Performance Analysis of Parabolic Trough Collector in Hot Climate
Performance Analysis of Parabolic Trough Collector in Hot Climate
Performance Analysis of Parabolic Trough Collector in Hot Climate
e-abs
and
the absorber surface Q
a-abs
. Part of the energy taken by the absorber is transferred to the
heat transfer fluid by forced convection V
a-f ;conv
, the remaining energy is transferred back to
the glass envelope by radiation Q
a-e;rad
and natural convection Q
a-e;conv
and lost through the
support brackets by conduction Q
cond;bracket
as well. The heat loss from the absorber in the
form of radiation and natural convection is conducted by the glass envelope and lost to the
environment by convection Q
e-sa;conv
and radiation Q
e-s;rad
, together with the energy absorbed
by the glass envelope Q
e-abs
. In order to obtain the partial differential equations that govern
the heat transfer phenomena, an energy balance is applied over a section of the solar
receiver. The detailed analysis of the partial differential equations and method of solution
can be found in Rohsenow et al. [29]. After applying the energy balance on a control volume,
assuming unsteady state and incompressible fluid, the following partial differential equation
can be obtained:
.
conv , f a
2
f
f f , p f
f
f , p f a , i
Q +
2
V
+ T C
z
m =
t
T
C A
(4)
a , i f
f
f
A
m
= V
(5)
where
f
m
a-f;conv
is the heat transfer by convection from the absorber to
HTF per unit length and is given by:
) -T (T k Nu = Q
f a f f
'
conv , af
(6)
British Journal of Applied Science & Technology, 4(14): 2038-2058, 2014
2044
where Nu
f
is the Nusselt number, k
f
is the thermal conductivity of HTF, and T
a
is the absorber
wall temperature. The equation for fully developed turbulent flow (Re
D
> 2300) and
convective heat transfer in circular ducts introduced by Kaka et al. [30] is adapted:
( )( )
( )
0.11
w
3 / 2 2 / 1
f
D f
uf
)
Pr
Pr
(
1) (Pr 2 / C 12.7 + 1
Pr 1000 Re 2 / C
= N
(7)
f
a , i f
D
D V
= Re
(8)
where C
f
is the friction coefficient (Fanning friction factor), Re
D
is the Reynolds number
based on the absorber inner pipe diameter (D
i,a
), Pr is the Prandtl number, and v
f
is the
kinematic viscosity of HTF. This equation is valid for 2300 s Re
D
s 5 x 10
6
and 0.5 s Pr s
2000. The thermal properties should be evaluated at the bulk mean heat transfer fluid
temperature, except Pr
w
which is evaluated at the absorber wall temperature. The
convection heat transfer coefficient for rough tubes can be estimated using the friction
coefficient from Colebrook and White as referred by Rohsenow et al. [29]:
70 Re 5
C Re
35 . 9
+
D
2
Ln 1.7372 - 3.48 =
C
1
f D a i, f
(9)
With
f
f
V
= Re
(10)
For laminar flow, (Re
D
s 2300), the Nusselt number on walls with uniform temperature is
given by Nu
f
= 3.66. Two heat transfer mechanisms occur between the absorber and the
glass envelope: convection heat transfer and thermal radiation. Convection heat transfer
depends on the annulus pressure; experimental work as stated in Rohsenow et al. [29] has
shown that heat transfer loss is independent on the annulus vacuum pressure for pressures
above 1Torr. At pressures below 1Torr, molecular conduction is the heat transfer
mechanism while for pressure above 1Torr, natural convection takes place. Fig. 1 shows the
control volume used for the absorber analysis.
Fig. 1. Control volume for absorber analysis
British Journal of Applied Science & Technology, 4(14): 2038-2058, 2014
2045
Applying the energy balance on the control volume:
'
bracket cond,
'
rad e, - a
'
conv e, - a
'
conv f, - a
'
abs a,
a
a a
a
a , p a a
Q
Q - Q Q Q +
z
T
k
z
A =
t
T
C A
-
- -
(11)
where A
a
is the absorber cross section area, ( ) ( )
2
a , i
2
a , o a
D - D /4 = A
, D
o,a
is the outer pipe
diameter, D
i,a
is the inner pipe diameter, Q
a;abs
is the solar absorption in the absorber per
receiver length, Q
a-f,conv
is the heat transfer by convection from absorber to heat transfer fluid
per unit length, Q
a-e,conv
is the heat transfer by convection from absorber to envelope per unit
length, Q
a-e,rad
is the heat transfer by radiation from absorber to glass envelope per unit
length, and Q
cond,bracket
is the heat conduction through support brackets per unit length.
