value of heat transfer characteristics for hot and cold fluid, nusselt number, pressure drop and effectiveness. It is shown that Nusselt number, pressure drop and effectiveness are depends on the mass flow rate. This paper is concerned with the plate type heat exchanger and investigation is carried out by considering the different works published in the field of heat exchangers. Key words : Plate heat exchanger (PHX), heat exchanger (HXCH), chevron angle, gaskets, plate, flow arrangement, domestic hot water (DHW). 1. Introduction Plate heat exchanger (PHE) are the exchangers having a number of plates combined together having different corrugation for transfer the heat between the two fluid. Plate heat exchanger are widely used in the industry and are commonly designed with the corrugated channel surface resulting in the enhanced heat transfer performance by increasing the area over which heat transfer take place and generating a vigorous mixing of effect within the working fluid. Component of plate type heat exchanger are shown in figure. As shown in Figure 1, the plate heat exchanger is basically a series of individual plates pressed between two heavy end covers. These plates are gasketed, welded or brazed together depending on the application of the heat exchanger. The basic geometry of plates used in plate heat exchanger is shown in figure 2. Stainless steel is a commonly used metal for the plates because of its ability to withstand Invertis Journal of Renewable Energy, Vol. 7, No. 3, 2017 ; pp. 137-141 high temperatures, its strength, and its corrosion resistance[1]. The entire assembly is held together by the tie bolts. This has a major advantage over a conventional heat exchanger in that the fluids are exposed to a much larger surface area because the fluids spread out over the plates. This facilitates the transfer of heat, and greatly increases the speed of the temperature change. The concept behind a heat exchanger is the use of pipes or other containment vessels to heat or cool one fluid by transferring heat between it and another fluid. In most cases, the exchanger consists of a coiled pipe containing one fluid that passes through a chamber containing another fluid. The walls of the pipe are usually made of metal or another substance with a high thermal conductivity to facilitate the interchange, whereas the outer casing of the larger chamber is made of a plastic or coated with thermal insulation, to discourage heat from escaping from the exchanger. DOI No. : 10.5958/2454-7611.2017.00019.4
- 28. 138 Numan Ansari, Mohd.Islam and Sayyed Haider Chevron Angle , Typically varying from 20° to 65°, is the measure of softness (small , low thermal efficiencyand pressure drop) and hardness (large , high thermal efficiency and pressure drop) of thermal and hydraulic characteristics of plates. Some authors define "/2- " as the chevron angle[1]. Pressure drop (p) is in direct relationship to the size of the plate heat exchanger. If it is possible to increase the allowable pressure drop, and incidentally accept higher pumping costs, then the heat exchanger will be smaller and less expensive. As a guide, allowable pressure drops between 20 and 100 kPa are accepted as normal for water/water duties. 2. PHE Parts and their function Plate with gaskets Plate heat exchangers consist of pressed plates; the separating gaskets between each plate; the end plates used to clamp the plate stack together and the frame to hold the plate stack in place. The plate is pressed metal. A wide range of metals and corrugation shapes can be used that suit the chemical, flow and corrosive properties of the media passing across the plates[4]. Plates for an exchanger have indentations and corruga-tions to encourage more turbulent flow across them and to make thinner films of media to promote better heat transfer. The headers to feed and remove the fluids pass through all plates. Unwanted headers are blocked off on a plate. Gaskets, each plate have a gasket that produces a sealing and channel system through the entire plate pack in which the two heats exchanging media flow in a counter-current direction. The circular portion of a gasket stops the fluid from going across the heat transfer plate and sends it to next open channel. The remaining portion or field gaskets directs the opposing fluid across the heat transfer surface. The gaskets separate the plates and create the thin chamber through which the fluid film swirls and flows. They also serve the purpose of directing the media from the entry port to the exit port. The frame is made up of thick steel pressure retaining parts, the fixed cover and the movable cover that when pulled together with the tightening bolts from the pres-sure retaining structure for the plates. The carrying bar and guide bar act as a carrier and guide both the plates and movable cover. The heat exchangers plates, which make up the heat transfer surface, are clamped between two plates of steel with the use of the tightening bolts. The heat exchanger construction allows a plate heat exchanger to be easily opened for inspection and cleaning. In a plate heat exchanger corrugated plate like this one are used between the barrier of hot fluid and cold fluid. The corrugated plate increases the area of heat transfer. PHE plate usually made of advance material like titanium which makes plate stronger and durable. The thickness of plate is 0.5 mm to 0.6 mm. Fig. 1. Various parts of PHE Fig. 2. Geometry of plate [1]
- 29. Study and analysisof plate type heat exchanger 139 Flow arrangement, Plate heat exchangers are very easy counter flow where hot and cold fluid flow in opposite direction as shown in figure. The plate is placed in such a way that the flow bond by each successive plate alter-nates the hot fluid and cold fluid. The heat exchangers plates with gaskets are arranged in an alternating pattern of left hand flow and right hand flow to direct the fluids in an opposing direction within the heat exchanger. The complete assembly of all the plates and gaskets is called the "Plate Pack". 3. Modelling of PHX and analysis of heat transfer performance The experimental set up presents the principal geometry parameters and illustrates the experimental setup established to investigate the heat transfer characteristics in the corrugated channel for different flow conditions. The basic components of the experimental apparatus include a water loop, an air loop, and a measurement system. The water loop comprises a water tank containing a heater, a pump, a flow meter, and a temperature controller. Fig. 5. Flow arrangement of PHE Fig. 3. Plate with gaskets [4] port ring gasket Flow ports Fig. 4. Gaskets [4] Fig. 6. Experimental set up of PHE Importantly, all of the components in the water system are thermally insulated such that wall temperature of the corrugated channel can be maintained at a nearly constant temperature. The air loop consists of the test section containing the corrugated channel, a blower, an air flow meter, and a number of valves which enable the flow rate to be adjusted. Additionally, a flow straightner can be installed at the entrance of the test section to maintain a uniform inlet flow. The test section is constructed from Galvanized Iron (GI) sheets, each with corrugated surface on both the sides. The two cases are clipped between two corrugated plates, metal plates to form the corrugated channel through which the working fluid is passed. During experiments, hot water was flowed through the two hollow cases to maintain the channel surfaces at an approximately constant temperature and thermocouples wrapped
- 30. 140 Numan Ansari, Mohd.Islam and Sayyed Haider in copper tubes and inserted in upper metal plates were used to record the corresponding air temperature. The side effect was avoided by specifying a small aspect ratio for the channel such that variations in the channel height could be neglected. In the experiments, the temperature distribution in the horizontal, middle plane of the channel was monitored using thermocouples positioned at different locations along the length of the channel in the water loop and in air loop.The temperature of the inlet and outlet water is measured using mercury thermometers. The pressure of the air at the inlet and outlet of the test section is measured using the pressure gauges. 4. Results and Discussion The experiments described herein represent the study of fully developed heat transfer for water flowing in a cor-rugated duct. The duct had a corrugation angle of 29 degree and interval spacing equal to the corrugation height. The Reynolds number based on hydraulic diameter was varied from 1500 to 25,000, and the prandtl number was varied from 4 to 8. Table 2. Counter Flow Condition (Air) Air Position Distance Xm TX Tb hx Nux × 103 channel of X (cm) ºC ºC (KW/m2k) Ma thermocouple (cm) 2.0×10–3 kg/s 1 9.7 29.1 42 43.75 7.13 45.43 2 48.5 66.5 45.5 47.25 9.43 66.93 3 84.5 49 hxm=8.28 Nux=56.18 Table 1. Counter Flow Condition (Water) Surface temperature for 100 LPM Water Position Distance Xm TX Tb hx Nux × 103 channel of X (cm) ºC ºC (KW/m2k) Mw thermocouple (cm) 0.133 kg/s 6 12 11.25 70 67.5 88.30 8.79 5 10.5 29.25 65 66.5 38.25 12.73 4 48 46.75 68 66 85.43 8.43 3 45.5 66.25 64 65 18.57 6.01 2 87 86.20 66 64.75 81.93 7.34 1 85.5 63.5 hxm=62.49 Nux=8.66 Experimental investigation are done by changing the mass flow rate in the experimental setup and study the pressure drop and heat transfer characteristics in the corrugated channel of the plate heat exchanger. The experimental data are substituted into the correlations to identify the characteristics and effect on the Nusselt number by changing the mass flow rate. The analytical results show that local Nu and effectiveness are affected by the change into the flow condition for a given set. 5. Conclusions Although a conclusion may review the main points of the paper, do not replicate the abstract as the conclusion. A conclusion might elaborate on the importance of the work or suggest applications and extensions. Authors are strongly encouraged not to call out multiple figures or tables in the conclusion- these should be referenced in the body of the paper. 6. Conclusions The authors wish to thank A, B, C. This work was supported in part by a lished manuscript).
