Experimental study of heat transfer coefficient with rectangular baffle fin of solar air heater

Abstract

This paper presents an experimental analysis of a single pass solar air collector with, and without using baffle fin. The heat transfer coefficient between the absorber plate and air can be considerably increased by using artificial roughness on the bottom plate and under the absorber plate of a solar air heater duct. An experimental study has been conducted to investigate the effect of roughness and operating parameters on heat transfer. The investigation has covered the range of Reynolds number Re from 1259 to 2517 depending on types of the configuration of the solar collectors. Based on the experimental data, values of Nusselt number Nu have been determined for different values of configurations and operating parameters. To determine the enhancement in heat transfer and increment in thermal efficiency, the values of Nusselt have been compared with those of smooth duct under similar flow conditions.

Keywords

Nusselt number / flow rate / heat transfer / heat transfer coefficient / thermal efficiency / forced convection

Introduction

Several investigators studied forced convection heat transfer in smooth and roughened ducts, and much useful information is available in the literature. The use of artificial roughness on a surface is an effective technique to enhance heat transfer to fluid flowing in ducts. In this study, two modes of the baffle fin are tested: under and on an absorber plate.

The absorber plate and bottom plate of solar air heaters having transversal baffle with rectangular form on one side of the plate (flow side) give better thermal performance than smooth plate. Artificial roughness under the absorber plate and on the bottom plate has been used to create turbulence near the wall or to break the boundary layer. Thus, the artificial roughness can be employed for the enhancement of heat transfer coefficient between the absorber plate and air, thereby improving the thermal performance of solar air heaters (SAHs).

Solar air collector is an effective device to harness solar radiation for space heating and other purposes. Studies show that the efficiency of solar air collector can be improved by designing the baffle fin. The solar air collectors, because of their simple construction and low cost, are extensively used in the world for heating purposes. In this study, a test of solar air collector was performed based on the heating of air by different configurations of the roughness, and the surface area for heat exchange. To this end, a traversal baffle fin with rectangular form is one of the important and attractive design improvements proposed to improve thermal performance.

There is an extensive literature dealing with boundary layer flow and heat transfer of a flat plate oriented at an angle of attack to an oncoming stream. Yeh and Lin [1] investigated theoretically and experimentally the effect of parallel barriers on the collector efficiency of flat-plate SAHs and discovered that the collector efficiency increases theoretically as the number of barriers increases. Sparrow and Tien [2] experimentally investigated the forced convection heat transfer at an inclined and yawed square plate. The common practice in predicting performance of solar energy system is to solve a set of several inter-related steady-state heat balance equations representing various components [3]. The convective heat transfer coefficient, h (absorber plate to flowing air) is an important parameter required for mathematical modeling, computer simulation and performance prediction [4]. Many researchers, while predicting the performance of solar dryers systems, are constrained to utilize convective heat transfer correlations published for standard geometries of system and specific test conditions [5–7]. Comparison of results reveals that the thermal efficiency of a single pass is higher with an increase in the flow rate. Increasing the absorber area or fluid flow heat-transfer area will increase the heat transfer to the flowing air and the pressure drop in the collector, thereby increasing the required power consumption to pump the air flow crossing the collector [8,9].

On the other hand, several configurations of absorber plates have been designed to improve the heat transfer coefficient. Artificial roughness obstacles and baffles in various shapes and arrangements were employed to increase the area of the absorber plate. As a result, the heat transfer coefficient between the absorber plate and the air pass is improved [10]. Omojaro and Aldabbagh [11] conducted an experimental investigation of the thermal performance of a single and double pass SAH with fins attached and a steel wire mesh as absorber plate, in which the bed heights were made to be 7 cm and 3 cm for the lower and upper channels, respectively. The result of a single or double SAH, when compared with conventional SAH, shows much more substantial enhancement in thermal efficiency.