Stainless steel is normally used as the absorber tube material. Thermal radiation analysis for
one surface implies that all surfaces that can exchange radiative energy with the surface
must be considered simultaneously. How much energy two surfaces exchange depends on
their size, separation distance and orientation. In order to carry out the radiative heat transfer
analysis, the view factors for a short annulus proposed by Siegel and Howell [31] are
employed. The heat transfer mechanisms from the glass envelope to the surroundings are
convection and radiation. Convection heat transfer distinguish two cases: Wind guided
(forced convection) and no wind (natural convection). The radiation heat transfer is basically
between the glass envelope and either the sky or the collector surface. The maximum
radiation heat loss takes place when the solar receiver is assumed to be surrounded only by
the sky. The energy balance on the control volume results in the following partial differential
equation:
'
rad s, - e
'
conv sa, - e
'
rad' e, - a
'
conv e, - a
'
abs e,
e
e e
e
e , p e e
Q Q - Q + Q + Q +
z
T
k
z
A =
t
T
C A
-
(12)
where Q
e,abs
is the solar absorption in the envelope per receiver length, Q
a-e,conv
is the heat
transfer by convection from the glass envelope to the surrounding air per unit length, and, Q
e-
s,rad
is the heat transfer by radiation from the glass envelope to the sky per unit length. The
heat transfer by convection per unit length from the glass envelope to the surrounding air is
calculated as:
) T (T D h = Q
e e o, e
'
conv sa, - e
-
(13)
with
D
k Nu
= h
o,e
e e
e
(14)
For no wind conditions, Kaka et al. [30] recommended the following expression for a
horizontal cylinder under natural convection:
, )
2
6 / 1
9 / 16
16 / 9
D
e
0.559/Pr 1
Ra
0.387 0.60 Nu
+
+ =
(15)
British Journal of Applied Science & Technology, 4(14): 2038-2058, 2014
2046
In order to simplify the model, it is assumed that half of the receiver surface is surrounded by
the mirror and the other half by the sky. The heat flux and radiosity are calculated for each
surface and a set of energy balance equations for internal and external surfaces are written.
In this work, it is assumed that the collector mirror temperature is approximately the ambient
temperature. When the collector surface is not included in the analysis, maximum radiation
heat transfer loss occurs, and this means that the glass envelope is assumed to be totally
covered by the sky. So, the heat flux (
esi
) for the area considered is expressed by Stefan-
Boltzmanns equation:
) T - (T = q
4
sky
4
esi esi esi
(16)
where is Stefan-Boltzmann constant (5.67 10
8
W /m
2
K
4
), c is the object emissivity, T
esi
is body temperature, T
sky
ambient temperature.
Several relationships have been proposed to connect T
sky
for clear skies, to other
meteorological variables. In the absence of meteorological data such as relative humidity,
dew point temperature, etc., a simple relation given by Swinbank [32] can be used:
T 0.0553 = T
1.5
sky
(17)
The energy absorbed in the solar receiver is affected by optical properties and imperfections
of the solar collector ensemble. For a concentrating collector, the effective optical efficiency
is defined as long as the direct beam radiation is normal to the collector aperture area. When
the beam radiation is not normal, the angle of incidence modifier (IAM) is included to account
for optical and geometric losses due to angles of incidence greater than 0. The IAM
depends on geometrical and optical characteristics of the solar collector. It is defined as the
quotient between the transmittance-absorptance product at the angle of the incidence of
radiation and that at normal incident radiation.
3. NUMERICAL MODEL
The set of partial differential equations (PDE) was discretized for steady state conditions by
using the finite difference method and taking into account the dependence of thermal
properties on temperature. Turning to the heat transfer fluid, discretization by backward
differencing creates a set of algebraic equations. For the absorber and the envelope, the
discretization is carried out using the central difference and thus, another set of algebraic
equations is obtained. Finally, the boundary conditions for each element are set down.
The proposed model includes the thermal interaction between absorber-envelope, and
envelope-envelope. These thermal radiation losses were not included in other existing
models. To account for the thermal interaction between adjacent surfaces, a comprehensive
radiative analysis was implemented for heat losses in the absorber and the glass envelope.
