Issue
Eur. Phys. J. Appl. Phys.
Volume 87, Number 3, September 2019
Materials for energy harvesting, conversion, storage and environmental engineering (Icome 2018)
Article Number 30902
Number of page(s) 12
Section Physics of Energy Transfer, Conversion and Storage
DOI https://doi.org/10.1051/epjap/2019190032
Published online 26 November 2019

© EDP Sciences, 2019

1 Introduction

The performance improvement of solar collectors consists in reducing the thermal losses of the absorber with a judicious collector design, high energy efficiency and pertinent selection of components. It is worth mentioning that a number of studies have been achieved in the field of solar collectors. Duffie et al. [1] developed a model for planar solar collectors where the temperatures of the absorber, back plate and glazing cover are calculated by considering the stationary one-dimensional regime. The effects of the solar collector surrounding conditions, such as the speed and direction of speed, age of the collector surface, convection thermal losses, thermal inertia and incidence angle, were numerically explored and experimentally validated by Rodriguez et al. [2,3]. In order to study the thermal behavior of solar collectors, Molero et al. [4] developed an unsteady three-dimensional mathematical model to study several design configurations. Minn et al. [5] studied the temperature of the absorber and its influence on the fluid flow inside the tubes. Letz et al. [6] performed an analysis study regarding the behavior of an air collector under natural and artificial sunlight, using the nodal method, in order to solve the thermal balance equations. It turned out that the solar collector orientation has an important impact on the global efficiency of the collector. Moreover, Dang et al. [7] conducted an experimental study related to the effect of the south orientation on the performance of the solar collector. In order to find a low-cost, optically efficient and long-lasting absorber coating, Henry et al. [8] employed black nickel (Nis ZnS), black chrome (CrO) and black iron (FeO) as a plated coating. The effects of the properties and number of layers of glass used on the thermal performance of the solar collector were investigated. Similarly, Youcef-Ali [9] carried out an experimental work on solar collectors with double and triple glazing layers. In order to investigate the effect of the free space between two glazing layers on the thermal performance of a planar solar collector, Ben Guehza et al. [10] carried out a theoretical investigation on a double-glazed collector for the purpose of minimizing the thermal losses through convection. Furthermore, Njomo [11] conducted a numerical study of the thermal behavior of a combined plastic-glass cover air collector, where plastic is used as a protection material against breakage. For this reason, polymer materials began to be seriously considered in the design of solar collectors. Therefore, Cristofari et al. [12] investigated the performance of a solar thermal installation where the collector is entirely built with polymer materials. Benyelles et al. [13] made an attempt to reduce the thermal losses using silica aerogel as an insulating material for covering the absorber. For their part, Vestlund et al. [14] replaced the air enclosed inside the free space between the absorber and the glass cover with an inert gas; they found out that the thermal losses could be reduced by approximately 20% when using inert gasses like Argon, Krypton and Xenon. However, much work is still needed in order to determine the influence of the thickness of free space between the absorber and glass as well as the effect of varying the gas temperature. To improve the heat-exchange efficiency, chicanes were added inside the free space between the absorber and glass. This innovative idea was proposed by Francia [15] in 1961. Several studies have raised questions of considerable interest for using several chicanes and studying related parameters such as the number of chicanes, their shape and geometry, inclination angle and finally the material these chicanes are made of [1619]. Other works were interested in the effect of the entire solar collector inclination and the Rayleigh number on the flow regime. Thus, for a free space cavity with an aspect ratio equal to 44, Hollands et al. [20] were able to determine the critical Rayleigh number for horizontal, vertical and inclined positions. Many works [2137] concentrated on the issue of convective heat transfer as well as on related topics, such as the residence time of a working fluid inside the free space, design parameters of specific channels inside that space, inclination of the whole solar collector with respect to the vertical direction, addition of chicanes as well as their arrangement and shapes, the temperature gradient between the cavity sides and even the entire shape of the solar collector. In 2012, D'Antoni et al. [38] went through the literature produced to date on the high capacity of solar thermal collectors in order to highlight the wide range of possible variants and applications and to share the information gathered here for future developments. Some researchers were interested in designing specific solar collectors for improving the global solar collector efficiency [3942]. Similarly, in 2017, Suresh et al. [43] suggested that the solar conversion efficiency is much higher for process heating than for electricity production; it is worth noting that industrial process heating applications account for a significant part in industrial energy consumption. In Algeria, Ihaddadene et al. [44] and Kabeel et al. [13,19,4555] conducted a series of studies on the flow rate of a working fluid, the automatic control of a solar collector position with respect to sun, the heat transfer problem, solar collector shape (parabolic, plane, etc.) and refrigeration. These same authors carried out several experiments on solar collectors, by varying either the type of material used (heat absorber and insulator, glazing, etc.) and its thickness, or by introducing new concepts on the device itself (chicane, mini-concentrator, etc.) [5662].