Paisarn [12] numerically studied the performance and entropy generation of the double-pass flat plate SAH with longitudinal fins. Predictions were made at air mass flow rate ranging between 0.02 kg/s and 0.1 kg/s. The fins served as heat transfer augmentation features in SAHs; however, they increased the pressure drop in flow channels. The results showed that high efficiency of the optimized fin improved the heat absorption and dissipation potential of a SAH [13]. The double flow SAH was designed with fins attached over and under the absorbing plate. This resulted in a considerable improvement in collector efficiency of double flow SAHs with fins compared to single flowing, operating at the same flow rate [14]. An experimental investigation was carried out on the thermal performance of the offset rectangular plate fin absorber plates with various glazing [15], in which, the offset rectangular plate fins which were used in heat exchangers, were experimentally studied. Because the offset rectangular plate fins were mounted in staggered pattern and oriented parallel to the fluid flow, high thermal performances were obtained with low-pressure losses. Karim and Hawlader [16] conducted experiments to study the performance of three types of SAH, namely flat plate, finned and V-corrugated SAHs.

The V-corrugated collector was found to be the most efficient while the flat plate collector the least efficient. Another work used the cross-corrugated absorbing plate and bottom plate to enhance the turbulence and the heat transfer rate inside the air flow channel and tested its thermal performance [17,18]. A novel solar air collector of pin-fin integrated absorber was designed to increase the thermal efficiency [19]; in the performance analysis of varying flow rates on PZ7-11.25 pin-fin arrays collector, the correlation equation for the heat transfer coefficient was obtained and the efficiency variation vs. air flow rate was determined. In another work, the results were compared with those obtained with a solar air collector without fins, using two types of absorbers selective (in copper sun) or not selective (black-painted aluminum) [20].

Andoh et al. [21] presented a solar water heater designed with a local vegetable material as insulating material. They focused on the comparative thermal performance of this collector and another collector, identical in design, fabrication, and operating under the same conditions, using glass wool as heat insulation. Chabane et al. [22–27] studied the effect of the mass flow rate in the range from 0.0078 kg/s to 0.0166 kg/s on the solar collector with longitudinal fins. The flat-plate SAH [28–32] was considered to be a simple device consisting of one (transparent) cover situated above an absorbing plate with the air flowing under the absorber plate [31,32]. The conventional flat-plate SAH was investigated for heat-transfer efficiency improvement by introducing forced convection [33,34] extended heat-transfer area [35,36] and increase of air turbulence [37,38].

The value of a convection heat transfer coefficient depends upon the physical configuration as well as upon several properties of the fluid occupied. Empirical correlations are available to estimate heat transfer coefficients for a variety of forced convection heat transfer configurations and will be presented and discussed in this study.

This paper presented the results of an experimental investigation of the performance of a novel flat plate SAH with several obstacles (Type I, Type II, and Type III) and without obstacles (Type IV). It was found that the optimal value of efficiency was determined by the SAH with Type II absorbent plate in flow channel duct for all operating conditions and the collector supplied with obstacles appeared significantly better than that without obstacles [39,40].

Experimental

Thermal analysis and uncertainty

Heat transfer coefficient

The convective heat transfer coefficient h_w for air flowing over the outside surface of the glass cover depends primarily on the wind velocity V_wind. McAdams [41] obtained the coefficient from the experiment as

(1)h_{\mathrm{w}}=\mathrm{5.7}+\mathrm{3.8}{V}_{\mathrm{w}\mathrm{i}\mathrm{n}\mathrm{d}},

where the units of h_w and V_wind are W/(m²·K) and m/s, respectively. An empirical equation for the loss coefficient from the top of the solar collector to the ambient was developed by Klein [42]. The heat transfers coefficient between the absorber plate and the airstream is always low, resulting in low thermal efficiency of the SAH. Increasing the area of the absorber plate shape will increase the heat transferred to the following air.

