2nd Canadian Solar Buildings Conference
Calgary, June 10 – 14, 2007
ANALYSIS OF HIGH EFFICIENCY
SOLAR AIR HEATER FOR COLD CLIMATES
Ahmed M. Qenawy and A. A. Mohamad
Mechanical and Manufacturing Department, Schulich School of Engineering,
University of Calgary, Calgary, AB, Canada, T2N 1N4
Tel.: +1 (403) 220-5787, Fax: +1 (403) 282-0486, amqenawy@ucalgary.ca
Type of Paper: Refereed
ABSTRACT
Minimizing solar collector heat losses and increasing
the heat transfer coefficient inside the solar collector,
especially in cold climate countries, is crucial. Since
ambient temperature has a significant effect on the
performance of solar air heaters, the effect of ambient
and inlet temperature to the solar air heater on the
performance of solar air heater equipped with porous
material are discussed. An analysis and design of the
solar air heater to be operated in cold climatic regions
is described. A mathematical model developed for
this study. The model accounts for changes of air
thermo-physical properties with temperature. Effect
of different design and operating parameters such as
mass flow rate and ambient temperature on the solar
collector performance are analyzed. The air mass
flow rate is found to have a great impact on the solar
collector performance especially for low ambient
temperatures. Optimum values of air mass flow rate
are suggested to maximize the performance of the
solar collector. Other important design factors such
as collector length and spacing between collector
covers are investigated. . It is shown that a collector
thermal and exergetic efficiency of about 85% and
39%, respectively, could be obtained for ambient
temperature of 20 oC, while the thermal efficiency
will be around 55% and the exergetic efficiency of 40
% for ambient temperature of -10 oC.
INTRODUCTION
Increased energy prices and the continuous reduction
of the Earth’s conventional fuels resources as well as
the increased world-wide global warming have been
the motivation for the recent growing interest in
alternative sources of energy, such as solar energy.
The development of new and renewable energy
technologies is important for the future of a balanced
global energy economy [1].
Solar air systems have some advantages compared
with those requiring liquid heat transfer media.
Freezing and boiling problems are eliminated.
Heated air can be used directly in air heating systems
without the need for external fluid loop. Corrosion
problems are considerably reduced. Solar air heaters
could be used for a wide variety of industrial
applications such as crop drying [2-4], adsorbent
beds regeneration in separation processes [5] and
meeting buildings heating load [6, 7].
Disadvantages of solar air heaters when compared
with solar liquid heaters like a rise from air leakage,
relatively high pumping power and low heat transfer
coefficient.
Several improvements have been suggested in
literature to enhance the performance of the system.
The use of porous material inside the collector is one
method to enhance solar air heater efficiency as
suggested by Lansing [8]. The use of porous media
tends to increase the surface per unit volume ratio
substantially. Porous media usage within solar air
heaters was reported to improve its performance [9].
Different porous materials have been suggested, such
as wire mesh [10-12], solid or hollow spheres [13],
iron foils [14] and crushed glass [15]. The solar
collector thermal performance is enhanced using
porous materials; however the pressure drop is
increased especially at high flow rates. Minimizing
pumping power by different backing arrangements
was discussed by Mittal [16].
It is also suggested to insert an absorbing plate into
double-pass channel in a flat-plate solar air heater
with air recycling. This design is found to improve
the collector efficiency due to the better value of the
heat transfer coefficient [17].
The major heat losses from flat-plate solar collectors
are through the top cover because the sides and the
bottom of the collector can be well insulated. In order
to minimize the heat losses double glazing were
recommended [18]. Recently, it has been suggested
to use transparent insulating materials (TIM) instead
of conventional glass covers. TIM can reduce the
collector top heat losses due to its low thermal
conductivity; however it is characterized also with
lower transmitivity compared with glass. Therefore,
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2nd Canadian Solar Buildings Conference
Calgary, June 10 – 14, 2007
the transmitted part of the solar radiation to the air
inside the solar collector is lower [19].
It is also suggested to insert an absorbing plate into
double-pass channel in a flat-plate solar air heater
with air recycling. This design is found to improve
the collector efficiency due to the better value of the
heat transfer coefficient [9]. The advanced solar air
heater analysis [9] showed that the thermal efficiency
can reaches 80% by combining high permeable
porous layer with double pass. Experimental work of
[20] proved the analysis of [9].