- 31. Study and analysisof plate type heat exchanger 141 References [1] Bridle J.S., "Probabilistic Interpretation of Feedforward Classification Network Outputs, with Relationships to Statistical Pattern Recognition," Neurocomputing-Algorithms, Architectures and Applications, Fogelman- Soulie F. and Herault J., eds., NATO ASI Series F68, Berlin: Springer-Verlag, (1989) 227-236. (Book style with paper title and editor). [2] Chen W.-K., Linear Networks and Systems. Belmont, Calif.: Wads-worth, (1993) 123-135. [3] Poor H., "A Hypertext History of Multiuser Dimensions," MUD History, http://www.ccs. neu.edu/home/pb/mud-history.html. 1986. (URL link *include year). [4] Elissa K., "An Overview of Decision Theory," unpublished. (Unplublished manuscript). [5] Nicole R., "The Last Word on Decision Theory," J. Computer Vision, submitted for publication. (Pending publication). [6] Kaufman C.J., Rocky Mountain Research Laboratories, Boulder, Colo., personal communication, (1992). (Personal communication). [7] Coming D.S. and Staadt O.G., "Velocity- Aligned Discrete Oriented Polytopes for Dynamic Collision Detection," IEEE Trans. Visualization and Computer Graphics, 14(1), (2008) 1-12, doi:10.1109/TVCG.2007.70405. (IEEE Transactions ) [8] Bingulac S.P., "On the Compatibility of Adaptive Controllers," Proc. Fourth Ann. Allerton Conf. Circuits and Systems Theory, (1994) 8-16. (Conference proceedings). [9] Goto H., Hasegawa Y. and Tanaka M., "Efficient Scheduling Focusing on the Duality of MPL Representation," Proc. IEEE Symp. Computational Intelligence in Scheduling (SCIS '07), (2007) 57-64, doi:10.1109/SCIS.2007. 367670. (Conference proceedings). [10] Williams J., "Narrow-Band Analyzer," PhD dissertation, Dept. of Electrical Eng., Harvard Univ., Cambridge, Mass., (1993). (Thesis or dissertation). [11] Reber E.E., Michell R.L. and Carter C.J., "Oxygen Absorption in the Earth's Atmosphere," Technical Report TR-0200 (420- 46)-3, Aerospace Corp., Los Angeles, Calif., Nov. (1988). (Technical report with report number) [12] Hubert L. and Arabie P., "Comparing Partitions," J. Classification, 2(4), (1985) 193-218. (Journal or magazine citation). [13] Vidmar R.J., "On the Use of Atmospheric Plasmas as Electromagnetic Reflectors," IEEE Trans. Plasma Science, 21(3), (1992) 876-880, available at http://www.halcyon.com/pub/ journals/21ps03-vidmar, (URL for Transaction, journal, or magzine). [14] Martinez J.M.P., Llavori R.B., Cabo M.J.A.and Pedersen T.B., "Integrating Data Warehouses with Web Data: A Survey," IEEE Trans. Knowledge and Data Eng., preprint, 21 Dec. 2007, doi:10.1109/TKDE.2007.190746.(PrePrint).
- 32. 142 Hasan Shamim, ShadabAhmad and Shah Alam Study of CFD approach to discretise first derivative of partial differential equation HASAN SHAMIM, SHADAB AHMAD and SHAH ALAM Jamia Millia Islamia, New Delhi *E-mail: hasanshmm@gmail.com Abstract The process of converting governing partial differential equations into algebraic equation is known as discretisation. The finite difference method is one of the most powerful and simple method to convert partial differential equation (PDE) into algebraic equation. The accuracy of finite difference method increases with refining grid size. In this present study the finite difference method (FDM) is used to determine the shear stress over a flat plate. This shear stress is function of velocity gradient as given by Newton's Law of viscosity. The first derivative of velocity gradient is replaced by forward difference, Reward difference and the central difference approaches. The exact solution of equation is compared with approximate solution and the errors are determined for different values of grid points spacing. It is found that central approach is more accurate than forward and reward approaches. The effect of order of accuracy such as O(y), O(y)2, O(y)3 has been also studied. It shows as order of accuracy increases, computational errors decreases. Key words : Discretisation, PDE, Finite difference method, Forward difference, Reward difference, Central difference, Order of accuracy, Mesh spacing, 1. Introduction Computational Fluid Dynamics (CFD) provides a qualitative (sometimes even quantitative) prediction of fluid flows by means of mathematical modelling, Numerical methods and software tools[1]. These advantages of CFD over experiment based approach are given as : Substantial reduction of lead time and cost of new designs. Ability to study systems where controlled experiments are difficult or impossible to perform. Ability to study systems under hazardous conditions. The main problem with CFD is that code user must have high skilled position. The assumptions, made regarding type of flow e.g. two dimensional, three dimensional, Compressible, Non compressible, Invertis Journal of Renewable Energy, Vol. 7, No. 3, 2017 ; pp. 142-146 Viscous, Non viscous. Laminar, Turbulent should be realistic and must also be able to handle convergence, consistent, stability associated problems in the solution[2]. However there are several applications of CFD in different fields of daily life and engineering. While analysing problems with CFD approach, we get PDE[3]. These equations are first converted into algebraic equations;only then solution of PDE is possible. This process is known as Discretisation. There are three different methods to discretise the PDEs into algebraic equations. These methods are finite difference method (FDM), Finite element method (FEM) and Finite volume method (FVM)[4]. In this present analysis we are using FDM to discretise PDE and observe the accuracy of solution. 2. Statement of Problem In this we want to determine shear stress over a flat plate, when air is flowing over it. The velocity profile is vertical direction is assumed as : DOI No. : 10.5958/2454-7611.2017.00020.0
- 33. Study of CFDapproach to discretise first derivative of partial differential equation 143 = 1 – y L u c e Where c = 480 L is the length of plate (Let L = 3 cm) We determine shear stress at a distance of 0.35 cm with step size of 0.10 cm, 0.15 cm, 0.20 cm, 0.25 cm, 0.30 cm, 0.35 cm, 0.40 cm. Dynamic viscosity of air at 15°C, µ = 1.81×10–5 Pa-s 3. Methodology i. On the basis of above parameters, We want to find out shear stress near wall by using Newton's law of viscosity, Given as : wall = du dy wall ii. The velocity gradient which is first order differential equation is discretised by using Finite difference approach. iii FDM for 1st order accuracy i.e. forward, reward and central difference approaches are used to discretise the equation. The expressions are as given below[1,5]: 1 , i, j– + O( y) i ju uu dy y Forward , i, j – 1– + O( y) i ju uu dy y Reward 1 2 , i, j – 1– + O( y) * i ju uu dy y Central iv FDM for 2nd order accuracy i.e. forward, reward/backward & central difference approaches are used to discretise the equation[6]. The expressions are as given below:- 2 3 2 , i, j + 1 , 2 + 4* – * + O( y ) * i j i ju u uu dy y Forward 23 2 , i, j – 1 , 2 * – 4* + O( y ) * i j i ju u uu dy y Reward 2 1 24 4 , i, j + 1 , , – 2 – + 4* – * + O( y ) * i j i j i ju u u uu dy y Central v. FDM for 3rd order accuracy i.e. forward, reward/backward & central difference approaches are used to discretise the equation[7, 8]. The expressions are as given below : 3 13 18 11 6 , i, j + 2 , , 3 * + 9* * – * + ( y ) * i j i j i ju u u uu dy y Forward 2 311 9 2 6 , i, j –1 , – , – 3 * – 18* * – * + ( y ) * i j i j i ju u u uu dy y Reward 3 1 1 2 32 18 18 9 2 12 , i, j +2 , , , , – 3 * – 9* * – * * – * + O( y ) * i j i j i j i j i ju u u u u uu dy y Central vi The approximate solution of equation is compared with exact solution at (y=.35) for different step sizes (y) i.e. 0.10, 0.15, 0.20, 0.25, 0.30, 0.35, 0.40. vii The effect of order of accuracy on solution has been also observed. The order of accuracy is taken as O(y), O(y)2, O(y)3 for FORWARD, BACKWARD/REWARD & CENTRAL difference method. viii The errors are also represented in graph for different step sizes and order of accuracy.