Collector thermal efficiency

The efficiency of a solar collector is the ratio of the amount of useful heat collected to the total amount of solar radiation striking the collector surface during any period of time [43–45]:

(2)\eta ={\displaystyle \frac{\mathrm{Solar}\mathrm{}\mathrm{energy}\mathrm{}\mathrm{collected}}{\mathrm{Total}\mathrm{}\mathrm{solar}\mathrm{}\mathrm{striking}\mathrm{}\mathrm{collector}\mathrm{}\mathrm{surface}}}={\displaystyle \frac{Q_{\mathrm{u}}}{I\times A_{\mathrm{c}}}}.

where A_c is the area of the collector, I is global irradiance incident on SAH collector, and the useful heat collected for an air-type solar collector can be expressed as

(3)Q_{\mathrm{u}}=\stackrel{.}{m}{C}_{\mathrm{p}}\left(T_{\mathrm{out}}-T_{\mathrm{in}}\right),

where C_p is the specific heat of the air. The equation for mass flow rate (m) is

m=\rho \times Q_{\mathrm{v}},

where ρ is the density of air, which depends on the air temperature and Q_v is the volume flow rate which depends on the pressure difference at the orifice measured from the inclined tube manometer and temperature.

The fractional uncertainty of the efficiency from Eq. (3) is a function of ΔT, m, and I, considering C_p and A_c as constants. So, collector thermal efficiency becomes

(4)\eta ={\displaystyle \frac{m{C}_{\mathrm{p}}\left(T_{\mathrm{out}}-T_{\mathrm{in}}\right)}{I{A}_{\mathrm{c}}}}.

Thermal performance of conventional SAHs

Figure 1 illustrates the thermal network of a conventional smooth plate SAH. The thermal performance of the flat plate SAH could be observed by considering the energy balance between the solar energy absorbed by the absorber plate and the useful thermal energy output of the system accompanied by some losses. The energy balance equation could be written as

(5)Q_{\mathrm{a}}=A_{\mathrm{c}}\left[IR\left(\tau \alpha \right)\right]=Q_{\mathrm{u}}+Q_{\mathrm{l}},

where Q_a is the energy absorbed by the absorber plate, A_c is the area of the absorber plate, I is the intensity of insolation, R is the conversion factor to convert radiation on the horizontal surface to that on the absorber plane, τα is the effective transmittance absorptance product of the glass cover-absorber plate combination, Q_u is the useful energy gain, and Q_l is the energy loss from the collector.

The useful energy gain can be expressed in terms of inlet air temperature T_in and other system and operating parameters as

(6)Q_{\mathrm{t}}=A_{\mathrm{c}}{F}_{\mathrm{R}}\left[IR\left(\tau \alpha \right)-U_{\mathrm{l}}\left(T_{\mathrm{in}}-T_{\mathrm{a}}\right)\right],

(7)F_{\mathrm{R}}={\displaystyle \frac{\stackrel{·}{m}{C}_{\mathrm{p}}}{A_{\mathrm{c}}{U}_{\mathrm{l}}}}\left[\mathrm{1}-\exp \left({\displaystyle \frac{F^{\prime }{U}_{\mathrm{l}}{A}_{\mathrm{c}}}{m{C}_{\mathrm{p}}}}\right)\right],

where F_R is the heat removal factor of the collector, which indicates the thermal resistance is met by the absorbed solar energy in reaching the flowing air; U_l is the overall loss coefficient; T_in and T_a are the inlet air and ambient temperatures, respectively; F´ is termed as the efficiency factor of the collector, which provides the relative measurement of the thermal resistance between the absorber plate and the ambient air to that of the thermal resistance between the air flowing through the collector and the ambient air.

The collector efficiency factor (F´) is expressed as

(8)F^{\prime }={\displaystyle \frac{\mathrm{1}}{\left(\mathrm{1}+U_{\mathrm{l}}/h_{\mathrm{e}}\right)}},

where h_e is the effective heat transfer coefficient between the absorber plate and flowing air. The thermal efficiency of the collector is the ratio of useful heat gain to the incident solar energy falling on the collector.