The present work deals with a comparative experimental study of two models of flat-air solar collectors, namely a solar collector with Black Mini Solar Concentrators (BMSC), already studied by authors in [47], and another model similar to the first one, with the same geometry and dimensions, with Mini Solar Concentrators supplied with Mirrors (MSCMR) instead of black mini solar concentrators.

2 Characteristics of the collectors and description

The two models of solar collectors mentioned above are identical. These are simple solar air collectors. The difference lies in the design of each of the mini-concentrators implanted in the two solar collectors, namely the BMSC or the MSCMR. It was agreed to first present the dimensions and characteristics of the collector, and then describe the models used, namely BMSC and MSCMR.

The external dimensions of the collectors under study are shown in Figure 1a; they are 210 cm long, 112 cm wide and 13.4 cm thick. The net dimensions, i.e., the test vein where the air circulates, of the collector are 200 cm for the length, 100 cm for the width and 6.4 cm for the thickness. In addition, a rotation speed fan of 1300 rpm and 12 cm in diameter was used to investigate the influence of forced convection on the performance of the solar collector (Fig. 1b).

Figures 2a and b respectively shows the two solar collector models under study, namely the solar collector with black mini solar concentrators (BMSCs) and the one with mini solar concentrators supplied with mirrors (MSCMRs).

The collector consists of a single 5-mm-thick transparent glass cover, a thin black absorbing aluminum sheet metal plate. The rear and side insulation is ensured by a 40 mm polystyrene layer, covered with a 30 mm thick glass wool layer. The whole is placed in a wooden case. The specifications and characteristics of the solar air collector are given in Table 1.

thumbnail Fig. 1

(a) Solar collector under study (without mini concentrators). (b) Fan at the exit of the collector.

thumbnail Fig. 2

(a) Solar collector with black mini solar concentrators (BMSCs). (b) Solar collector with mini solar concentrators and mirrors (MSCMRs).

Table 1

Characteristics of solar collector component.

3 Description and arrangement of BMSC and MSCMR

3.1 Black mini solar concentrators

Flat plate air solar collectors have low thermal performance compared to liquid collectors. This is mainly attributed to the low thermo-physical properties of air. A large number of researchers have used different methods to increase the performance of these collectors. One of these methods recommends the addition of obstacles or fins, also called baffles, of various shapes, inside the airflow vein.

The black mini solar concentrators (BMSCs) have a curved shape, with a radius of curvature of 1 cm, a length of 8 cm and a width of 3.5 cm (Fig. 3). The BMSCs are fixed directly to the absorber and are distributed over its entire surface according to the distribution presented in Figure 4. The total number of baffles is 176. These baffles are spread over 27 rows; they are arranged in such a way that a row comprises 06 baffles and the preceding one 07 baffles, alternately, so as to obtain the configuration illustrated in Figure 4.

The distance (pitch y) separating two consecutive rows is equal to 5 cm; the distance (pitch x) between two consecutive baffles of the same row is also equal to 5 cm (see Fig. 5). The arrangement of these BMSCs in this way forces the air to flow between the BMSCs of row (i) and to bypass the BMSC of row (i + 1) that lies frontally to the air flow.

thumbnail Fig. 3

Dimensions of BMSCs.

thumbnail Fig. 4

Arrangement of BMSCs within the absorber of the solar collector.

thumbnail Fig. 5

Pitch separating two consecutive BMSCs.