Ambient temperature of cold climate regions may be
down to about -25 oC. As the ambient temperature
decreases, the driving force for heat loss also
increases causing degradations in the thermal
performance of the collectors. In this study the
performance of high efficiency solar air heater
introduced by Mohamad [9] is analyzed for cold
climate conditions. It is aimed to select the
appropriate design parameters to maximize collector
efficiency. Two scenarios are studied, the air drawn
from the building is circulated through the collector
and air directly drawn from the ambient is passed
through the collector. The first configuration
simulates the condition of using solar air heater as a
part of heating system, while the second
configuration, the collector is used as ventilation
auxiliary heater.
ANALYSIS
Following Mohamad’s [10] solar air heater
configuration and analysis, the solar collector
configuration is shown in Fig. 1. The heater has a
two counter flow air passes. Air is forced to flow
through the gap between the two glass covers.
Therefore, energy extracted from the glass covers
will be used to preheat the air stream. A wire mesh
(high permeable) is used as absorber. Wire mesh has
high surface area per unit volume, which enhances
the rate of heat transfer from mesh to the air.
For purpose of analysis the following assumptions
were considered:
-Air is an ideal gas,
-Steady air flow rate,
- Constant porous media thermal conductivity,
- Quasi-steady heat transfer,
- Negligible conduction resistance in glass covers,
Top glass cover:
Energy conservation on the glass cover dictates that:
Iαg = ha ( Tc1 - Ta ) + hfc1 ( Tc1 - Tf 1 )
+hr ,c1c2 ( Tc1 - Tc2 ) + hr ,ca ( Tc1 - Ta )
(
)
⎡
⎛ 2 ⎞⎤
−1⎟⎥
⎝ εc ⎠⎦
hr,c1c2 =σ T c1 +T c2 ⎢(Tc1 +Tc2 ) ⎜
2
2
⎣
⎡
⎛ 1 ⎞⎤
2
2
hr,ca =σ( T c1 +T a ) ⎢(Tc1 +Ta ) ⎜ −1⎟⎥
⎝ εc ⎠⎦
⎣
(1)
(2)
(3)
I is the solar radiation incident on the surface, W/m2.
αg is the glass cover absorptivity. The left hand side
of the Eq. 1 represents the absorbed portion of the
solar radiation in the top glass cover. The right hand
side represents heat transfer from the glass cover.
The first and second terms are the convection heat
transfer from the glass cover to the ambient air and
the first air stream. While, the third and fourth terms
represent the radiation heat loss to the surrounding
sky and the second glass cover respectively.
First Air-pass:
Applying energy conservation principle for the air
passing between the glass-covers yields:
& p,1
mC
dTf 1
dx
= hfc1 (Tc1 − Tf 1 ) + hfc 2 (Tc 2 − Tf 1 )
(4)
& is the mass flow rate of fluid per unit
Where m
width of the solar collector, Cp1 is the specific heat of
the air evaluated at the average temperature of the
first air stream, Tf1 is the air temperature in the first
pass, Tc1 is the first cover temperature, hfc1 is the heat
transfer coefficient between the air in the first stream
and the first cover and hfc2 is the heat transfer
coefficient between the air in the first stream and the
second cover.
Second glass cover:
Energy conservation on the glass cover dictates that:
Iαgτ g = hr ,c1c2 (Tc2 − Tc1 ) + hfc2 (Tc2 − Tf 1 )
+hr ,cp (Tc2 − Tp ) + hfc2b (Tc2 − Tf 2 )
(5)
Where,
(
hr,cp =σ T c2 +T p
2
2
⎡
) ⎢(Tc2 +Tp )
⎣⎢
⎛ 1 1 ⎞⎤
⎜⎜ + −1⎟⎟⎥
⎝ εc εp ⎠⎦⎥
(6)
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2nd Canadian Solar Buildings Conference
Calgary, June 10 – 14, 2007
Where, τg is the glass transmitivity and Tf2 is the air
temperature in the second air passage.
Second air-pass:
Applying energy conservation principle for the air
passing between the second glass-cover gives:
& p,1
mC
dTf 1
dx
= hfc1 (Tc1 − Tf 1 ) + hfc 2 (Tc 2 − Tf 1 )
(7)
EXERGY ANALYSIS
parameters of the solar collector. Integration of Eqs.