Therefore,

(9)\eta ={\displaystyle \frac{Q_{\mathrm{u}}}{I{A}_{\mathrm{c}}}}=\left[\tau \alpha -{\displaystyle \frac{U_{\mathrm{l}}\left(T_{\mathrm{in}}-T_{\mathrm{a}}\right)}{I}}\right].

According to Eq. (9), the thermal efficiency of the solar collector could be improved by increasing the value of F_R which depends on the efficiency factor F´ of the collector. By enhancing the heat transfer coefficient between the absorber plate and air, higher values of F´ could be achieved. The roughening of the absorber surface has been found to be a convenient and effective technique to enhance the convective heat transfer rates from the absorber surface to the air.

The outlet temperature of air in terms of F_R is

(10)T_{\mathrm{out}}=T_{\mathrm{in}}+F_{\mathrm{R}}{\displaystyle \frac{\left[I\left(\tau \alpha \right)-U_{\mathrm{L}}\left(T_{\mathrm{in}}-T_{\mathrm{a}}\right)\right]}{m{C}_{\mathrm{p}}}}.

The thermo-physical properties of air have been taken at the corresponding mean temperature T_f from the following relations of thermo-physical properties, obtained by correlating data from NBS (US) [46]:

C_{\mathrm{p}}=\mathrm{1006}{\left({\displaystyle \frac{T_{\mathrm{f}}}{\mathrm{293}}}\right)}^{\mathrm{0.0155}},k=\mathrm{0.0257}{\left({\displaystyle \frac{T_{\mathrm{f}}}{\mathrm{293}}}\right)}^{\mathrm{0.86}},

(11)\mu =\mathrm{1.81}\times {\mathrm{10}}^{-\mathrm{5}}{\left({\displaystyle \frac{T_{\mathrm{f}}}{\mathrm{293}}}\right)}^{\mathrm{0.735}},\rho =\mathrm{1.204}\left({\displaystyle \frac{\mathrm{293}}{T_{\mathrm{f}}}}\right).

Equations used for calculation

The following equations were used for calculating the mass flow rate of air, m, heat energy transfer, Q_u, mean’s heat transfer coefficient of fluid, h [47].

(12)m=Cd{\left[{\displaystyle \frac{\mathrm{2}\rho \left(\delta p_{\mathrm{0}}\right)}{\left(\mathrm{1}-{\beta }^{\mathrm{4}}\right)}}\right]}^{\mathrm{0.5}},

(13)h={\displaystyle \frac{Q_{\mathrm{u}}}{A_{\mathrm{c}}\left(T_{\mathrm{p}}-T_{\mathrm{f}}\right)}},

where T_p and T_f are average values of the absorber plate temperature and air temperature, respectively. The average temperature of the plate is determined by the temperature recorded at three different locations along the test section of the absorber plate. It is found that the temperature of the absorber plate varies predominantly in the flow direction only and is linear. The air temperature is determined as an average of the temperatures measured at three central locations over the test length of the duct along the flow direction. The Nusselt number is calculated by

(14)Nu={\displaystyle \frac{h{D}_{\mathrm{h}}}{k}}.

The Prandtl number, a dimensionless number approximating the ratio of momentum diffusivity (kinematic viscosity) and thermal diffusivity, can be expressed as

(15)Pr={\displaystyle \frac{C_{\mathrm{p}}\mu }{k}}.

Description of SAH considered in this work

A schematic view of the constructed single flow under an absorber plate in duct of solar collector is shown in Fig. 2, and the schematic of a solar collector and the view of the absorber plate in the collector box are demonstrated in Fig.3. In this study, three configurations of the absorber plates were used with baffle fin fixed on the absorber plate and on the bottom plate. The absorbers were made of galvanized iron sheet with black chrome selective coating. The plate thickness of three collectors was 0.5 mm. The cover glass was 5 mm in thickness. Single transparent cover was used for the three collectors. Thermal losses through the collector backs were mainly caused by the conduction across the insulation (thickness 4 cm), and those caused by the wind and the thermal radiation of the insulation were assumed to be negligible. After installation, the two collectors were left operating for several days under normal weather conditions for weathering processes.