3.2 Mini solar concentrators with mirrors (MSCMR)

These are mini-concentrators of cylindro-parabolic shape, similar to the BMSC (same shape and dimensions), except that the MSCMR has got two faces; an inner face with a reflecting mirror for focusing the incident radiation and then reflecting it towards the absorber. The second external face is painted in a matte black color in order to increase the temperature inside the collector (Fig. 6).

The comparison between the two collector models under study is made on the basis of average temperature calculations, for each type of collector, during the four measurement days.

thumbnail Fig. 6

Typical mini-concentrators.

4 Governing equations and experimental set up

4.1 System of equations

The performance improvement of a solar collector requires the prediction of the thermal losses within the collector. In order to determine the heat losses, it is required to take into account the different thermal exchanges taking place within the system:

Heat losses in a flat collector are composed of three parts, i.e. the front heat loss, the rear heat loss and the side heat loss which is neglected in comparison to the first two types of heat loss because the collector is supposed to be well insulated.

4.1.1 Front thermal losses of the collector

These losses consist of two parts, and are attributed to heat exchanges between glass and the ambient air and also between the absorber and glass.

4.1.1.1 Loss between glass and the surrounding

Glass is subjected, on the one hand, to a convective heat exchange with the ambient air, which is expressed as:(1) hc,ga is the coefficient of convective heat transfer between the glazing and ambient air. It is expressed by Mc Adams as [63]:(2) VV is the wind speed.

On the other hand, glass is also subjected to a radiative thermal exchange with the sky. This is expressed as:(3)where εg is the glass emissivity, σ is Stephan-Boltzman constant, Tg is the glass temperature and Ts the temperature of sky.

The heat losses between glass and ambient air may be written as:(4)with(5)

The temperature of the sky, which corresponds to the temperature of the celestial vault, is given in terms of the ambient temperature Ta. Several expressions can be used to calculate this temperature. The most widely employed is the one proposed by Swinbank [64]:(6)

4.1.1.2 Thermal losses between the absorber and glass

Convective heat exchange: The heat flux exchanged by convection between the absorber and glazing is written as:(7)

Radiative heat exchange: The heat flux exchange by radiation between the absorber and glazing is written as:(8)with:(9)

The overall coefficient of thermal losses Uav before is then be given by:(10)

4.1.2 Rear thermal losses of the collector

Rear thermal losses are small compared to the front thermal losses because the collector is generally well insulated at the rear. The expression that allows evaluating the heat exchange coefficient is given by:(11)where h is the coefficient of thermal exchange between the ground and the rear side of the collector, and(12)where λis is the thermal conductivity of the insulator and eis is its thickness.

To these losses may be added the lateral thermal losses which are relatively small compared to the rear thermal losses, because the lateral surface of the collector is not important.

4.2 Materials and experimental setup

The operation and operating mode of the two solar collector models are similar. At first, the experiments are carried out in the case of natural convection, where the ambient air is sucked in from the outside (Fig. 7) and passes through the solar collector and consequently heats up; it then leaves the solar collector through its outlet.

In the present case, the suction of air inside the solar collector occurs as a result of the temperature difference between the inlet and the outlet of the collector. Afterwards, the experiments are performed in the case of forced convection. The fluid then follows the same path, except that in this case, it is displaced by the fan which is placed at the outlet of the solar collector. The air speed at the outlet is 1.5 m/s.

A series of experiments were conducted in an open site, away from obstacles in order to avoid the mask effect.

Various measurement campaigns were carried out according to the solar irradiation (during the day) and to the flow rate variation.

The measurements were taken every 30 minutes, between 9 am and 4 pm. This period of time is amply sufficient to take the measurements because, within this time range, the solar irradiation varies considerably. The measurements were performed on Monday 13 April, 2017. Temperatures were measured using Nickel-Chromium/Nickel-Aluminum Type K thermocouples, with an accuracy of 0.1 °C.