- 34. 144 Hasan Shamim, ShadabAhmad and Shah Alam 4. Result Table 1. Errors in forward difference approach for 1st, 2nd & 3rd order of accuracy S.l. No. Step size Error (%) Error (%) Error (%) O(Y) O(Y)2 O(Y)3 1 0.1 1.65 0.04 0.001 2 0.15 2.46 0.08 0.003 3 0.2 3.26 0.14 0.007 4 0.25 4.05 0.22 0.013 5 0.3 4.84 0.31 0.022 6 0.35 5.61 0.42 0.035 7 0.4 6.38 0.54 0.051 Table 2. Errors in reward difference approach for 1st, 2nd & 3rd order of accuracy S.l. No. Step size Error (%) Error (%) Error (%) O(Y) O(Y)2 O(Y)3 1 0.1 -1.69 0.04 -0.001 2 0.15 -2.54 0.09 -0.003 3 0.2 -3.41 0.16 -0.008 4 0.25 -4.28 0.25 -0.016 5 0.3 -5.17 0.36 -0.028 6 0.35 -6.07 0.50 -0.046 7 0.4 -6.97 0.66 -0.070 Table 3. Errors in central difference approach for 1st, 2nd & 3rd order of accuracy S.l. No. Step size Error (%) Error (%) Error (%) O(Y) O(Y)2 O(Y)3 1 0.1 -0.02 0.04 -3.70E-05 2 0.15 -0.04 0.08 -1.88E-04 3 0.2 -0.07 0.15 -5.94E-04 4 0.25 -0.12 0.23 -1.45E-03 5 0.3 -0.17 0.33 -3.01E-03 6 0.35 -0.23 0.46 -5.58E-03 7 0.4 -0.30 0.60 -9.54E-03 Table 1. 1st order accuracy for forward, reward and central difference accuracy S.l. No. Step size Error (%) Error (%) Error (%) forward reward central 1 0.1 1.65 -1.69 -0.02 2 0.15 2.46 -2.54 -0.04 3 0.2 3.26 -3.41 -0.07 4 0.25 4.05 -4.28 -0.12 5 0.3 4.84 -5.17 -0.17 6 0.35 5.61 -6.07 -0.23 7 0.4 6.38 -6.97 -0.30
- 35. Study of CFDapproach to discretise first derivative of partial differential equation 145 Fig. 2. Errors in forward (series 1), reward (series 2) and central (series3) approach using O( y)2 Table 2. 2nd order accuracy for forward, reward and central difference approach S.l. No. Step size Error (%) Error (%) Error (%) forward reward central 1 0.1 0.04 0.04 0.04 2 0.15 0.08 0.09 0.08 3 0.2 0.14 0.16 0.15 4 0.25 0.22 0.25 0.23 5 0.3 0.31 0.36 0.33 6 0.35 0.42 0.50 0.46 7 0.4 0.54 0.66 0.60 Fig. 1. Errors in forward (series 1), reward (series 2) and central (series3) approach using O( y) ERROR(%)ERROR(%)
- 36. 146 Hasan Shamim, ShadabAhmad and Shah Alam Table 3. 3rd order accuracy for forward, reward and central difference approach S.l. No. Step size Error (%) Error (%) Error (%) forward reward central 1 0.1 0.001 -0.001 -3.70E-05 2 0.15 0.003 -0.003 -1.88E-04 3 0.2 0.007 -0.008 -5.94E-04 4 0.25 0.013 -0.016 -1.45E-03 5 0.3 0.022 -0.028 -3.01E-03 6 0.35 0.035 -0.046 -5.58E-03 7 0.4 0.051 -0.070 -9.54E-03 Fig. 3. Errors in forward (series 1), reward (series 2) and central (series3) approach using O( y)3 ERROR(%) 5. Conclusion The effect of grid spacing and order of accuracy has been observed by calculating shear stress. It is found the central difference approach has least error for fix step size compare to forward and backward difference approach. Higher order of accuracy has least error in each approach. References [1] Anderson D. John, "Computational Fluid Dynamics: Basic with applications" Tata McGraw Hill, (2012). [2] Versteeg H. K. and Weeratunge Malalasekera, "An Introduction to Computational Fluid Dynamics" (2007). [3] Ferziger J.H. and Peric M., Computational Methods for Fluid Dynamics. Springer, (1996). [4] Hirsch C., Numerical Computation of Internal and External Flows. John Wiley & Sons, Chichester, I & II, (1990). [5] Wesseling P., Principles of Computational Fluid Dynamics. Springer, (2001). [6] Donea J. and Huerta A., Finite Element Methods for Flow Problems. John Wiley & Sons, (2003). [7] Ohner R.L., Applied CFD Techniques: An Introduction Based on Finite Element Methods. John Wiley & Sons, (2001). [8] Cuvelier C., Segal A. and Van Steenhoven A.A., Finite Element Methods and Navier- Stokes Equations. Kluwer, (1986). [9] CFD-Wiki http://www.cfd-online.com/Wiki/ Main Page.
- 37. Thermal design ofliquid cooled charge air cooler: A computational approach 147 Thermal design of liquid cooled charge air cooler: A computational approach TRIBHUWAN CHANDRA JOSHI* and YOGENDRA SINGH KUSHWAH Subros Technical Centre, India *E-mail: tribhuwan.joshi@subros.com Abstract Successive advent of vehicle emissions norms & strong emphasis on fuel efficiency have enforced vehicle manufactures to deploy turbo charging of intake air prior to intake manifold. However introduction of turbochargers also signify need of a heat exchanger called charge air cooler or intercooler to reduce the temperature of turbocharged air to optimum limit prior to intake manifold of engine. The major limitation of conventional air to air charge air coolers is that they yield low heat transfer rate to weight ratio & low heat transfer rate to frontal area ratio characteristics. Beside this at higher boost pressure and higher temperature conditions there are metallurgical challenges for conventional air to air charge air coolers. Therefore; design optimization aimed at improved heat transfer from charge air is the need of strict environmental norms (Euro V onward). Rapid growth of automobile sector coupled with cutting edge technology development is also leading demand for highly efficient as well as compact solutions for charge air cooler. This provide motivation and challenge to design highly optimized new technology based system, which can provide increased rate of heat transfer with lesser pressure drop & overall compact size. To overcome discussed limitations of conventional air to air charge air cooler; the present research explores thermal design of liquid cooled charge air cooling device with unique charge air flow mechanism. The new design of liquid cooled charge air cooler provides multiple flow passes in order to achieve advantages of both cross flow as well as counter flow type heat exchangers. Multiple flow passes basically ensure maximum interaction of two heat exchanging fluids; hence heat exchanger effectiveness gets improved. Charge air is introduced to the heat exchanger with the help of an inlet tank header. The flow of charge air around the heat exchanger is precisely controlled by a casing, which ensures a uniform guided flow to successive passes of heat exchanger. An outlet tank header serves the purpose of supplying charge air to the engine intake manifold. A Computational Fluid Dynamics (CFD) methodology is used to optimize flow dynamics of two working fluids, namely liquid water and charged air coming out of turbocharger. Key words : Automotive, CFD, Heat Transfer, Liquid Cooled Charge Air Cooler, Radiator, Emission, Turbocharger 1. Introduction With the rapid advancement of emission norms, Turbochargers has become the integral part of engine system these days. It not only helps to reduce the pollutant emissions but also improves the fuel efficiency. Turbochargers are used to compress the ambient air to a certain boost pressure prior to its admission Invertis Journal of Renewable Energy, Vol. 7, No. 3, 2017 ; pp. 147-157 into the intake manifold in internal combustion engines. So by compressing the air in turbocharger, large amount of charge air is being inducted into combustion chamber, So as to improve combustion efficiency along with reduced pollutant emission. However along with the compression, it also raises the temperature of air; which in turns results in decrease in density of combustion air. To overcome this limitation a heat exchanger, DOI No. : 10.5958/2454-7611.2017.00021.2
- 38. 148 Tribhuwan Chandra Joshiand Yogendra Singh Kushwah commonly known as Charge Air Cooler is being used between turbo-charger and the intake manifold. Wherein high temperature & high pressure air loses its heat to cooling media (air or water). This basically results in increased density of combustion air & ultimately improved combustion/fuel efficiency of vehicle. Further high boost pressure requirements of intake air to meet latest emission norms (Euro V/ Euro VI) has emphasized the need of highly efficient and durable design of heat exchangers. This will ensure increased heat transfer rate along and less charge air side pressure drop & overall compactness. This requirement has enforced the technology developers to explore liquid based charge air cooling instead of traditional air to air charge air coolers. To achieve the aforementioned objective, some work has been done by researchers in the areas of liquid cooled charge air coolers, Lamich et al. [1] has disclosed a cross flow type liquid charge air cooler, where liquid water is flowing in a U flow manner through equally spaced tube banks, whereas charge air flows through inter tube spacing to make a Charge air flow path. Maceroni et al.[2] has disclosed a Liquid cooled charge air cooler directly integrated with Intake manifold, in this arrangement charge air from inlet tank flows through heat exchanger, where it loses its heat with liquid water flowing through heat exchanger tubes. The majority of the work reported in existing literature suggests either straight or U flow type flow arrangement for liquid water, whereas charge air flows through tube banks in the direction normal to tube bank, to form a Cross flow type heat exchanger. Thus better heat transfer characteristics possible with counter flow type arrangements can't be utilized. Hence there is a scope of flow path optimization, So as to achieve advantages of both cross as well as counter flow type arrangements. Present study explores an innovative charge air flow mechanism, to form a cross- counter flow type arrangement[3]. So as to ensure increased heat transfer rate along with comparable charge air side pressure drop and overall heat exchanger compactness. A CFD methodology to evaluate the various performance parameters of heat exchanger is used. Fin Surfaces associated with the heat exchanger has been modeled as a shell region[4], So as to take into account conduction heat transfer though fins. The CFD modeling strategy for heat exchangers has been correlated with test results of baseline liquid to air heat exchanger (radiator configuration) & it has been found that CFD results are in close agreement with test results. In a different study to develop correlation between liquid water temperature and cooling air Anders et al.[5] have also used CFD modeling strategy for estimating the liquid water temperatures at radiator inlet and found that estimated values were within ± 4°C of the experimental data. 2. Theoretical Background Overall heat transfer rate (Q) for heat exchangers can be expressed by equation (1), (2) & (3). Q = .Cmin. (Tca_in-Tc_in)/3600 (1) Charge Air Side : Qca = mca .Cp_ca . (Tca_in-Tca_out) /3600 (2) Liquid Water Side : Qc = mc .Cp_c . (Tc_in-Tc_out) /3600 (3) In perfect isolated system, heat gained by the cold fluid (liquid) will be equal to heat lost by the hot fluid (charge air). Qc = Qca (4) Where Heat exchanger effectiveness () is defined as the ratio of the actual heat transfer rate (Q) to the maximum possible heat transfer rate (Qmax); and Cmin is minimum of heat capacities of both the fluids: = Q/Qmax (5) Qmax = Cmin. (Tca_in-Tc_in) (6) Cmin =Min. (mca .Cp_ca, mc .Cp_c) (7) In current study a detailed investigation of various flow configurations & their impact on various design parameters has been discussed. Where : WHX : Weight of heat exchanger (kg) VHX : Heat exchanger volume (m3) W* H* D* fp : Width* Height* Depth* fin pitch (all in mm)