The thermocouples were positioned evenly on the top surface of the absorber plates, at identical positions along the direction of flow for both collectors. The inlet and outlet air temperature were measured by two well insulated thermocouples. The output from the thermocouples was recorded in degrees Celsius by using a digital thermocouple thermometer DM6802B, with a measurement range from -50°C to 1300°C (-58°F to 1999°F), a resolution of 1°C or 1°F, an accuracy of ±2.2°C or ±0.75% of reading, and by using a non-contact digital infrared thermometer temperature laser gun TM330, with an accuracy of ±1.5°C/±1.5%, a measurement range from -50°C to 330°C (-58°F to 626°F), a resolution of 0.1°C or 0.1°F, and an emissivity of 0.95. A digital thermometer measured the ambient temperature with sensor in display LCD CCTV-PM0143 placed in a special container behind the collectors. The total solar radiation incident on the surface of the collector was measured with a Kipp and Zonen CMP 3 Pyran-ometer placed adjacent to the glazing cover, at the same plane, facing due south. The measured variables were recorded at intervals of 15 min, including insolation, inlet and outlet temperatures of the working fluid circulating through the collectors, ambient temperature, absorber plate temperatures at several selected locations, and volume flow rates (Lutron AM-4206M digital anemometer). All tests began at 9:00 am and ended at 4:00 pm.

The layout of the solar air collector studied is depicted in Figs. 2 and 4. The collector A served as the baseline one, whose solar collecting area was 1.95 m (length) × 0.89 m (width); installation angle was 34.8° from horizontal; height of the stagnant air layer was 0.025 m; thermal insulation board EPS (expanded polystyrene board), with a thermal conductivity of 0.037 W/(m·K), was put on the exterior surfaces of the back, and side plates, with a thickness of 40 mm; the baffle fin were 0.89 m in length, 1.5 cm in height and located in a direction 4/5 of length the solar collector, as displayed in Fig. 4.

To improve the thermal performance of the solar collector and to choose the best position of the baffle plate located on the bottom or under the absorber, very thin metal plate was introduced into the airstream

α_abs = 0.95, absorptivity coefficient of the absorber;

α_v = 0.06, absorptivity coefficient of the glass;

ϵ_pl = 0.25, emissivity of the back plate;

ϵ_abs = 0.95, emissivity of the absorber painted matt black;

ϵ_b = 0.93, emissivity glazing;

τ_v = 0.84, transmissivity coefficient of the glazing.

Results

The objective of this research was to study experimentally the flow forced convection heat transfer of the air flowing through offset plates located between parallel plates and heated by radiation heat flux. An outdoor test facility was designed and fabricated to generate heat transfer data for the flow in a rectangular duct with baffle fin surface at different volume flow rates for a range of surface roughness.

The roughness was controlled by different solar collectors. Experimental data were collected for smooth duct for ensuring the accuracy of the experimental data and also to compare the heat transfer characteristics of roughened and smooth duct. Two plates having different position of the baffle fin were tested for different flow rates. Three flow rates corresponding to a flow Reynolds number of 1259, 1888 and 2517 were used for each test set and data were collected under steady-state condition.

Based on this experimental investigation, it is found that the transversal baffle fin in the rectangular duct has a better thermal efficiency and heat exchange than in the smooth plate, when the heat transfer coefficient is increased by a rate of 4.62 and 13.48, corresponding to the volume flow rate of 80 m³/h in the last position of length collectors fold increase in the thermal efficiency as compared to a smooth duct for the range of present investigation. It can be seen from Fig. 5 that the performance is significantly improved relative to the solar collector with a smooth plate. Besides, it can be noted that a 53% yield can be achieved by a baffled solar collector at a volume flow rate of 60 m³/h compared with the smooth plate at a volume flow rate of 80 m³/h.