These thermocouples have a diameter of 0.5 mm, with a measurement temperature range between −50 °C and 400 °C. About 34 K-type thermocouples were placed within the collector; one thermocouple was positioned at the input of the solar collector and a second one at its output. Another thermocouple was used to read the glazing temperature, another one to measure the ambient temperature. Also, fifteen thermocouples were distributed over the surface of the collector absorber and fifteen others were spread over the air vein inside the solar collector (Fig. 8). The quantification of the thermal field (the distribution of temperature within each of the two studied models of solar collectors) was achieved by means of an infrared camera of the “Fluke TI10” type.

On the other hand, the global solar radiation was measured using a calibrated reference solarometer “Hand Pyranometer 4890.20”. This device was placed on the same plane as the collector with respect to the solar radiation direction, as it is clearly shown in the following figure. The different measured quantities are given in the following table, while specifying the number of measuring points for each quantity.

A given value of the temperature of the absorber or of the air inside the solar collector − presented on the graphs − represents the average temperature calculated on the basis of fifteen local values distributed over the entire surface of the absorber or within the air vein inside the solar collector (Fig. 4).

thumbnail Fig. 7

Diagram of the experimental setup.

thumbnail Fig. 8

Location of the different thermocouples inside the solar collector.

Table 2

Measurements.

5 Results and discussion

5.1 Glazing temperature in the case of free convection

The glazing temperatures for both collector models are given in Figures 9 and 10, respectively, for the case of free convection and forced convection. In Figure 9, it should be noted that the temperature difference is almost zero at the beginning of the measurements, i.e. between 8 h 30  am and 10  am. There is a temperature difference of 06 °C during the time interval between 10:00 and 15:30. Elsewhere, there is no difference in temperature.

thumbnail Fig. 9

Glazing temperature in the case of free convection.

5.2 Glazing temperature in the case of forced convection

Results of comparison of the glazing temperatures of the two collector models under study, in the case of forced convection, are given in Figure 10. It should be noted that the temperature difference is almost zero at the beginning of the measurements (between 8 am and 10 am); it begins to increase until reaching a value of 14.7 °C at 1 pm, after which this difference starts decreasing to finally reach a constant value around 7 °C.

thumbnail Fig. 10

Glazing temperature in the case of forced convection.

5.3 Absorber temperature in the case of free convection

The comparison of the absorber temperatures for the two models of collectors under study in the case of free convection is shown in Figure 11. It can be seen that the temperature difference is almost zero at the beginning of the measurements (between 08.00 and 10.00) and also between 14.30 and 16.00. It is easy to note a temperature difference of 15 °C at 10.30, 26 °C at 11.00 and 12 °C at 12.00; this gives an average of 17.5 °C which is higher than that of the collector with reflective baffles.

thumbnail Fig. 11

Temperature of the absorber (free convection).

5.4 Temperature of the absorber in the case of forced convection

The comparison of the absorber temperatures for the two models of collectors under study in the case of forced convection is illustrated in Figure 12. It is interesting to note that the difference in temperature is almost zero at the beginning of the measurements (8.00 until 10.00). However, a difference of 17 °C is observed between the two collector models between 11.00 and 13.00, and then a difference of 11 °C between 13.00 and 15.30.

thumbnail Fig. 12

Temperature of the absorber (forced convection).

5.5 Air temperature inside the solar collector in the case of free convection

The comparison of air temperatures inside the solar collector in the case of free convection for the two collector models under study is illustrated in Figure 13. It can be noted that the temperature difference is negligible at the beginning of the measurements (from 8.00 am to 10.00 am); then, this difference rises to 15 °C between the two collector models during the rest of the measurement time.

thumbnail Fig. 13

Temperature of air (free convection).

5.6 Air temperature inside the collector in the case of forced convection

The comparison of air temperatures inside the solar collector in the case of forced convection is shown in Figure 14. It can be seen that the temperature difference tends to zero at the beginning of the measurements (from 8 am to 10 am). This difference becomes equal to 22.5 °C during the rest of the measurement time.

thumbnail Fig. 14

Air temperature (forced convection).

5.7 Air temperature at the collector outlet in the case of free convection

The comparison of air temperatures at the outlet of the collector in the case of free convection for the two collector models under study is illustrated in Figure 15. It can be seen that the temperature difference is almost zero at the beginning of the measurements (between 8.00 and 10.00) and between 13.00 and 16.00 as well; this temperature difference is equal to 10 °C in the time interval between 10.00 am and 1 pm.

thumbnail Fig. 15

Air temperature at the solar collector outlet (free convection).