- 39. Thermal design ofliquid cooled charge air cooler: A computational approach 149 Q : Heat rejection capacity (kW) Qc : Liquid water (liquid) side heat rejection capacity (kW) Qca : Charge air side heat rejection capacity (kW) : Heat exchanger effectiveness mca : Mass flow rate of charge air (kg/h) Va : Flow velocity of cooling air (m/s) Cp_ca : Specific heat of charge air (kJ/kg. K) Ta_in : Cooling air inlet temperature (ºC) Tca_in : Charge air inlet temperature (ºC) Tca_out : Charge air out temperature (ºC) mc : Mass flow rate of liquid water (kg/min) Cp_c : Specific heat of liquid water (kJ/kg. K) Tc_in : Liquid water inlet temperature (ºC) Tc_out : Liquid water out temperature (ºC) (cs) : Heat exchanger Section BC : Boundary conditions Temp. : Temperature (ºC) 3. CFD Modeling This section presents CFD model considered for the analysis, governing assumptions and validation of CFD results with test data. Modeling of Heat exchanger geometry has been done in commercially available preprocessing software called ANSA (Version 15.1.2). A commercial CFD code called Star CCM V 11.04 has been used as a solver and as a post processer. Volume meshing has been carried out in the same software. For the simplicity of the analysis following assumptions has been considered. Steady state based solver has been considered for both heat exchanging fluids. Within heat exchanger liquid water side as well as charge air side flow is incompressible. Liquid water flow rate through heat exchanger tubes is uniformly distributed Charge air side inlet/outlet headers have been considered as longitudinally extending rectangular ducts. Further size of rectangular duct is defined by available free flow area with various flow configurations. To take into account effect of change in fluid properties, all the properties specified to the CFD code are based on estimated mean temperature values of hot and cold fluid. 3.1 Heat Exchanger Geometry Modeling & Result Deduction Methodology For the simplification of the analysis & to save the overall computational time, a small section (Fig. 1) of heat exchanger has been modeled. This model consists of geometry of two fins separated apart by a tube. Further both the fins are covered by half tube sections on both the sides. Liquid water flows through both single tube & half tube sections & spacing between tubes forms the flow path for air. Inlet mass flow rate for both the working fluids (i.e. liquid water & air) has been calculated based on overall mass flow rate through the complete heat exchanger. Complete flow length (i.e. tube length for the liquid water side & core depth for the air Side) for both the heat exchanging mediums has been modeled, So as to closely estimate the outlet temperatures for both the fluids leaving the heat exchanger. Fig. 1. Tube, fin Heat Exchanger cut Section for CFD modeling A lot of complexity involved with design of charge air side inlet-outlet headers; which depends on specific packaging space & piping layout associated with turbo charger & air intake manifold, Therefore; in current study charge air inlet and outlet headers for various flow configurations mentioned in section 4, has been considered as rectangular ducts extended from available free flow area for a particular flow configuration (D1-D4). This will ensure uniform flow through inlet headers. As complete flow length of both the fluids being considered in CFD model, hence outlet temperatures for both the fluids calculated from CFD analysis can
- 40. 150 Tribhuwan Chandra Joshiand Yogendra Singh Kushwah be used for calculating the various performance parameters of complete heat exchanger. 3.2 Solver Numeric & Boundary Condition's Governing momentum and energy equations are Reynolds-averaged Navier-Stokes equations and realizable k- turbulence model with all Y+ wall treatment being used for modeling turbulent nature of flow. All the heat exchanger surfaces, where modeling of wall thickness & subsequently mesh generation was possible have been modeled as a solid. However, in case of fins, where modeling of thickness is not possible; all the fin walls has been modeled as a zero thickness surfaces. Further to take into account heat conduction through fins, three dimensional Shell feature of Star CCM software has been used. Following type of BC specification method has been used for the analysis. Inlet boundaries for both the fluids have been defined as Mass Flow inlets with associated fluid temperatures. However outlet boundaries have been defined as Pressure outlets. 3.3 Validation of CFD Approach & Results To validate the CFD modeling strategy, an available reference heat exchanger also called as radiator having certain tube fin configuration has been first tested in calorimeter lab (Fig. 2). CFD results have been extracted for a range of liquid water flow rates with fixed liquid water in temperature. However, cooling air side parameters at the inlet have been kept fixed for all the cases. Typical specifications & Boundary conditions for reference heat exchanger have been specified in Table 1. Fig. 3 shows temperature distribution at the centre plane of the heat exchanger obtained from CFD simulations. From temperature profile it is quite evident that high temperature water is flowing through tubes, whereas cooling air is allowed to flow through inter tube spacing & external fins. Table 1. BC & Corce size specification for radiator sample Radiator Core Air Liquid Flow Liquid Air Velocity (100% Water) Inlet Temp. Inlet Temp. Width : 295 mm; Height : 202.6 mm; Depth : 16 mm; Fin Pitch :2.4 mm 4.5 m/s (40 - 60) kg/min 85 °C 20 °C Fig. 2. Component Calorimeter test setup for radiator testing Fig. 3. Temperature distribution across the HX obtained from CFD The comparisons of CFD results and calorimeter lab have been presented in Fig. 4 & 5. From Fig. 4, it is evident that liquid water out temperature obtained from CFD are in close agreement with calorimeter results. Further liquid water side heat rejection rate (Fig. 5) calculated from CFD results are closely matching with test results with an accuracy of ± 5 %.
- 41. Thermal design ofliquid cooled charge air cooler: A computational approach 151 Fig. 4. Liquid water out Temp Vs Liquid water mass flow rate Fig. 6. Single Pass Cross Flow Heat Exchanger Fig. 5. Heat Rejection Capacity Vs Liquid water mass flow rate 4. Investigations of various flow configurations of liquid cooled charge air cooler Based on physics several flow configurations are possible in liquid cooled charge air coolers. To identify the optimum flow configuration following four flow configurations (D1 to D4) has been studied. A "flat-tube & fin" type heat exchanger having core Size of Width 295mm*Height 202.6mm*depth 16mm*fin pitch 2.4 mm, is being considered for all the flow configurations. Further In case of configuration D2, two heat exchanger core's are spaced at a gap of 4 mm. 4.1 Flow Configuration 1 (D1) Fig. 6 shows the flow configuration for single pass cross flow heat exchanger (D1). In this arrangement charge air (fluid 1) enters normal to tube bank, whereas liquid water (fluid 2) is flowing through tubes. 4.2 Flow Configuration 2 (D2) Fig. 7 shows the flow configuration for two banks cross flow heat exchanger (D2). In this arrangement charge air (fluid 1) enters normal to front tube bank and exits through rear tube bank, whereas liquid water (fluid 2) is allowed to flow from rear tube bank first and then it is transferred to front tube bank via intermediate headers. Fig. 7. Two Bank Cross Flow Heat Exchanger 4.3 Flow Configuration 3 (D3) Fig. 8 shows the flow configuration for two pass cross-counter flow heat exchanger (D3). In this arrangement charge air (fluid 1) is being circulated around the tube bank in two numbers of passes. Charge air is first allowed to flow through 50% of
- 42. 152 Tribhuwan Chandra Joshiand Yogendra Singh Kushwah total available core area. It is then guided to second pass via outer casing to form a two pass flow arrangement. The second fluid i.e. liquid water flows in tubes. The direction of flow of both the fluids is such that it forms cross-counter flow arrangement. 5. Results and Discussion All the flow configurations discussed in section 4, has been simulated for a fixed charge air inlet temperature (Tca_in) of 200 °C, liquid water in temperature (Tc_in) of 65 °C & liquid water mass flow rate of 40 kg/min. However charge air mass flow rate has been varied from 200 to 370 kg/h. Further for simplicity of the analysis liquid water is considered as 100% water. Charge air side velocity stream lines through all the flow configurations have been presented. In case of flow configuration 1 & 2 (Fig. 10 & 11), there are no recirculation zones hence lead to highly uniform charge air flow through heat exchanger surfaces. Further in case of design configuration 3 & 4 (Fig 12 & 13), charge air flow path has been optimized in such a way that, it ensures less recirculation zones, Fig. 9. Three Pass Cross- counter Flow Heat Exchanger Fig. 8. Two Pass Cross- counter Flow Heat Exchanger 4.4 Flow Configuration 4 (D4) Fig. 9 shows the flow configuration for three pass cross-counter flow heat exchanger (D4). In this fluid 1 i.e. charge air is being circulated around the tube bank in three passes. Charge air is first allowed to flow through one third of total available core area. It is then guided to successive passes to form a three pass flow arrangement. The second fluid i.e. liquid water flows in tubes. The direction of flow of both the fluids is such that it forms cross- counter flow arrangement. Fig. 10. Velocity stream lines through centre plane of flow configuration 1, (D1), Charge air mass flow rate: 250 kg/h
- 43. Thermal design ofliquid cooled charge air cooler: A computational approach 153 Fig. 12. Velocity stream lines through centre plane of flow configuration 3, (D3), Charge air mass flow rate: 250 kg/h) Fig. 11. Velocity stream lines through centre plane of flow configuration 2, (D2), Charge air mass flow rate: 250 kg/h hence improved flow uniformity through individual passes. 5.1 Liquid water Outlet Temperature Liquid water out temperatures has been plotted for all the design configurations. Fig. 14 shows variation of liquid water out temperature with charge air mass flow rates. Further at low charge air side mass flow rate (~200 kg/h) liquid water out temperature is less for all the flow configurations.