It is also noted that the obtained with a baffle fixed on the bottom plate is slightly greater than that of a solar collector with the baffle fixed under the absorber plate.

Discussion

Figure 6 shows the average temperature distribution in the solar collector and the variation of the average temperature corresponding to the absorber plate. The difference can be seen in Fig. 6 (a), (b), (c) as a function to volume flow rates of 40 m³/h, 60 m³/h and 80 m³/h, respectively. This difference is caused by the configurations used under the weather condition and the position of the baffle in the duct, which cools the absorber plate and ensures a better heat exchange with the air. Table 1 lists the average temperature of the absorber plate for each solar collector configuration, corresponding to the volume flow rates of 40 m³/h, 60 m³/h and 80 m³/h, and the solar collector length from 0 m to 1.95 m with a tilt angle β = 34.8°.

It can be seen from Table 1 that the average temperature of the absorber plate for each solar collector configuration, corresponding to the volume flow rates of 40 m³/h, 60 m³/h and 80 m³/h, increases the volume flow rate, which impacts directly the temperature under an absorber plate and on the bottom plate. The temperature of the absorber plate corresponding to the volume flow rates of 40 m³/h, 60 m³/h and 80 m³/h in the median length of the solar collectors depending on configuration A, B and C, are (T_ab = 60°C, 70°C and 70°C), (T_ab = 73°C, 71°C and 65°C) and (T_ab =80°C, 72°C and 65°C), respectively. The change in volume flow rate has a significant influence on the temperature of the absorber plate and results in an increase of 13°C and 20°C for each solar collector when the flow rate is increased from 40 m³/h to 80 m³/h (Reynolds number ranging from 1259 to 2517). The present study indicates that the changes in temperatures and volume flow rates can have an effect on heat transfer for each design of SAH.

Figure 7 presents the average air temperature of a fluid as a function to the length from 0 m, 0.975 m and 1.95 m, corresponding to three configurations at volume flow rates of 40 m³/h, 60 m³/h and 80 m³/h. It can be seen that the curves of the temperature fluid in x₂= 0.975 m as a function to the volume flow rates of (T_f = 58°C, 58°C and 51°C) corresponding to smooth plate and (T_f = 64°C, 59°C and 57°C) of a solar collector with the baffle fin fixed under the absorber plate, and (T_f = 59°C, 55°C and 56°C) corresponding to solar collector with the baffle fin fixed on the bottom plate. Table 2 shows the average air temperature of the absorber plate for each solar collector configuration, corresponding to the volume flow rates of 40 m³/h, 60 m³/h and 80 m³/h, and the solar collector length from 0 m to 1.95 m with a tilt angle β = 34.8°. It is seen from Table 2 that the average temperature of the bottom plate is higher than that of the absorber plate, which means that the fluid between the bottom plate and the absorber plate takes the heat away from the absorber plate. It can be concluded that the temperature fluid in the duct takes away more heat from the absorber plate in configuration C when the baffle fin is fixed under the absorber plate.

Figure 7 demonstrates the air temperature as a function to the length of the solar collector from x₁ to x₃, that is, T_ab(x₁)<(T_ab(x₂), T_ab(x₃))<T_ab(x₂), corresponding to the configuration of the solar collectors. It is observed clearly from Fig. 7 that the fluid takes away some heat from each location a length of a solar collector exceptional in a point x₂. It can be concluded that the increase in flow rates can affected the average temperature of an absorber plate, exceptional in x₂, about this location x₂. It can be seen that the average temperature of the absorber plate takes the maximum value for each configuration used.

The difference among (a), (b) and (c) in Fig. 7 results from fixing the baffle under an absorber plate and on the bottom plate, for the best thermal performance of SAH. The air is distributed very well and takes away more heat from the bottom plate and the absorber plate.