5.8 Air temperature at the collector outlet in the case of forced convection

The comparison of air temperature at the outlet of the two collector models in the case of forced convection, for both collector models, is shown in Figure 16. It can be seen that the temperature difference is almost zero only at the beginning of the measurements (between 8.00 and 10.00); after that, this difference is 13 °C between the two collector models, during the entire period of the day.

thumbnail Fig. 16

Air temperature at the solar collector outlet (forced convection).

5.9 Evolution of the different temperatures for MSCMR and BMSC collectors, in the case of free convection

In order to better estimate the forward losses, the different temperatures (glazing temperature, absorber temperature, air temperature inside the collector and finally the exit air temperature) were grouped for the same type of collector (MSCMR or BMSC) and for the same kind of convection (free or forced).

The evolution of the different temperatures for the MSCMR collector, in the case of free convection, is given in Figure 17. There is a temperature difference between the absorber and the glazing, and between the outlet temperature of air and that of the glazing, respectively, of 44 °C, with a rate of 44%, and 28 °C, with a rate of 33%. Similarly, the evolution of the different temperatures for the BMSC collector, in the case of free convection, is given in Figure 18. There is a difference in temperature between the absorber and the glazing, and between the temperature of air at the outlet of 44 °C, with a rate of 44%, and 31 °C, with a rate of 38%.

A comparison between the different rates given above allows us to say that the forward losses (which strongly depend on the absorber and glazing temperatures) for the two collectors MSCMR and BMSC, and for the case of convection natural, are almost identical (44% and 33% for MSCMR, against 47% and 38% for BMSC).

thumbnail Fig. 17

Different temperatures for collector MSCMR (free convection).

thumbnail Fig. 18

Different temperatures for collector BMSC (free convection).

5.10 Evolution of the different temperatures for MSCMR and BMSC collectors, in the case of forced convection

The evolution of the different temperatures for the MSCMR collector, in the case of forced convection, is given in Figure 19. There is a temperature difference between the absorber and the glazing and between air temperature of the outlet and the glazing, respectively, of 37 °C, a rate of 39%, and 21 °C, a rate of 27%. Similarly, the evolution of the different temperatures for the BMSC collector, in the case of forced convection, is given in Figure 20. There is a difference in temperature between the absorber and the glazing and between the outlet temperature of air and glazing, respectively, of 44 °C, with a rate of 46%, and 30 °C, with a rate of 37%.

A comparison between the different rates given above allows us to say that forward losses (case of forced convection) for the MSCMR collector have dropped considerably as compared to those of the BMSC collector (39% and 27% for MSCMR against 44% and 37% for BMSC).

thumbnail Fig. 19

Different temperatures for collector MSCMR (forced convection).

thumbnail Fig. 20

Different temperatures for collector BMSC (forced convection).

5.11 Qualitative comparison of the two models of solar collectors

The quantification of the thermal field (distribution of temperature inside each of the two collectors under study) was recorded by means of a “Fluke TI10” type infrared camera. Figures 21 and 22 illustrate, respectively, the temperature distribution within the solar collector with reflective baffles as well as that in the solar collector with mini-concentrators. It is clear that the MSCMR collector model has a better temperature distribution than the BMSC solar collector, especially at the top, i.e. in the area near the output of the collector.

Temperatures are high, especially on the left side of the collector, which means that the collector must be oriented in the right-to-left direction.

thumbnail Fig. 21

Distribution of air temperature inside the collector with BMSC (forced convection).

thumbnail Fig. 22

Distribution of air temperature inside the collector with MSCMR (forced convection).

6 Discussion of results

The different temperatures recorded indicate that the values found for the MSCMR collector are much better than those of the BMSC collector. It should be noted that there was an increase in the outlet temperature of 10% over the entire measurement time interval in the case of forced convection. In addition, an almost identical increase in the outlet temperature was observed in the case of free convection, but only in the time interval between 10 and 13 h.