- 44. 154 Tribhuwan Chandra Joshiand Yogendra Singh Kushwah Fig. 13. Velocity stream lines through centre plane of flow configuration 4, (D4), Charge air mass flow rate: 250 kg/h) Fig. 14. Liquid water out temperature over a range of charge air mass flow rates for all the design configurations. Fig. 15 shows charge air out temperature variation with charge air mass flow rates. It is evident from the Fig. 15 that, design configuration D2 & D4 results into lowest charge air out temperatures. The maximum difference between However at high mass flow rates, D2 & D4 have higher liquid water out temperatures, hence better effectiveness. 5.2 Charge Air Outlet Temperature Liquid water out temperatures has been plotted Fig. 15. Charge air out temperature over a range of charge air mass flow rates
- 45. Thermal design ofliquid cooled charge air cooler: A computational approach 155 Fig. 16. Charge air side heat rejection capacity over a range of charge air mass flow rates Fig. 17. Normalized heat rejection capacity over heat exchanger weight outlet temperatures between D2 & D4 configurations is ~ 2 °C at a charge air mass flow rate of 370 kg/h. However for other two configurations the difference of outlet temperatures with respect to D4 is in the range of 5 to 10 °C. This shows that D2 & D4 have better effectiveness. 5.3 Charge Air Heat Rejection Capacity Fig. 16 shows the variation of charge air heat rejection capacity with mass flow rate. It is quite evident that both D2 & D4 has the excellent heat exchange capabilities. Further heat rejection capacity of Design configuration 2 (D2) is slightly higher than D4. This can be attributed to the modeling strategy for inlet & outlet headers as discussed in section 3.1. However practically inlet & outlet headers will be coupled with piping's which associated with turbo charger & air intake manifold respectively. Therefore; it will impose certain degree of flow non uniformity to available heat exchanger area. More the inlet free flow area for charge air, higher will be the non uniformity. In case of D2 flow uniformity will be higher as entire frontal area of heat exchanger will become part of inlet header. Hence will lead to reduced flow uniformity & heat transfer rate. Whereas in case of D4 only one third of frontal area of heat exchanger is available as an inlet header, hence will lead to relatively improved flow uniformity & heat transfer rate. 5.4 Normalization of Heat Rejection Capacity over Heat Exchanger Weight and Volume Normalization of heat exchange capacity over heat exchanger weight & volume has been done. It is quite evident from Fig. 17 & 18 that, design con- figuration 4 (D4) exhibits excellent Qca/WHX & Qca/ VHX characteristics. Whereas design configuration 2 (D2) has the lowest Qca/WHX & Qca/VHX values. 5.5 Qualitative Evaluation of various Design con- figurations In line with the above discussion, various aspects involved with heat exchanger design have been summarized in Table 2. All the flow configurations have been rated relatively from 1 to 4 for listed key heat exchanger parameters. It is evident from Table 2, that flow configuration D4 have highest ratings due to its low weight, high heat rejection capacity & low liquid water side pressure drop. Further from vehicle integration point of view also, it will be most feasible
- 46. 156 Tribhuwan Chandra Joshiand Yogendra Singh Kushwah Table 2. Qualitative evaluation of various design configuration Fig. 18. Normalized heat rejection capacity over heat exchanger solution, due to available scope of compact design of inlet & outlet headers. However charge side pressure drop rating is lower in comparison of D1 & D2. Hence charge air flow casings and inlet/ outlet headers will be the critical aspects for vehicle. Further design configuration D2 has received the lowest rating despite of having very good heat rejection capability. This is primarily due to increased heat exchanger weight along with increased liquid water side pressure drop & overall cost. 6. Conclusions Current research explores a CFD modeling approach for heat exchanger, particularly for "flat tube- fin" type heat exchanger evaluation. The CFD results were first validated with test data obtained for a radiator sample. The comparison shows that liquid water out temperatures obtained from CFD analysis are closely matching with test data. Further heat rejection rate predicted from CFD is within ± 5 % of test data. The main focus of present research was to design and opti-mized flow configuration associated with liquid cooled charge air cooler. A detailed analysis of existing flow configurations (D1 & D2) and proposed flow configurations (D3 & D4) has been carried out using validated CFD approach. The comparison of Qca/WHX & Qca/VHX indicates that design configuration D3 & D4 has the highest heat rejection per unit weight, and heat rejection per unit volume characteristics at specified values of charge air side mass flow rates. Based on various performance parameters, flow configuration D4 is rated as optimized flow configuration considering weight and vehicle integration aspects. Future work will involve integrating liquid cooled charge air (configuration D4) to a vehicle to evaluate impact on vehicle level parameters e.g. fuel efficiency and exhaust emissions. 7. Acknowledgment The authors would like to acknowledge prototype and testing teams of Subros Technical Centre for extending their valuable support. The authors would also like to thank Mr. A. K. Jaiswal
- 47. Thermal design ofliquid cooled charge air cooler: A computational approach 157 (SVP, Subros Technical Centre) for supporting this research work of new technology development. References [1] U.S. Patent Number 7,793,710 B2, Intercooler Apparatus and Method, Bernhard Lamich & Denis Bazika dated Sep. 14, 2010 [2] U.S. Patent Number 2013/0220289 A1, Intake System with An Integrated Charge Air Cooler, Karen Elizabeth dated Aug., 2013 [3] Kays W.M. and London A.L., Compact Heat Exchangers, Third Edition. [4] Star CCM+ User Guide, Version 10.04.009 [5] Anders Jerhamre and Anders Jonson, Development and Validation of Liquid water Temperatures and Cooling Air Flow CFD Simulation at Volvo Cars, SAE Technical paper 2004-01-0051
- 48. 158 Jitendra Singh andShahbaz Mozammil Utilisation of exhaust gases of automobile JITENDRA SINGH1 and SHAHBAZ MOZAMMIL2 1 Department of Mechanical Engg. Faculty of Engg. & Technology, Jamia Millia Islamia, New Delhi-110025, India 2 Department of Mechanical Engg. , Al-Falah School of Engineering & Technology, Haryana-121004, India *E-mail: singharyajeet@gmail.com Abstract Recent trend about the best ways of using the deployable sources of energy in to useful work in order to reduce the rate of consumption of fossil fuel as well as pollution. Out of all the available sources, the internal combustion engines are the major consumer of fossil fuel around the globe. Out of the total heat supplied to the engine in the form of fuel, approximately, 30 to 40% is converted into useful mechanical work. The remaining heat is expelled to the environment through exhaust gases and engine cooling systems. Therefore we have to utilise this energy. This paper includes Utilising the waste energy by using the vapour absorption system and exhaust gas recirculation system. Vapour absorption system required low grade energy for operation. Hence in such types of systems, a physicochemical process replaces the mechanical process of the vapour compression refrigeration system by using energy in the form of heat rather than mechanical work. This heat obtained from the exhaust of high power internal combustion engines. And also using the Exhaust Gas Recirculation NOx is to be control. The exhaust gases mainly consist of carbon dioxide, nitrogen etc. and the mixture has higher specific heat compared to atmospheric air. Re-circulated exhaust gas displaces fresh air entering the combustion chamber with carbon dioxide and water vapour present in engine exhaust. As a consequence of this air displacement, lower amount of oxygen in the intake mixture is available for combustion. Key words : Waste heat, EGR system, VARS , Refrigerants, NOx, pollution. 1. Introduction The vapour absorption refrigeration system is one of the oldest methods of producing refrigerating effect. The principle of vapour absorption was first discovered by Michael Faraday in 1824 while performing a set of experiments to liquefy certain gases. The first machine was based on Vapour Absorption Refrigeration machine was developed by a French scientist, Ferdinand Carry, in1860. This system may be used in both the domestic and large industrial refrigerating plants. refrigerant commonly used in vapour absorption system is Ammonia. This system uses Heat energy instead of mechanical energy as in vapour compression system, in order to change the condition of refrigerant required for the operation of the refrigeration cycle. In automobile air conditioning load factors are constantly and rapidly changing as the automobile moves over highways at Invertis Journal of Renewable Energy, Vol. 7, No. 3, 2017 ; pp. 158-164 different speeds and through different kind of surroundings. As the car moves faster there is greater amount of infiltration into the car and the heat transfer between the outdoor air and the car surface is increased. The sun baking down on a black top road will raise its temperature to 350°C - 450°C approximately and thus increases the amount of heat transmitted into car. Therefore, the car is subjected to varying amounts of heat load when its orientation changes during the journey. An automobile engine utilizes only about 35% of available energy and rests are lost in the form of heat and mechanical losses to cooling and exhaust system. If one is adding conventional air conditioning system to automobile, it further utilizes about 4-5% of the total energy. Therefore automobile becomes costlier, uneconomical and less efficient. The conventional air conditioning system in car decreases the life of engine and increases DOI No. : 10.5958/2454-7611.2017.00022.4