Figure 8 presents the relationship between the local wall heat transfer coefficient and the configurations of the solar collectors, flow rates and length of solar collectors. The experimental results show that the local wall heat transfer coefficient is strongly dependent on the mean length, the flow rates, and the position of the baffle fin in the ranges of the parameter variations studied. The local wall heat transfer coefficient increases rapidly when the flow rates increases, because the characteristic length increases. It is seen apparently that the local wall heat transfer coefficient increases with the flow rate. The heat transfer coefficient for each solar collector configuration, corresponding to the volume flow rates of 40 m³/h, 60 m³/h and 80 m³/h, and solar collector length from 0 m to 1.95 m with a tilt angle β = 34.8° is presented in Table 3. It can be seen that the local wall heat transfer coefficient slightly varies from the smooth plate, and other configuration is very well exceptional in the solar collector with the baffle fixed under an absorber plate because the area of heat exchanges is augment when the flow rate is in the range of 40 m³/h–80 m³/h. From Fig. 8, it can be seen that the relationships is nonlinear between the wall heat transfer coefficient, the length and flow rates. A significant improvement in performance can be noticed relative to the solar collector with the smooth plate. Besides, a 53% yield is achieved by a baffled solar collector at a flow rate of 60 m³/h compared to the smooth plate at a flow rate of 80 m³/h. It is also noted that the obtained by fixing the baffle on the bottom plate is slightly greater than that of the solar collector with the baffle fixed under the absorber plate.

Figure 5 shows the variation of the thermal efficiency with volume flow rates. The thermal efficiency used to evaluate the performance of the SAH is calculated. It can be seen from Fig. 5 that the thermal efficiency increases with the increase in flow rate as a function of the modes of the configuration used in the solar collectors. The efficiencies of the solar collectors with baffle are higher than that of the collector with smooth plate. Figure 5 shows the comparison of the thermal efficiency for three difference regimes of volume flow rates for each solar collector with, and without using baffle fin. The results of each SAH are tabulated in Table 3.

Evidently the mean highest thermal efficiency (η = 58%) at a solar intensity I = 937 W/m² corresponds to the baffle fixed under the absorber plate. The solar collector with the baffle fixed on the bottom plate (η = 56.7%) corresponds to the solar intensity, I = 887 W/m² at the same volume flow rate 80 m³/h and 34.8° tilt angle. The day 05/12/2010, corresponding to smooth plate with a solar irradiation and a thermal efficiency equal to (I = 766 W/m², η = 56.8%), respectively, represents the median highest thermal efficiency. The mean lowest thermal efficiency of the baffle under the absorber plate, on the bottom plate and smooth plate equaling (η = 32.8%, 34% and 35.4%), respectively, corresponds to the volume flow rate 40 m³/h.

The performance curves of three modes of the solar air collectors were tested and are shown in Fig. 8, based on the performance curves at a tilt angle of 34.8° [22]. The thermal efficiency with the baffle fixed under the absorber plate is higher than with the baffle fixed on the bottom plate and smooth plate corresponding to volume flow rates. The SAH heated the air much more at the lower air flow rate because the air had more time to get hot inside the collector.

Figure 9 shows the variation of the ambient, outlet and inlet temperature. The average temperatures of the inlet, outlet and ambient for each solar collector configuration, corresponding to the volume flow rates of 40 m³/h, 60 m³/h and 80 m³/h, and a solar collector length from 0 m to 1.95 m with a tilt angle of β = 34.8° are given in Table 4. It can be seen from Fig. 9 that the outlet temperature increases with the decrease in volume flow rate. For a specific volume flow rate at a constant ambient temperature, the outlet and inlet temperatures increase with the increase in solar intensity and volume flow rate. Again, it can be clearly observed that the baffle fin facilitates the increase in the outlet-air temperature. In general, the inlet temperature increases linearly at the volume flow rates of 40 m³/h to 80 m³/h, as shown in Fig. 9.