The different temperatures recorded for the MSCMR collector remained almost constant and homogeneous over a long period. However, these temperatures dropped sharply in the case of the BMSC collector when solar irradiation decreased, which means that the MSCMR collector accumulates more heat than the BMSC collector. This is certainly due to the thermal inertia of the MSCMRs due to their proper weight (176 × 20 g = 3520 gram for MSCMR against 176 gram only for BMSC).

It is worth noting that the MSCMR collector causes a 10% increase in air temperature at the outlet of the collector as compared to that of the collector BMSC; this is true for the 2 cases of free and forced convection. In addition, the collectors BMSC and MSCMR lead to an air temperature increase at the collector outlet equal to 24% and 27%, respectively, compared to that given by a simple solar flat-plate air collector, in the case of free convection. Similarly, the BMSC and MSCMR collectors lead to an air temperature increase at the collector outlet, estimated at 28% and 35%, respectively, with respect to that of a simple solar flat-plate air collector, in the case of forced convection. These findings may be explained by the fact that the addition of BMSCs within the absorber leads to a higher convective heat exchange surface; this additional surface represents 53% of the total surface area of the absorber. However, in addition to the exchange surface provided by the heat transfer fluid and the absorber, there is another heat exchange surface that is composed of several small surfaces in the direction of the air flow. Consequently, the convective heat exchange coefficient between the absorber supplied with BMSCs and the heat transfer fluid. In this case, the value hc,pg will be greater than that of the absorber without BMSCs due to the additional heat exchange surface provided by the BMSCs. Moreover, the presence of BMSCs leads to the creation of turbulence, which strongly favors heat exchange.

Furthermore, placing MSCMRs within the collector does not only increase the heat exchange surface between air and the absorber, but generates turbulence as well. Note that the mirrors of the MSCMRs contribute to the concentration of radiation incident on the absorber and consequently leads to increased temperature.

The temperature of the MSCMR collector glazing was on average higher than that of the BMSC collector by approximately 14.7 °C. In addition, the MSCMR collector losses were high, which caused the temperature at its output to considerably decline. Consequently, this led to lower performance. Note that the double glazing could be a solution to remedy this problem.

7 Conclusion

The present study aimed to make a comparative experimental study between two models of flat plate solar collectors, namely BMSC and MSCMR. This comparison was made in both cases of convection, i.e. free convection and forced convection. The comparison between the two models of collectors under study was carried out based on the calculations of average temperatures during the measurement periods, for each model. The relevant temperature profiles were obtained by means of the method used in estimating the average temperature of the absorber and also the average temperature of the air stream inside the collector. These temperatures were estimated based on an average of fifteen temperatures spread over the entire surface of the absorber and over the total air vein inside the collector.

The different temperatures recorded indicate that those of the MSCMR collector are much higher than those of the BMSC collector. Indeed, there is an 10% increase for the outlet temperature, in the case of forced convection over the entire measurement time interval. Similarly, there is also almost the same increase in the rate of the outlet temperature in the case of free convection but only during the time interval between 10 h and 13 h.

The different temperatures relating to the MSCMR collector remained almost constant and homogeneous for a long time. However, for the BMSC collector, these temperatures dropped sharply once solar irradiation began to regress. This should be very important for industrial applications such as food drying where the outlet temperature must be constant.

A comparison between the different temperatures of the same collector (MSCMR or BMSC) and for the same kind of convection (free or forced) allows us to say that the front losses, which strongly depend on the temperatures of the absorber and the glazing, for the MSCMR and BMSC collectors and for the case of natural convection are almost the same; they are 44% and 33% for collector MSCMR against 47% and 38% for collector BMSC. However, in the case of forced convection, these rates dropped significantly for the MSCMR collector compared to BMSC, with 39% and 27% for MSCMR, and 44% and 37% for BMSC. Double glazing could be a solution to remedy this temperature increase of the glazing. Despite this, the outlet temperature remained very important especially for collector MSCMR; it is of the order of 110 °C for natural convection and 100 °C for forced convection, while the ambient temperature is about 30 °C.

Furthermore, the infrared camera allowed quantifying the temperature distribution in each collector, namely BMSC and MSCMR. It turned out that the MSCMR collector had a better temperature distribution. This temperature distribution was particularly high at the outlet of the MSCMR collector.