A review of the equations for convective heat transfer loss was performed as well, and new
equations were incorporated to the present model. The effects of heat conduction in the
collector tube wall and the mixed convection in the inner tube, which have been neglected in
previous studies, are also taken into consideration in the present model. The resulting set of
nonlinear algebraic equations is solved simultaneously using numerical techniques. These
equations are solved using an implicit Euler scheme. Linear system equations are written
British Journal of Applied Science & Technology, 4(14): 2038-2058, 2014
2047
using the matrix-vector notation, Ax = b, where A is the matrix of coefficients for the system,
x the column vector of the unknown variables x
1
,, x
n
, and b a given column vector.
Collector thermal efficiency (q) is defined as the ratio of energy collected by the working fluid
to the direct normal solar radiation incident upon the collector aperture. It is typically
determined by testing a collector over a range of high temperatures and is expressed as:
A G
) T (T m
=
a
f i, f o,
-
(18)
Where where
f
m
a-e,conv
is the heat transfer by convection from the glass envelope to the surrounding
air per unit length, and, Q
e-s,rad
is the heat transfer by radiation from the glass envelope to
sky per unit length and Q
cond,bracket
is the heat conduction through support brackets per unit
length.
A simulation model compatible with the transient simulation program (TRNSYS) was
developed to determine the thermal performance of a typical parabolic trough configuration
under Kuwait climatic conditions. A TRNSYS Studio project is designed by setting up the
connecting components graphically in the simulation study as explained by Klein et al. [33].
Parameters treated were: area of collectors absorber, overall heat loss coefficient from the
absorber, reflectivity of the reflecting surface and absorptivity and emissivity of the absorber.
The parametric study was conducted for different mass flow rates and concentration ratios
utilizing hourly solar radiation data for Kuwait. The processed solar insolation data consist of
the beam radiation and diffuse radiation as well as the total radiation on a tilted surface
measured in W/m
2
.
4. RESULTS AND DISCUSSIONS
In order to validate the present heat transfer model, it was first compared with experimental
data obtained by Dudley et al. [34]. In addition, to corroborate the improvement in the
equations and radiation analysis proposed here, Garca Valladares et al. [35] compared the
numerical model results against those from other solar receiver heat transfer models. The
experimental results were obtained from a solar collector assembly module (LS-2) placed on
an AZTRAK rotating platform at the SNL. Two different selective coatings were used in this
test: black chrome and cermet. Cermet has better radiative properties (low emissivity) at
high temperatures than black chrome and does not oxidize if the vacuum is lost.
The Sandia test was performed for both: Full sun and no sun conditions and different
scenarios for the annulus of the heat collection element (HCE). Among them: vacuum intact
(In vacuo), with annulus pressure at 10
-4
Torr, lost vacuum (annulus filled with ambient air),
British Journal of Applied Science & Technology, 4(14): 2038-2058, 2014
2048
and glass cover completely removed (bare tube). Fig. 2 presents the results obtained for the
collector efficiency. As seen in the figure, the current calculated collector efficiency follows
the trend of the experimental values and all of the results are within the error bars. As
expected, the higher efficiencies are obtained when the annulus is In vacuo, but in both
cases, air and vacuum, the collector efficiency drops gradually at high temperature. This
behavior is more noticeable for black chrome coating.
Fig. 2. Comparison of collector efficiency as calculated with the proposed model to
experimental data [34] and NREL model [35]
In case of Cermet coating, the model developed by Garca-Valladares and Velsquez [35]
shows some discrepancies at low temperatures in the collector efficiency. These
discrepancies may be attributed to their assumptions: negligible conduction at each end of
each trough and that only radiation heat loss takes place between the receiver and the glass
envelope for In vacuo annular space. These assumptions were made neither in the present
nor in the NREL model. So, the present model and NREL model introduced by Dudley et al.
[34] give almost similar collector efficiency values.
The angle of incidence modifier, K
(
i
), enables the performance of the collector to be
predicted for solar angles of incidence other than 0 (normal). Simulations using the present
numerical model are carried out setting up a value of
i
and then calculating K
.