(4) and (6) yields the air temperature distribution in
the solar collector as well as the exit air temperature.
Solving this model requires information about the
heat transfer coefficient inside as well as outside the
solar collector.
In this study, the heat transfer coefficient between the
first glass cover and the ambient air is(7)
assumed to be
10 W/m2K [22]. Whereas, the heat transfer
coefficient inside the solar collector between the
flowing air and the heated glass surfaces are
calculated from [23]:
The second law of thermodynamics deals with the
quality of energy rather than its quantity. The
importance of this law application has recently
increased due to the increased awareness of the
world’s limited energy resources. The second law of
thermodynamics can be used as a powerful tool for
optimization of energy conversion systems.
Employing the entropy production concept, sources
of energy degradation within systems could be
analyzed and identified.
For the solar air heater, the exergy gain by the air
passing through the solar collector can be calculated
from:
Where Nu is the Nusselt number, Re is the Reynolds
number, Pr is the Prandtl number and Dh is the
hydraulic diameter of the air passage channel. The
overall heat transfer coefficient between the back of
the collector and the ambient is assumed to 1.0
W/m2K.
The extra pressure drop introduced by this design of
the solar collectoer is expected not to be high as
suggested by [9].
Egain = E f 2 − Ei
RESULTS AND DISCUSSION
(8)
Ef2 is the exergy of the air exiting from the solar
collector. Ei is the exergy of the entering stream.
Considering air as an ideal gas, Eq. 8 can be rewritten as [21]:
⎛T
⎞
& p 2 (Tf 2(9)
& p 2 ln ⎜ f 2 ⎟
Egain = mc
− Ti ) − To mc
T
i
⎝
⎠
(9)
The exergy of the solar radiation incident on the solar
collector can be estimated using the following
equation [21]:
T
Ein = IA ⎛⎜1 − o ⎞⎟
Ts ⎠
⎝
(10)
Where, Ts is the apparent sun temperature which is
approximately set equal to 6000 K.
SOLUTION METHODOLOGY
Equations (1) through (10) give complete model
equations of the thermal performance of the solar air
heater. These equations identify the performance
Nu =
hf Dh
k
= 0.0333Re0.8 Pr1/3
(11)
The following (8)
values are used in solving the
mathematical model. The transmitivity of the glass
cover (τg) and the absorpitivity of the glass cover (αg)
are assumed 0.92 and 0.06, respectively. The
emissivity of the glass cover (εg) is set to 0.92. All
calculations are performed for 2 m long and 1 m
width solar collector. The channel height in both
directions of air flow are set equal to 0.05 m. The
solar radiation intensity of 950 W/m2K is assumed.
The inlet air temperature is otherwise stated is set to
20 oC.
The performance of the solar collector is defined by
first (ηI) and second low (ηII) efficiencies as,
ηI =
Qgain
Qin
(12)
and
ηII =
Egain
Ein
(13)
The results for solving the pre-described
mathematical model are obtained and are shown on
Fig. 2. Figure 2 shows a comparison between the
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2nd Canadian Solar Buildings Conference
Calgary, June 10 – 14, 2007
experimental data [20] with the model prediction.
The agreement between the model prediction (Eq. 17) and experimental data is within experimental
uncertainties. The experimental setup consists of a
solar collector consisted of a rectangular collection
section, as well as a triangular outlet nozzle and fan.
Seven copper constantan type J thermocouples were
used in the collector. Five of them were spaced 20
cm apart down the length of the collector. The
remaining 2 thermocouples were placed 5.1 cm and
7.6 cm respectively, in the centerline of the outlet air
duct. The solar intensity was measured using a small
pyrometer which was fixed into the upper corner of
the insulation attached to the collector, parallel to the
face of the collector [20]. Figure 2 and Figure 3 show
this agreement for two day measurements Mar. 1,
2003 and Mar. 2, 2003 respectively.
For the verified model and for this paper objective, in
the following paragraphs, the effect of ambient air
temperature on the performance of the solar air
heater is studied with constant inlet temperature to
the collector (as a part of heating system). Then, the
effect of ambient temperature on the performance of
the solar air heater is studies with air drawn from
ambient (as a part of ventilation system).