Nomenclature

Latin notation

eis : Insulator thickness (M)

hc,ga : Coefficient of convection heat exchange between glazing and ambient air (W/m2 K)

hr,ga : Coefficient of radiation heat exchange between glazing and ambient air (W/m2 K)

hc,pg : Coefficient of convection heat exchange between absorber and glazing (W/m2 K)

hr,pg : Coefficient of radiation heat exchange between absorber and glazing (W/m2 K)

hd : Coefficient of heat exchange between ground and rear side of collector (W/m2 K)

K : Thermal conductivity of insulator (W/m2 K)

Qc,ga : Convective heat flux exchange between glazing and ambient air (W)

Qr,ga : Radiative heat flux exchange between glazing and ambient air (W)

Qc,pg : Convective heat flux exchange between absorber and glazing (W)

Qr,pg : Radiative heat flux exchange between absorber and glazing (W)

Ta : Ambient temperature (K)

Tg : Glazing temperature (K)

Tp : Absorber temperature (K)

Ts : Temperature of sky (K)

VV : Speed of wind (m/s)

Uav : Overall heat loss coefficient before (W/m2 K)

Uarr : Overall heat loss coefficient after (W/m2 K)

Greek notation

λis : Thermal conductivity of insulator (W/m K)

εg : Emissivity of glazing (–)

εp : Emissivity of absorber (–)

σ : Stefan-Boltzmann constant (W/m2 K4)

Abbreviations

A: Ambient

Av: Before

Arr: After

C: Convection

G: Coating

Is: Insulator

P: Absorber

R: Radiation

S: Sky

BMSC: Black mini solar concentrator

MSCMR: Mini solar concentrator with mirror

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Cite this article as: Zakaria Sari Hassoun, Khaled Aliane, Comparative study of two types of mini chicanes concentrators implanted in solar collector, Eur. Phys. J. Appl. Phys. 87, 30902 (2019)

All Tables

Table 1

Characteristics of solar collector component.

Table 2

Measurements.

All Figures

thumbnail Fig. 1

(a) Solar collector under study (without mini concentrators). (b) Fan at the exit of the collector.

In the text
thumbnail Fig. 2

(a) Solar collector with black mini solar concentrators (BMSCs). (b) Solar collector with mini solar concentrators and mirrors (MSCMRs).

In the text
thumbnail Fig. 3

Dimensions of BMSCs.

In the text
thumbnail Fig. 4

Arrangement of BMSCs within the absorber of the solar collector.

In the text
thumbnail Fig. 5

Pitch separating two consecutive BMSCs.

In the text
thumbnail Fig. 6

Typical mini-concentrators.

In the text
thumbnail Fig. 7

Diagram of the experimental setup.

In the text
thumbnail Fig. 8

Location of the different thermocouples inside the solar collector.

In the text
thumbnail Fig. 9

Glazing temperature in the case of free convection.

In the text
thumbnail Fig. 10

Glazing temperature in the case of forced convection.

In the text
thumbnail Fig. 11

Temperature of the absorber (free convection).

In the text
thumbnail Fig. 12

Temperature of the absorber (forced convection).

In the text
thumbnail Fig. 13

Temperature of air (free convection).

In the text
thumbnail Fig. 14

Air temperature (forced convection).

In the text
thumbnail Fig. 15

Air temperature at the solar collector outlet (free convection).

In the text
thumbnail Fig. 16

Air temperature at the solar collector outlet (forced convection).

In the text
thumbnail Fig. 17

Different temperatures for collector MSCMR (free convection).

In the text
thumbnail Fig. 18

Different temperatures for collector BMSC (free convection).

In the text
thumbnail Fig. 19

Different temperatures for collector MSCMR (forced convection).

In the text
thumbnail Fig. 20

Different temperatures for collector BMSC (forced convection).

In the text
thumbnail Fig. 21

Distribution of air temperature inside the collector with BMSC (forced convection).

In the text
thumbnail Fig. 22

Distribution of air temperature inside the collector with MSCMR (forced convection).

In the text

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