- 49. Utilisation of exhaustgases of automobile 159 the fuel consumption, further for small cars compressor needs 3 to 4 bhp i.e. a significant ratio of the power output. Keeping these problems in mind, a car air conditioning system is proposed which is using exhaust heat. The advantages of this system over conventional air-conditioning system are that it does not affect designed efficiency life and fuel consumption of engine. G VICATOS, J GRYZAGORIDIS and S WANG[1] proposed a new system, Vapour Absorption Refrigeration of System based on Exhaust gas from an Internal Combustion Engine. Here energy from the Exhaust gas of an Internal Combustion Engine is used to power absorption Refrigeration system to Air condition an Ordinary passenger. In this theoretical Design is verified by the unit that is tested under both laboratory and road-test conditions, by using Refrigerant 134a and absorbent Dimethyl formamide (DMF) the cooling effect for a passenger in a vehicle is been very comfort. Dimethylformamide is a clear liquid, miscible with water and majority organic solvents and it is used as a solvent in the production of fibers, films, and surface coatings. Recent trend about the best ways of using the deployable sources of energy in to useful work in order to reduce the rate of consumption of fossil fuel as well as pollution. Out of all the available sources, the internal combustion engines are the major consumer of fossil fuel around the globe. Out of the total heat supplied to the engine in the form of fuel, approximately, 30 to 40% is converted into useful mechanical work. The remaining heat is expelled to the environment through exhaust gases and engine cooling systems, resulting in to entropy rise and serious environmental pollution, so it is required to utilized waste heat into useful work. The recovery and utilization of waste heat not only conserves fuel, usually fossil fuel but also reduces the amount of waste heat and greenhouse gases damped to environment. It is imperative that serious and concrete effort should be launched for conserving this energy through exhaust heat recovery techniques[2]. Such a waste heat recovery would ultimately reduce the overall energy requirement and also the impact on global warming. The Internal Combustion Engine has been a primary power source for automobiles and automotives over the past century. Presently, high fuel costs and concerns about foreign oil dependence have resulted in increasingly complex engine designs to decrease fuel consumption. For example, engine manufacturers have implemented techniques such as enhanced fuel-air mixing, turbo-charging, and variable valve timing in order to increase thermal efficiency. However, around 60-70% of the fuel energy is still lost as waste heat through the coolant or the exhaust. Moreover, increasingly stringent emissions regulations are causing engine manu- facturers to limit combustion temperatures and pressures lowering potential efficiency gains. As the most widely used source of primary power for machinery critical to the transportation, construction and agricultural sectors, engine has consumed more than 60% of fossil oil. On the other hand, legislation of exhaust emission levels has focused on carbon monoxide (CO), hydrocarbons (HC), nitrogen oxides (NOx), and particulate matter (PM). Energy conservation on engine is one of best ways to deal with these problems since it can improve the energy utilization efficiency of engine and reduces emissions. Given the importance of increasing energy conversion efficiency for reducing both the fuel consumption and emissions of engine, scientists and engineers have done lots of successful research aimed to improve engine thermal efficiency, including supercharge, lean mixture combustion, etc. However, in all the energy saving technologies studied. Engine exhaust heat recovery is considered to be one of the most effective. Many researchers recognize that Waste Heat Recovery from engine exhaust has the potential to decrease fuel consumption without increasing emissions, and recent technological advancements have made these systems viable and cost effective. This paper gives a comprehensive review of the waste heat from internal combustion engine, waste heat recovery system and methods of waste heat recovery system. Better fuel economy and higher power with lower maintenance cost has increased the popularity of diesel engine vehicles. Diesel engines are used for bulk movement of goods, powering Stationary/ mobile equipment, and to generate electricity more economically than any other device in this size range. In most of the global car markets, record diesel car sales have been observed in recent years. The exhorting anticipation of additional improvements in
- 50. 160 Jitendra Singh andShahbaz Mozammil diesel fuel and diesel vehicle sales in future have forced diesel engine manufacturers to upgrade the technology in terms of power, fuel economy andemissions[3]. 2. Vapour AbsorptionSystem An absorption refrigerator is a refrigerator that uses a heat source (e.g., solar, kerosene-fuelled flame, waste heat from factories or district heating systems) to provide the energy needed to drive the cooling system[4]. In the early years of the twentieth century, the vapour absorption cycle using water-ammonia systems was popular and widely used. Ammonia- water combination possesses most of the desirable qualities which are listed below: 1 m3 of water absorbs 800 m3 of ammonia (NH3). Latent heat of ammonia at-15°C = 1314 kJ/ kg. Critical temperature of NH3 =132.6°C. Boiling point at atmospheric pressure = –33.3°C The NH3-H2O system requires generator temperatures in the range of 125°C to 170°C with air cooled absorber and condenser and 80°C to 120°C when water-cooling is used. These temperatures cannot be obtained with flat-plate collectors. The coefficient of performance (COP), which is defined as the ratio of the cooling effect to the heat input, which vary between 0.6 to 0.7. Ammonia is highly soluble in water. 3. Refrigerant used for the absorption refrigera- tion systems The properties of the refrigerant combination are that, in liquid phase, they must have a margin of miscibility within the operating temperature range of the cycle. The mixture should also be chemically stable, non-toxic, and non-explosive. In addition to these requirements, the following aredesirable[5]. a. Refrigerant should have high heat of vaporization and high concentration within the absorbent in order to maintain low circulation rate between the generator and the absorber per unit of the coolingcapacity. b. Transport properties that influence heat and mass transfer, e.g., viscosity, thermal conductivity, and diffusion coefficient should befavorable. c. Both refrigerant and absorbent should be non- corrosive, environmental friendly, and economical. There are some 40 refrigerant compounds and 200 absorbent compounds available. However, the most common working fluids are water/ammonia and LiBr/water. Since the invention of absorption refrigeration systems, water/ammonia has been widely used for both cooling and heatingpurposes. The main properties are : Ammonia is the only refrigerant from inorganic group which was used universally for many applications and still used to great extent at the present time. It possesses many properties required for an ideal refrigerant. It has wide applications because of its low volumetric displacement, low cost, low weight of liquid refrigerant per ton of refrigeration and high efficiency. Presently it is widely used in cold storages, ice manufacturing plants and skating rinks due to its low production and maintenance cost[6]. Properties of Ammonia 1. Its toxic, flammable, irritating and fooddestroying 2. Anhydrous ammonia has no effect on lubricating oil but in presents of moisture, ammonia forms an emulsion with oil thatFig. 1. Vapour Absorption System
- 51. Utilisation of exhaustgases of automobile 161 causes operating difficulties. The formed emulsion becomes effective when the percentage of water exceeds 0.01%. It is not oil miscible therefore it will not dilute the oil in the crank case of thecompressor. 3. It is highly volatile and becomes explosive when mixed with air and compressed therefore air leaks must be avoided in ammonia refrigerationsystems. 4. Ammonia attacks on non-ferrous metals in the presents of water therefore copper and brass are never used with ammonia refrigerationsystem. 5. Ammonia can be used economically for -70°C evaporator temperature and its application for further low temperature becomes highly uneconomical and difficult to maintain low vacuum required in the evaporator. 6. Ammonia (refrigerant) and water (absorbent) are highly stable for a wide range of operating temperature andpressure. 7. Ammonia has a high latent heat of vaporization, which is necessary for efficient performance of the system. Its latent heat of vaporization at -15 ºC is 1315 kJ/Kg. 8. Its boiling point at atmospheric pressure is - 33.3 ºC & freezing point is -77ºC 9. It has highest refrigerating effect per Kg of refrigerant. The leakage of this refrigerant may be quickly & easily detected by the use of burning sulphur candle which in the presence of ammonia will form white fumes of ammonium sulphite. It is environmental friendly[7]. 4. Exhaust gas recirculation system When EGR system is applied, the engine intake consists of fresh air and recycled exhaust gas. Exhaust gases were tapped from exhaust pipe and connected to inlet airflow passage. An EGR control valve was provided in this pipe for EGR control. The exhaust gases were regulated by this valve and directly send to the inlet manifold, upstream of compressor or downstream of compressor. Sufficient distance for through mixing of fresh air and exhaust gases were ensured. The above shown system is also called as hot EGR because it not fitted with an EGR cooler which is used to cool the intake mixture. A typical EGR valve is shown in Fig 2. 4.1 Types of EGR systems Low pressure EGR Systems : If portion of turbine outlet exhaust gas is delivered to compressor inlet through the flow control valve then it called low Fig. 2. EGR System in a Turbocharged Engine Fig. 3. Low Pressure EGR System