Figure 4 shows the effect of different air mass flow
rate on the performance of the solar collector for
various ambient temperatures, for a fixed inlet
temperature of 20 oC. For building heating the return
air is from the heated spaces are readmitted to the
solar collector at almost fixed temperature. For
ambient temperature of 20 oC, the results show the
thermal efficiency slight increases with increasing
the mass flow rate. However, for lower ambient
temperatures the efficient tends to decrease with
increasing the mass flow rate. This could be reasoned
by recalling the inlet temperature is set to 20oC which
is higher than the ambient temperature, hence the
losses increase by increasing the flow rate. Figure 5
shows that the exergetic efficiency decreases as the
flow rate increase. These results can be explained
that quality of energy decreases as the mass flow rate
increase. In other words, the temperature of air
stream decreases as the flow rate increase.
Figure 6 shows performance cure for solar air heater
for different ambient temperature. It is clear that the
thermal efficiency of advance solar air heater is
superior to the conventional solar air heat for
different ambient temperatures. It is known that the
thermal efficiency of conventional flat plate air
heater is about 35% [23], while Fig. 6 shows
substantially higher value for the solar in the
analysis. The exergetic efficiency of the solar
collector is found to decrease with increasing the
mass flow rate, Fig. 7.
Figure 7 show the effect of the collector length to
width ratio of the performance of the solar collector.
Increasing the collector length which increases the air
residence time inside the solar collector tends to
increase the efficiency of the solar collector.
The same behavior is shown in Fig. 9 which shows
that the exergetic efficiency is improved for longer
solar collectors for all ambient conditions.
Another important parameter is the air gap in the first
pass (H1) and in the second pass (H2). The effect of
different design values of H1 and H2 on the solar
collector performance are shown on Fig. 10. For the
previously mentioned parameters, as shown in Fig.
10, the best arrangement for maximum efficiency is
for H1= 0.025m and H2=0.05m. This arrangement
insures that a higher velocity in the first air passage
than the second air passage. The configuration
suggests that large portion of the heat absorbed by
the first and second glass covered to be transferred to
the air through higher heat transfer coefficient.
Where as in the second passage which has lower air
velocity, the residence time is increased which helps
in getting out most of the energy absorbed by the
porous material.
The performance of the solar collector under typical
weather conditions is shown in Fig. 11. The solar
collector is of 2 m2 area (1 m width by 2 m length).
The air mass flow rate is 0.05 kg/s.m. The inlet
collector temperature is fixed at 20 oC. Typical
weather data is used for Calgary area as an example
of a cold climate city. It could be concluded that the
solar collector could be successfully operated under
the weather conditions [24, 25] shown in Fig. 11. The
maximum temperature that can be obtained from the
solar collector is estimated to be of 51.27 oC.
CONCLUSION
The performance of solar air collector equipped with
porous material is discussed. Efficiency up to 85%
could be obtained from such configuration. The
ambient temperature has a significant effect on the
performance of the solar collector. The lower the
ambient temperature, the higher the heat losses and
consequently the lower the efficiency. In order to
minimize the negative effect of low ambient
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2nd Canadian Solar Buildings Conference
Calgary, June 10 – 14, 2007
temperatures on the solar collectors performance,
other design parameters should be altered such as the
air mass flow rate. A gap thickness of 0.025 m for
the first pass and 0.05 m for the second pass ensures
maximum efficiency of the solar collector.
ACKNOWLEGEMENT
The authors would like to express their gratitude for
the financial support of the Canadian Solar Building
Research Network.
NOMENCLATURE
A
Cp
Dh
E
H
h
I
T
k
Area (m2)
Specific heat at constant pressure (kJ/kg)
Hydraulic diameter
Exergy (W)
Gap thickness (m)
Heat transfer coefficient (W/m2 K)
Solar radiation (W/m2)
Temperature (K)
Thermal conductivity (W/m K)
m& Air mass flow rate (kg/s)
Nu Nusselt Number
Pr Prandtl Number
Q Heat transfer rate (W)
Re Reynolds Number
Greek letters
α Absorptivity
ε Emissivity
η Efficiency (%)
σ Stefan-Boltzman constant (W/m2 K4)
Subscripts
c Glass cover
f
Air
i
Inlet
o Ambient
p Plate
REFERENCES
[1] Renewable Energy: RD&D priorities Paris
OECD/IEA, 2006.