The angle of incidence modifier represents the ratio between a specified thermal efficiency
value and the peak efficiency of the collector at zero incidence. Results of simulation are
presented in Fig. 3. Regression analyses provided the following equations for K
as a
function of
i
:
125 . 1 + ) ( 10 884 . 3 ) ( 10 2146 . 1 + ) ( 10 2.235 - = K
i
3 2
i
4 3
i
-6
U
- (20)
078 . 1 + ) ( 10 773 . 1 + ) ( 10 561 . 1 ) ( 10 9.272 = K
i
3 2
i
4 3
i
-7
G
- (21)
50
55
60
65
70
75
80
0 100 200 300 400
C
o
l
l
e
c
t
o
r
E
f
f
i
c
i
e
n
c
y
(
%
)
Temperature Above Ambient (0C)
Present Model
Experimental
[34]
British Journal of Applied Science & Technology, 4(14): 2038-2058, 2014
2049
Coefficients upon solving equations (20) and (21) were 0.958 and 0.961. Up to an angle of
incidence of approximately 25 the glass-shielded receiver performed slightly better but, for
greater angles, performance declined more rapidly and was inferior to that of the unshielded
receiver.
0 5 10 15 20 25 30 35 40 45 50 55 60
Angle of Incidence (deg)
0.50
0.60
0.70
0.80
0.90
1.00
1.10
I
n
c
i
d
e
n
t
A
n
g
l
e
M
o
d
i
f
i
e
r
Glass Shielded
Unshielded
Fig. 3. The angle of incidence modifier for unshielded and glass-shielded receivers
In Fig. 3, the calculated value of K
by about
25% for the unshielded receiver and by 17.7% for the glass shielded unit. Two factors are
primarily responsible for the decline in performance of a PTSC with increasing
i
,: the
geometric reduction in irradiance falling on the aperture as
i
increases or cosine effect and
the change in optical efficiency (due to differences in light interaction) with the reflective
surface of the collector, the glass shield (if present) and the absorber. Nothing can be done
to account for the first effect other than tilting the PTSC constantly so as to keep it
perpendicularly oriented to the sun.
The thermal losses of the collector receiver depend on operating temperature. The thermal
losses through PTSC change in different ways depending on the receivers configuration and
operational conditions. As shown in Fig. 4, the convection loss from the absorber tube to
supporting structures is the largest. It follows in decreasing order the radiation loss from the
glass envelope to ambient air, while the smaller loss is the convection loss from the glass
envelope to ambient air. Thermal losses are always present if there is a temperature
difference between receiver and ambient whether solar radiation is available or not.
Increased solar radiation results in increased solar energy absorbed by the collector. It
should be noted that thermal losses also increase due to the increased collector
temperature. However, this increase is smaller than the enhanced absorbed solar energy.
As shown from Fig. 5, PSTC efficiency increases with increasing solar radiation.
British Journal of Applied Science & Technology, 4(14): 2038-2058, 2014
2050
0 50 100 150 200 250 300 350 400
Receiver Operating Temperature (
0
C)
0.45
0.50
0.55
0.60
0.65
P
e
r
c
e
n
t
o
f
E
n
e
r
g
y
t
o
T
o
t
a
l
C
o
l
l
e
c
t
e
d
Loss due to Conduction to Brackets
Loss due to Convection to Ambient
Loss due to Radiation to Sky
Fig. 4. Heat loss from an absorber tube
0 50 100 150 200 250 300 350 400
Average Temperature above Ambient (
0
C)
0.30
0.35
0.40
0.45
0.50
0.55
0.60
E
f
f
i
c
i
e
n
c
y
(
-
)
1000 W/m2
700 W/m2
400 W/m2
Fig. 5. Variation of parabolic trough collector efficiency with solar radiation
The angle of incidence modifier (IAM) is a very significant factor impacting on the solar
efficiency, and can be approximately estimated as the cosine of the angle of incidence. IAM
depends on time of day, date, the location and orientation of the aperture, and whether the
collector is stationary or tracks the sun movement about one or two axes. Collector efficiency
reaches a maximum value only at zero angle of incidence. The efficiency of a PTSC
decreases when the angle of incidence of solar beam increases, as shown in Fig. 6. The
British Journal of Applied Science & Technology, 4(14): 2038-2058, 2014
2051
effect of the angle of incidence is to reduce radiation arriving at the absorber tube.
Therefore, at noon time PTSC has the smallest angle of incidence and the highest efficiency.
0 50 100 150 200 250 300 350 400
Average Temperature above Ambient (
0
C)
0.20
0.25
0.30
0.35
0.40
0.45
0.50
0.55
0.60
E
f
f
i
c
i
e
n
c
y
(
-
)
Incident angle=0 deg
Incident angle=25 deg
Incident angle=45 deg
Fig. 6. Variation of collector efficiency with angle of incidence
When the annulus between the receiver surface and the glass envelope is in vacuum state,
conduction and convection across the annulus are effectively eliminated. Once air is
introduced into the vacuum space, measured losses increase significantly since conduction
and convection begin to transfer heat to the glass envelope as shown in Fig. 7. Radiation
loss from the heated receivers metal surface to the glass envelope does not change
significantly when air is allowed into the annulus.