- 52. 162 Jitendra Singh andShahbaz Mozammil of exhaust gas. This type of loop can be used by directing exhaust from the turbine outlet to the inter- cooler outlet directly bypassing thecompressor. High pressure EGR systems : Another method of EGR is high pressure EGR loop. In high pressure EGR system, a flow passage is devised between the exhaust of engine (up-stream of theturbine) and the intake manifolds of engine (downstream of the super charging compressor). In this system the exhaust gas is recirculated from upstream of the turbine to downstream of the compressor or the downstream of the inter-cooler as shown in Figure. The compressor and intercooler are therefore not exposed to the exhaust gas. Such high pressure loop EGR system is only applicable when the turbine upstream pressure is sufficiently higher than the boost pressure (Compressor downstream pressure) i.e. if (P3-P2)>0. 5. Possibility of heat recovery and availability from I.C. engine Waste heat is heat, which is generated in a process by way of fuel combustion or chemical reaction, and then "dumped" into the environment even though it could still be reused for some useful and economic purpose. This heat depends in part on the tempera- ture of the waste heat gases and mass flow rate of exhaust gas. Waste heat losses arise both from equipment inefficiencies and from thermodynamic limitations on equipment and processes. For example, consider internal combustion engine approximately 30 to 40% is converted into useful mechanical work. The remaining heat is expelled to the environment through exhaust gases and engine cooling systems. It means approximately 60 to 70% energy losses as a waste heat through exhaust (30% as engine cooling System and 30 to 40% as environment through exhaustgas. Exhaust gases immediately leaving the engine can have temperatures as high as 842-1112°F [450- 600°C]. Consequently, these gases have high heat content, carrying away as exhaust emission. Efforts can be made to design more energy efficient reverberatory engine with better heat transfer and lower exhaust temperatures; however, the laws of thermodynamics place a lower limit on the pressure EGR loop. In low pressure EGR system, a flow passage is a devised between the exhaust of super charger turbine and the intake manifolds connected to the super charging compressor. The flow of EGR regulated is with a throttling valve showing in Figure above. The pressure differences generally are sufficient to drive the EGR flow of a desired amount except during idling. If the exhaust gas is recycled to the intake directly, the operation is called hot EGR. If the exhaust gas is recycled through EGR cooler, the operation is called cooledEGR[8]. For turbocharged I.C. engine modification is done in EGR system because a positive differential pressure between the turbine outlet and compressor inlet is generally available (p4-p1)>0. Furthermore, tailpipe pressure p1 can be elevated by partial throttling that ensures sufficient driving pressure for the EGR flow. The low pressure EGR loop is not applicable as the conventional compressor and inter- coolers are not designed to ensure the temperature Fig. 4. High Pressure EGR System
- 53. Utilisation of exhaustgases of automobile 163 temperature of exhaust gases. Fig. 5 show total energy distributions from internal combustionengine. 6. Result and Discussion 6.1 Benefits of 'waste heat recovery' canbe broadly classified in twocategories: Direct Benefits :Recovery of waste heat has a direct effect on the Combustion process efficiency. Thisis reflected by reduction in the utility consumption and process cost. Indirect Benefits: Reduction in pollution: A number of toxic combustible wastes such as carbon monoxide (CO), hydrocarbons (HC), nitrogen oxides (NOx), and particulate matter (PM) etc, releasing to atmosphere. Recovering of heat reduces the environmental pollutionlevels. Reduction in equipment sizes: Waste heat recovery reduces the fuel consumption, which leads to reduction in the flue gas produced. This results in reduction in equipment sizes. Reduction in auxiliary energy consumption: Reduction in equipment sizes gives additional benefits in the form of reduction in auxiliary energyconsumption.(9) In automobile engines significant amount of heat is released to the environment. For example, As much as 35% of the thermal energy generated from combustion in an automotive engine is lost to the environment through exhaust gas and other losses. The amount of such loss, recoverable at least partly or greatly depends on the engine load. Among various advanced concepts, Exhaust Energy Recovery for Internal Combustion (IC) engines has been proved to not just bring measurable advantages for improving fuel consumption but also increase engine power output (power density) or downsizing, further reducing CO2 and other harmful exhaust emissions correspondingly. Which was predicted that if 6% of the heat contained in the exhaust gases were converted to electric power, 10% reduction of fuel consumption can beachieved. 7. Conclusions The possibility to design a refrigeration unit inside an automobile using the waste heat from the engine of the vehicle based on Vapor Absorption Refrigeration System is realistic. Also keeping in mind the Environmental safety view, this system is Eco- friendly as it involves the use of Ammonia (a natural gas) as a refrigerant and is not responsible for Green House effect and OZONE layer depletion. In this way we can conclude, that out of the total heat supplied to the engine in the form of fuel combustion, approximately, 35% to 40% is converted into useful mechanical work; the remaining heat is categorized under the waste heat and expelled out of the system, resulting in the rise of entropy, so it is required to utilize this waste heat into useful work. Possible methods to recover the waste heat from internal combustion engine through the study on the performance and emissions of the internal combustion engine are discussed upon and can be designed. Waste heat recovery system is the best way to recover waste heat and saving thefuel[10]. References [1] Endo T., Kawajiri S., Kojima Y., Takahashi K., Baba T., Ibaraki S. and Takahashi T., "Study on Maximizing Exergy in Automotive Engines," SAE Int. Publication, 2007-01-0257, 2007. [2] Nantha Gopal K., Rayapati Subbarao, Pandiyarajan V. and Velraj R., "Thermo- dynamic analysis of a diesel engine integrated with a PCM based energy storage system," International Journal of Thermodynamics 13(1), (2010) 15-21. Fig. 5. Fuel Energy Content in I. C. Engine
- 54. 164 Jitendra Singh andShahbaz Mozammil [3] Hakan Özcan and Söylemez M.S., "Thermal balance of a LPG fuelled, four stroke SI engine with water addition," Energy Conversion and Management, 47(5), (2006) 570-581. [4] Sathiamurthi P., "Design and Development of Waste Heat Recovery System for air Conditioning," Unit European Journal of Scientific Research, 54(1), (2011) 102-110. [5] YadavR.J. and Verma R.S., "Effective Utilization of Waste Heat from Diesel Genset to Run Air Conditioning Plant", Advances in Energy Research, (2006). [6] Shah Alam, A proposed model for utilizing exhaust heat to run automobile air conditioner, joint international conference on "sustainable energy and environment (SEE 2006)" 21-23 Nov(2006), E-011(P). [7] Florides G.A., Kalogirou S.A., Tassou S.A.and Wrobel L.C., Design and construction of a LiBr- water absorption machine, Energy Conversion and Management, 44, (2003) 2483-2508. [8] Verband der Automobilindustrie, VDA Jahresbericht (2009). [9] Dietsche K.-H. and Klingebiel M., Bosch Automotive Handbook, 7th edition, (2007). [10] Balaji K. and Senthil Kumar R., Study of vapor absorption system using waste heat in sugar industry, IOSR Journal of Engineering, (2012), 2250-3021.
- 55. Life cycle analysisof multi-split Variable Refrigerant Flow (VRF) system : case study 165 Life cycle analysis of multi-split Variable Refrigerant Flow (VRF) system : case study MD. KHURSHID ALAM1 and SHAH ALAM2 1 Al-Falah University, Dhauj, Faridabad, Haryana 2 Jamia Millia Islamia, New Delhi *E-mail: sm.arzoo111@gmail.com Abstract In this paper a building having ground and first floor located in A.M.U. Aligarh has been selected for HVAC purposes. The total area of building is 18327 square feet. The total cooling load of ground floor and first floor is 149.5 TR. The total cooling load is ground floor is 89 TR and first floor is 61 TR. The total HP outdoor units in VRF system is installed load for the diversity of VRF units is 80 - 90%. The load is calculated on the basis of occupants, electric loads, exposed area to sun etc. We have found that total HP outdoor unit and TR indoor units are needed. The total consumption of electricity for running VRF units is also calculated. It is found that units electricity is consumed 3,30,000 units/year for VRF System and 4,22,500 units/year for split system. The life cycle analysis has been also done for useful life of 10 years for split units and 12 years for VRF units.The aim is to create thermally controlled environment within the space of a building envelope by designing and planning a HVAC system for the project with the objective that the system designed and built is cost-wise economical, energy efficient as well as simple, flexible with regard to its operation, maintenance. This document mentions the codes, standards and criteria that will generally be used in the design and constructions of HVAC system for this project. Key words : HVAC System, VRF System, Life Cycle Analysis. 1. Introduction Heating, ventilation, and air conditioning (HVAC) is the technology of indoor and vehicular environmental comfort. Its goal is to provide thermal comfort and acceptable indoor air quality. HVAC system design is a sub discipline of mechanical engineering, based on the principles of thermo- dynamics, fluid mechanics, and transfer. Refrigeration is sometimes added to the field's abbreviation as HVAC & R or HVACR, or ventilating is dropped as in HACR (such as the designation of HACR-rated circuit breakers). The primary function of all air-conditioning systems is to provide thermal comfort for building occupants. There are a wide range of air conditioning systems available, staring from the basic window- fitted unit to the small split systems, medium scale Invertis Journal of Renewable Energy, Vol. 7, No. 3, 2017 ; pp. 165-170 package units, large chilled water systems and very latest variable refrigerant flow (VRF) system. Split type air conditioning systems are one to one system consisting of one evaporator (fan coil) unit connected to an external condensing unit. Both the indoor and outdoor unit are connected through copper tubing and electrical cabling. The indoor part (evaporator) pulls heat out from the surrounding air while the outdoor condensing unit transfers the heat into the environment. The advantages of split systems are : Low initial cost, less noise and ease of installation; Good alternative to ducted systems; Each system is totally independent and has its own control. DOI No. : 10.5958/2454-7611.2017.00023.6