[2] T. Koyuncu, "Performance of various design of
solar air heaters for crop drying applications,"
Renewable Energy, vol. 31, pp. 1073-1088, 2006.
[3] S. Chemkhi, F. Zagrouba, and A. Bellagi,
"Drying of agricultural crops by solar energy,"
Desalination, vol. 168, pp. 101-109, 2004.
[4] D. R. Pangavhane, R. L. Sawhney, and P. N.
Sarsavadia,
"Design,
development
and
performance testing of a new natural convection
solar dryer," Energy, vol. 27, pp. 579-590, 2002.
[5] W. J. T. Barry Crittendn, Adsorption Technology
and Design. Boston Reed Educational and
Professional Publishing Ltd., 1998.
[6] M. H. Ahmed, N. M. Kattab, and M. Fouad,
"Evaluation and optimization of solar desiccant
wheel performance," Renewable Energy, vol. 30,
pp. 305-325, 2005.
[7] S. Pramuang and R. H. B. Exell, "The
regeneration of silica gel desiccant by air from a
solar heater with a compound parabolic
concentrator," Renewable Energy, vol. In Press,
Corrected Proof.
[8] F. L. Lansing, V. Clarke, and R. Reynolds, "A
high performance porous flat-plate solar
collector," Energy, vol. 4, pp. 685-694, 1979.
[9] A. A. Mohamad, "High efficiency solar air
heater," Solar Energy, vol. 60, pp. 71-76, 1997.
[10] L. Varshney and J. S. Saini, "Heat transfer and
friction factor correlations for rectangular solar
air heater duct packed with wire mesh screen
matrices," Solar Energy, vol. 62, pp. 255-262,
1998.
[11] I. T. Togrul and D. Pehlivan, "Effect of packing
in the airflow passage on the performance of a
solar air-heater with conical concentrator,"
Applied Thermal Engineering, vol. 25, pp. 13491362, 2005.
[12] V. K. Sharma, G. Rizzi, and H. P. Garg, "Design
and development of a matrix type solar air
heater," Energy Conversion and Management,
vol. 31, pp. 379-388, 1991.
[13] R. K. Swartman and O. Ogunlade, "An
investigation on packed-bed collectors," Solar
Energy, vol. 10, pp. 106-110, 1966.
[14] V. K. Sharma, S. Sharma, R. B. Mahajan, and
H. P. Garg, "Evaluation of a matrix solar air
heater," Energy Conversion and Management,
vol. 30, pp. 1-8, 1990.
[15] D. Singh and N. K. Bansal, "Analysis of a glass
solar air heater," Energy Conversion and
Management, vol. 23, pp. 231-236, 1983.
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2nd Canadian Solar Buildings Conference
Calgary, June 10 – 14, 2007
[16] M. K. Mittal and L. Varshney, "Optimal
thermohydraulic performance of a wire mesh
packed solar air heater," Solar Energy, vol. 80,
pp. 1112-1120, 2006.
[17] C. D. Ho, H. M. Yeh, and R. C. Wang, "Heattransfer enhancement in double-pass flat-plate
solar air heaters with recycle," Energy, vol. 30,
pp. 2796-2817, 2005.
[18] R. L. San Martin and G. J. Fjeld, "Experimental
performance of three solar collectors," Solar
Energy, vol. 17, pp. 345-349, 1975.
[19] W. Stahl, V. Wittwer, and A. Pfluger,
"Transparent insulation," Solar Energy Materials,
vol. 11, pp. 199-208, 1984.
[20] D. J. Nugent, "Experimental investigation of the
performance of an advanced solar air heater using
a porous matrix absorber, outdoor testing," in
Department of Mechanical and Manufacturing
Engineering, vol. Master. Calgary: University of
Calgary, 2003, pp. 104.
[21] Y. A. Çengel and M. A. Boles,
Thermodynamics : an engineering approach
Boston McGraw-Hill Higher Education, 2007.
[22] F. d. Winter, "Solar collectors, energy storage,
and Materials," in Solar Heat technologies:
Fundamentals and Applications, vol. 5. London,
England: The MIT press, 1990, pp. 1082.