0 50 100 150 200 250 300 350 400
Receiver Operating Temperature (
0
C)
0.40
0.45
0.50
0.55
0.60
0.65
0.70
0.75
P
e
r
c
e
n
t
o
f
E
n
e
r
g
y
t
o
T
o
t
a
l
C
o
l
l
e
c
t
e
d
Optical Losses
Vacuum
Air
Fig. 7. Collector efficiency with air in annulus
British Journal of Applied Science & Technology, 4(14): 2038-2058, 2014
2052
5. CASE STUDY
A case study is carried out to investigate the performance of solar heating and cooling
systems in Kuwait. The solar system is designed to satisfy a significant part of the nominal
space heating load, water heating load and cooling load of a typical Kuwaiti house. A
schematic diagram of the system studied is represented in Fig. 8. The system consists of
parabolic trough collectors integrated into the roof of the building, a storage tank, an
absorption chiller, heat exchanger and auxiliary units. The main purpose of this stage is to
investigate the feasibility of integrating parabolic trough collectors in an existing building in
order to convert it to nearly zero net energy building.
Buildings are responsible for a significant part of the total primary energy consumption.
Thus, researchers and designers thrive to develop energy efficient buildings from renewable
energy sources. Many new buildings incorporate solar energy technology to attain
environmental goals as opposed to conventional energy in order to reduce greenhouse and
other gases emission, mainly CO
2
as presented by Zhai et al. [36], Yin et al. [37], Thomas et
al.[38], Bourrelle et al.[39], Bucking et al. [40] and, Zeiter et al. [41].
A solar heating and cooling system operates in four different modes. When solar energy is
available for collection and there is load demand, heat is supplied directly from the collector
to the heating or cooling unit. When solar energy is available for collection and there is no
heat or cooling demand, heat is stored in the storage unit. On the other hand, if solar energy
is not available for collection and there is load demand, the storage unit then supplies heat to
the heating or cooling unit. However, if the storage temperature is not sufficient, the heating
or cooling load is supplied by the auxiliary source.
Fig. 8. Schematic diagram of a solar heating and cooling system
The solar cooling system consists of single-effect lithium bromide water absorption chiller.
The chiller model is based on a commercially available LiBr-H
2
O absorption chiller system,
Arkla model WF-36. The Arkla chiller has a nominal cooling capacity of three tons
(37980kJ/h). Units of different capacity are approximated by scaling the Arkla performance.
Hot water is supplied to the air conditioner at temperatures ranging between 87C and 93C,
leaving this unit 10C cooler than the supply and ending into the storage, or to the auxiliary
heater if storage temperature is below 77C. Whenever hot water from the storage
British Journal of Applied Science & Technology, 4(14): 2038-2058, 2014
2053
temperature is below 87C, auxiliary heat is supplied to raise its temperature to 87C. When
storage temperature is below 77C, it is not used, and the auxiliary heater carries the full
cooling load.
The performance of air conditioning systems is expressed by their coefficient of performance
(COP). COP determines how many units of cooling/heating is attained for every unit of input
energy. In addition, a constant COP model of a vapor compression air conditioner is
included as a secondary cooling auxiliary unit, so that the energy required to meet space
cooling is always provided even if the absorption machine cannot meet the full load.
5.1 Building load
The model dwelling under study is a typical Kuwaiti house in a remote area. The computer
program is provided with an estimate of typical loads encountered in the house. An hourly
load schedule is written to an input file. The program is provided with an estimate of the
number of hours that the load is utilized during a typical day for each load. The roof area is
large enough so that the parabolic trough collectors can be spaced widely to minimize
shading losses. The domestic water demand is based on a figure of 60 kg/day/person. The
system has been designed to supply cooling up to a load peak of about 21 kW (6 tons of
refrigeration), which occurs in August.
Different collector parameters and operating conditions were essayed to maximize the
annual energy generated by the parabolic trough collector. For building simulation, type 56
included in version 16 of TRNSYS developed by Klein et al. [33] is employed. This
component models the thermal behavior of a building with multiple thermal zones. The
building can be completely described using this component from a set of external files. The
files can be generated based on user supplied information by running the preprocessor
program called TRN Build.
From simulation with TRN build (Type 56) of TRNSYS to determine the buildings demand
for air conditioning, results indicate that this demand spans from April to October, with critical
periods of maximum load in the months of June, July and August. Study results showed that
the air conditioning system consumed about 83% of the total building energy requirement.