- 56. 166 Md. Khurshid Alamand Shah Alam Apart from this there are several disadvantages of this system given as : There is limitation on the distance between the indoor and outdoor unit i.e. refrigerant piping can't exceed the limits stipulated by the manufacturer (usually 100 to 150 ft) otherwise the performance will suffer; Maintenance (cleaning/change of filters) is within the occupied space; Limited air throw, which can lead to possible hot/cold spots; Impact on building aesthetics of large building because too many outdoor units will spoil the appearance of the building. VRF System : The term VRF refers to the ability of the system to control the amount of refrigerant flowing to each of the evaporators, enabling the use of many evaporators of differing capacities and configurations, individualized comfort control, simultaneous heating and cooling in different zones, and heat recovery from one zone to another. VRF systems operate on the direct expansion (DX) principle meaning that heat is transferred to or from the space directly by circulating refrigerant to evaporators located near or within the conditioned space. Refrigerant flow control is the key to many advantages as well as the major technical challenge of VRF systems. Note the term VRF systems should not be confused with the centralized VAV (variable air volume) systems, which work by varying the air flow to the conditioned space on variation in room loads. 2. Literature Review Roth et al. (2002)[1] had presented the paper on VRF installed costs are highly dependent on the application, construction and lay out of the building and whether the installation is new or retrofit. Lack of familiarity with the technology in the U.S will add to VRF costs. Total costs of VRF systems are likely to be about 5% to 20% higher than chilled water systems of similar capacity. Hai, Xiaohong (2006 )[2] et al. had discussed an effective method to combine the ice- storage technology with VRV system in commercial central air conditioning system and the system has the merits of flexible controls, simple operation and compact construction, which need not the big space and difficult construction like the big central ice- storage system. William Goetzler (2007)[3] et al. had presented the paper on VRF systems are not suitable for all commercial building applications. However, they are an excellent option for certain projects, and one more tool for engineers to consider. Johnson (2007)[4] et al. had presented the paper on Ozone depletion issues became an increasing concern at that time issues of a high refrigerant charge of multi split system was likely a strong negative for the system. Since that time, refrigerant developments, advances in charge management, control and inverter technology Increasing market acceptance of VRF technology. Morton Blatt, (2008)[5] et al. had presented It observed to VRF systems are enhanced versions of ductless multi-split systems. VRF systems are very popular in Asia and support from major U.S. Main impact of this system on the electric utility, application recommendations and technology attributes. John Rogers, (2008)[6] et al. had presented this paper reviews The India Low Carbon Growth study is developing a bottom-up model that covers the electricity supply, residential, nonresidential buildings, transport, industry, and agricultural sectors. In heating/cooling includes like electric water heater, air cooler, fans, air conditioning. It can be seen that by 2031 the difference amounts to 135,240 GWh/ year, equivalent to the total output of approximately thirty eight 500 MW power stations. Qiu Tu (2010)[7] et al. had presented the paper on the heating control strategy is key technology for stable and reliable heating operation of variable refrigerant flow air conditioning system with multi-module outdoor units. Li and Wu (2010)[8] et al. had presented the paper on the heat recovery VRF system can save up to 17% of the energy consumption compared to a heat pump VRF system set up. This heat recovery VRF system offered an additional advantage that it can provide cooling and heating for different zones at the same time. Tolga N. Aynur (2010)[9] et al. had presented the paper on VRF system to control refrigerant mass flow rate according to cooling and heating load. This detailed review indicates that the researchers focus on three main subjects: (a) Control strategies of the A Review of a HVAC With VRF System (IJIRST/ Volume 1 / Issue 10 / 003) All rights
- 57. Life cycle analysisof multi-split Variable Refrigerant Flow (VRF) system : case study 167 reserved by www.ijirst.org.Jignesh M.B. [10] has designed HVAC System foa a commercial building. 3. Methodology In this analysis we have taken first floor of a building located in A. M. U. Aligarh (U.P.). The area of building is 18327 square feet. First we got the plan of building. With the help of plan we have calculated all dimensions like length width of each wall, partition wall, window, floor area. Then according to the climate of AMU U.P., we select recommended outdoor condition with the help of Weather Data Handbook published by ISHRAE carrier hand book. We get all the specification such as height of walls, glasses and ceiling height from architecture drawing and calculated the area of all walls, partition walls and windows, of conference room which we can get fresh air, Dehumidified air and TR. On the basis of cooling load we have selected air conditioning system. Basis of Design Ambient Summer Conditions : DBT 110°F (43.3°C) WBT 75°F (23.9°C) RH 60% Ambient Monsoon Conditions : DBT 95°F (35.0°C) WBT 83°F (28.3°C) RH 88% Indoor Temperature : 73 ± 2°F in Office and 74 ± 2°F in Lift Lobby & Entrance Hall etc. RH : RH may exceed 65% in peak monsoon. Monsoon Reheat is not considered hence humidity may increase up to 65-70% during peak monsoon. Occupancy : 80 Sq.ft per person in office at chargeable area, chargeable area is 22% more than air conditioned area, 125 Sq.ft per person in entrance lobby, Equipment Load : 3.5 W/sq.ft in Office area Lighting Load : 1.25 W/ft² in Office, 1.25 W/ft² in Lobby Fresh Air : As per ASHRAE standard 62.1- 2007 (5 CFM per person+0.06 cfm/sqft) + 30 % higher to meet LEED requirement for Office area, lobbies Design Requirements It is often a practice to design an Air Conditioning system for peak load conditions. However, the average usage of an Air Conditioning system during a year is between 65% to 75% Hence, the proposed system to be adopted for this project will be planned & selected for Lowest capital cost. Flexibility to shut down in the areas that are not in use Energy efficiency Minimum Outdoor unit The Heat Load of Conference Room is given below : Table 1. Heat load calculation
- 58. 168 Md. Khurshid Alamand Shah Alam
- 59. Life cycle analysisof multi-split Variable Refrigerant Flow (VRF) system : case study 169 The layout of building floor is as given below: Life Cycle Analysis : Fig. 1. HVAC layout of first floor of building (Based on 2500 Hours per year) VRV System Air Cooled Split Units 1. Total Installed Load 120 TR 130 TR 2. Repair & Maintenance cost per Year 3,50,000/- 2,45,000/- 3 Power consumption/Hour (KWH) 132 Units 169 Units (@ 1.1 KW per Ton) (@ 1.3 KW per Ton) 4 Total Power consumption (Units Per Year) 3,30,000 Unit 4,22,500 Unit 5 Cost @ Rs 8/- per Unit in one Year 26,40,000/- 33,80,000/- Total running cost including Maintenance in 1 Years 29,90,000/- 36,25,000/- 9. Difference in running cost/Year -- Approx.6,35,000/- 10 Installation cost Approx. 79.2 Lacs Approx. 52.8 Lacs 11 Difference in Installation cost -- Approx. 26.4 Lacs Pay Back Period (Using VRV System) Around 4.2 Years LIFE CYCLE 12 YEARS 10 YEARS
- 60. 170 Md. Khurshid Alamand Shah Alam Life Cycle Analysis using Annual Cost Analysis method : C= M+E+I / (P/S, i %, n) – S/ (F/S, i %, n). F/S = (1+i)n–1, P/S= (1+i)n–1/ i(1+i)n, Option Installation Maintanance Electric Intrest Year Annual (Lac) (Lac) consumption (Lac) rate (Lac) expenses (Lac) SPLIT 52.8 2.45 33.8 0.085 10 42.61193 VRF 79.2 3.5 26.4 0.085 12 0.638348 4. Conclusions Based on comparative analysis of alternative with operating cost and life cycle cost, a VRF (variable refrigerant flow) system is selected for a particular project of 149.5 TR. Codes and Standards NBC - National Building Code ASHRAE - American Society of Heating, Refrigeration and Air-Conditioning Engineers. ARI - Air-conditioning and Refrigeration Institute NFPA - National Fire Protection Association UL - Underwriters' Laboratories AMCA - Air Movement and Control Association References [1] Roth Kurt, et al. "Energy consumption characteristics of commercial building HVAC system volume III: energy saving potential" TIAX LLC FOR DOE, 2002 www.eere.gov/ b u i l d i n g / i n f o / d o c u m e n t s / p d f s / hvacvolume2finalreport.pdf. [2] Hai Xiaohong, Tao Zhang, Yun Fanhua and Jun Shen, "Design and Research of the Commercial Digital VRV Multi-Connected Units With Sub-Cooled Ice Storage System" (2006). International Refrigeration and Air Conditioning Conference. Paper 759. http:// docs.lib.purdue.edu/iracc/759. [3] William Goetzler, Member ASHRAE, ASHRAE Journal, April 2007. [4] Johnson Spellman" ASHRAE headquarters building renovation, mechanicals narrative" June26, 2007. http;//images.ashrae.biz/ renovation/documents/o6js22_mech_design_ narrative_permit%20set_2.pdf. [5] Ammi Amarnath, "Variable Refrigerant Flow: An Emerging Air Conditioner and Heat Pump Technology" Electric Power Research Institute. Morton Blatt, Energy Utilization Consultant 2008, ACEEE Summer Study on Energy Efficiency in Buildings. [6] John rogers, "Residential Consumption of Electricity in India” Background Paper India: Strategies for Low Carbon Growth, July 2008 The World Bank. [7] Qiu Tua b, Ziping Fenga b, Shoubo Maoc, Kaijun Donga b, Rui Xiaoa b and Wenji Songa, "Heating control strategy for variable refrigerant flow air conditioning system with multi-module outdoor units" Guangzhou Institute of Energy Conversion, Chinese Academy of Science, Guangzhou 510640, China b Key Laboratory of Renewable Energy and Gas Hydrate, Chinese Academy of Science, Guangzhou 510640, China c Haier Air- Conditioning Electronic Co. LTD., Qingdao 266510, China [8] Li Y.M., Wu J.Y. and Shiochi S., "Experimental validation of the simulation module of the water-cooled variable refrigerant flow system under cooling operation", Applied Energy, 87, (2010) 1513-1521. [9] Tolga N. Aynur, "Variable refrigerant flow systems: A review", Center for Environmental Energy Engineering, Department of Mechanical Engineering, University of Maryland, 3157 Glenn Martin Hall Building, College Park, MD 20742, USA.
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