[23] F. P. Incropera and D. P. DeWitt, Fundamentals
of heat and mass transfer 6th ed. Hoboken, NJ
John Wiley, 2007.
[24] University of Calgary Weather Station.
University of Calgary Campus, Calgary, Alberta,
Canada, 2000.
[25] "The National Climate Data and Information
Archive of Canada, 2000."
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2nd Canadian Solar Buildings Conference
Calgary, June 10 – 14, 2007
Solar Radiation
Porous
Absorber
Glass Covers
Air Inlet
Air Exit
Fig. 1 Advanced solar air-heater [10].
15
Experimental [20]
Calculated
o
Tf2, Colector outlet temperature, C
10
5
0
12:00
12:50
13:40
14:30
15:20
16:10
17:00
17:50
Time
-5
-10
-15
Fig. 2 Comparison between the model
results and experimental data, Mar. 1, 2003.
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2nd Canadian Solar Buildings Conference
Calgary, June 10 – 14, 2007
30
Experimental [20]
Calculated
25
o
Tf2, Colector outlet temperature, C
20
15
10
5
0
8:30
9:20
10:10
11:00
11:50
12:40
13:30
14:20
15:10
16:00
16:50
17:40
Time
-5
-10
-15
Fig. 3 Comparison between the model
results and experimental data, Mar. 2, 2003.
16
20 oC
o
10 C
o
0 C
o
-10 oC
-20 C
14
ηII, %
12
10
8
6
4
2
0
0.00
0.05
0.10
0.15
0.20
m, kg/s.m
Fig. 4: Effect of air mass flow rate
on solar collector thermal efficiency.
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2nd Canadian Solar Buildings Conference
Calgary, June 10 – 14, 2007
η, %
100
90
80
70
60
50
40
30
20
10
0
0.00
o
20 oC
10 o C
0 C
-10 oC
-20 oC
0.05
0.10
0.15
0.20
m, kg/s.m
Fig. 5: Effect of air mass flow rate
on solar collector exergetic efficiency.
100
90
80
η, %
70
60
50
20 oC
10 oC
0 oCo
-10 oC
-20 C
40
30
20
10
0
0.00
0.02
0.04
0.06
(T
f2
0.08
0.10
0.12
− Ti ) I , m2 .K / W
Fig. 6: Performance curves for different
ambient temperature.
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2nd Canadian Solar Buildings Conference
Calgary, June 10 – 14, 2007
16
14
12
ηII, %
10
8
Ta=20 oC
6
4
Ta=10 oC
Ta=0 oC
Ta=-10 oC
2
Ta=-20 oC
0
0.01
0.03
0.05
(T
f2
0.07
0.09
0.11
0.13
− Ti ) I , m2 .K / W
Fig. 7: Performance curves for different
ambient temperature.
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2nd Canadian Solar Buildings Conference
Calgary, June 10 – 14, 2007
90
80
70
η, %
60
50
40
30
20ooC
10 oC
0 Co
-10oCC
-20
20
10
0
0.0
0.2
0.4
0.6
0.8
1.0
1.2
1.4
1.6
1.8
2.0
1.8
2.0
LW
Fig. 8: Efficiency versus length to width
ratio
3.0
o
20 oC
10 oC
0 C
-10 ooC
-20 C
2.5
ηII, %
2.0
1.5
1.0
0.5
0.0
0.0
0.2
0.4
0.6
0.8
1.0
LW
1.2
1.4
1.6
Fig. 9: Exergetic efficiency versus
length to width ratio.
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2nd Canadian Solar Buildings Conference
Calgary, June 10 – 14, 2007
95
H1=0.05,H2=0.05m
H1=0.05,H2=0.025m
90
H1=0.025,H2=0.05m
H1=0.025,H2=0.025m
85
η, %
80
75
70
65
60
0.00
0.04
(T
0.08
0.12
− Ti ) I , m .K / W
2
f2
Fig. 10: Efficiency versus gap thickness.
100
I, W/m2
Tf2, oC
90
80
70
60
50
40
30
20
10
0
3
5
7
9
11
13
Time, hr
15
17
19
21
Fig. 11: Solar air heater performance for typical
weather conditions, Calgary, Canada, Jul. 5, 2000 .
12