The remaining energy consumption is distributed between space heating 11%, and domestic
water heating, 6%. For hot thermal storage, a stratified liquid storage tank, with two flow inlet
and two flow outlet, Type 60, is adapted. Several data files of single-effect absorption chillers
employing LiBrH
2
O solution as working fluid, (type 107) according to Yazaki chiller WFC
SH10 are used. Other model types essayed were: Type 15, TMY2 and Type 3b for pump
simulation.
5.2 Solar Fraction
The thermal performance of a solar system is usually measured by the solar fraction (F).
Solar fraction is defined as the fraction of the load met by solar energy. Fig. 9 shows the
variation of solar fraction of space heating (F
s
), domestic water heating (F
D
), and cooling
load (F
Ac
) with collecting area. As seen from the figure, the solar space heating load and
domestic water heating load are completely satisfied for areas of about 29 m
2
. Conversely,
the space cooling load requires much greater areas. The minimum required collector area is
about 82 m
2
which can supply the cooling load of a typical residential house for sunshine
hours under all hot weather conditions in Kuwait, with a maximum cooling load of
approximately 21 KW (6 ton refrigeration).
British Journal of Applied Science & Technology, 4(14): 2038-2058, 2014
2054
0 10 20 30 40 50 60 70 80 90 100
Area (m
2
)
0.0
0.2
0.4
0.6
0.8
1.0
S
o
l
a
r
F
r
a
c
t
i
o
n
(
-
)
F
AC
F
D
F
s
F
D
Water Heating Load
F
S
Space Heating Load
F
AC
Cooling Load
Fig. 9. Solar fraction variation with collecting area
The coefficient of performance (COP) of the absorption chiller is approximately 0.66 which is
within the accepted practical values of the conventional lithium bromide system.
A computer routine was developed to determine the yearly reduced CO
2
emission as a result
of integrating parabolic trough collectors on the house roof. The annual variation of CO
2
decreased emission vs. collector area is presented in Fig. 10. In this figure at collecting area
values satisfying house cooling load, the calculated CO
2
emissions reduction was
12.3tonne/year.
30 40 50 60 70 80 90 100 110 120 130 140
Collector Area (m
2
)
11.2
11.4
11.6
11.8
12.0
12.2
12.4
12.6
A
v
o
i
d
e
d
C
o
2
e
m
i
s
s
i
o
n
(
t
o
n
n
e
/
y
e
a
r
)
Fig. 10. CO
2
emissions reduction variation with collecting area at zero azimuth Angle
British Journal of Applied Science & Technology, 4(14): 2038-2058, 2014
2055
6. CONCLUSIONS
This paper investigates the performance of parabolic trough collectors as well as solar
heating and cooling systems in Kuwait climate. A theoretical model that takes into
consideration some factors not included in previous models is proposed. Based on present
results, the following conclusions can be drawn:
- The performance of parabolic trough collectors can be significantly enhanced by
optimizing its parameters as well as operating conditions. Reducing the heat transfer
losses can significantly improve collector efficiency.
- Convection loss from the absorber tube to the supporting structures is the largest
among the other losses (conduction and radiation).
- The angle of incidence modifier is an important factor impacting on the solar
efficiency. At temperature 150C, parabolic trough collector efficiency decreases
from about 0.55 at angle of incidence 0 to about 0.35 at angle of incidence 45.
- At noon time, PTSC has the smallest angle of incidence and the highest efficiency.
- When the annulus between the receiver surface and the glass envelope is In vacuo,
conduction and convection across the annulus are effectively eliminated.
- Space and domestic water heating loads can be completely provided by parabolic
trough collectors.
- The minimum required parabolic trough collecting area is about 82 m
2
to supply
cooling loads of a typical residential house under all climatic conditions in Kuwait.
- Total reduction in CO
2
emission may reach to some 12.3tonne/year per building.
- The results of the present study should encourage widespread utilization of solar
energy systems which will help in keeping our environment healthy and clean.
ACKNOWLEDGEMENTS
The authors would like to express their sincere gratitude to the Public Authority for Applied
Education and Training (PAAET), Kuwait for supporting and funding this work, Research
Project No. (TS11-09), Research Project Title Assessment of Parabolic Trough Collector
Performance in Kuwait Climate.
COMPETING INTERESTS
Authors declare that there are no competing interests.
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_________________________________________________________________________
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