The thermal efficiency of the heater improves with the increase in volume flow rates due to an enhanced heat transfer to the air flow and the temperature difference decreases at a constant tilt angle of 34.8°. The solar intensity is at its highest value at about 12:50 as is expected. The solar intensity decreases as the time passes in the afternoon.

Figure 10 illustrates spatial variation of Nusselt number for different values of volume flow rates and corresponding to different configurations in the solar collectors. As can be noticed, the Nusselt number increases along the air heater. The minimum value of the Nusselt number occurs at the entry section of the air heater which is mostly affected by the heat transfer from the wall and inside the circumference of the air heater. Further, it may be concluded that the heat transfer increases with the increase in volume flow rates, because the increasing volume flow rate leads to a slight temperature increase in the absorber plate and near wall air, and consequently, increases the air viscosity. This increase in air viscosity affects the wall shear stress and decreases the local Re as well which causes an increase in thermal boundary layer thickness, and results in an increase in the convective heat transfer coefficient.

The collector efficiency η is plotted against (T_in – T_a)/I in Fig. 11. The slope of this line (–F_RU_L) represents the rate of heat loss from the collector. For example, the collector with baffle has less of a slope than those without baffle. The curves of Fig. 11 represent changes in the performance of the solar collector studied for three configurations available on the reduced parameter (T_in – T_a)/I.

The best configuration is the one that gives the value of the best performance which is the collector configuration with the baffle plate fixed at the bottom:

F_R(τ_vα_ab) = 0.625 is the rate of heat gain and the highest loss rate F_RU_L = 13.436 lowest where the thermal efficiency is 62.50%, as presented in Table 6.

Two interesting operating points can be observed from Fig. 11. The first is the maximum collection efficiency, called the optical efficiency. This occurs when the fluid inlet temperature equals the ambient temperature (T_i = T_a). For this condition, the ΔT/I value is zero and the intercept is F_R(τα). The other one is the intercept with the ΔT/I axis. This point of operation can be reached when useful energy is no longer removed from the collector, a condition that can happen if fluid flow in the collector stops (power failure). In this case, the optical energy coming in must equal the heat loss, requiring that the temperature of the absorber increases until this balance occurs. This maximum temperature difference or “stagnation temperature” is defined by this point.

Figure 12 shows the solar intensity versus standard local time of the day for three days in which the experiment was carried out. The solar intensity increases from the early hours of day with about 695 W/m², 738 W/m² at 9:30 am and 766 W/m² at 10:45 am, respectively on February 28, April 03, and March 01, 2011 to a peak value at noon and then reduces later on during the day (Fig. 12). The highest daily solar radiation obtained with the single pass solar air collector was 941 W/m², 950 W/m² at 12:00 am and 811 W/m² at 11:45 am. Calculating the mean solar intensity for each day, there was stability in the solar radiation as all mean averages within the same and close range. The mean average solar intensity for the days of the experiment was 858 W/m², 852 W/m² and 760 W/m², respectively for the single pass solar air collector. The obtained result shows a stable amount of solar radiation measured for each day of the experiment.

Conclusions

The present studied aims to analyze a thermal efficiency of SAH. The comparison of solar collectors with smooth plate and with baffle indicates that the efficiency of the solar air collector depends significantly on the solar radiation, the volume flow rate and the position of a baffle when is fixed on the duct of solar collectors. The efficiency of the collector improves by increasing the volume flow rate from 40 m³/h to 80 m³/h due to enhanced heat transfer to the air flow. The efficiency of the solar air collector is proven to be higher. The highest collector efficiency and air temperature rise were achieved by the baffled collector with a tilt angle of 34.8°, whereas the lowest values were obtained from the collector with smooth plate.

Based on the experimental results obtained for the three modes of solar collectors, a forced convection SAH was manufactured and tested in the Laboratory of Mechanical Engineering, University of Biskra, Algeria, and the following conclusions can be drawn:

1) Air mass flow rate and solar radiation are principal factors which affect the performance of SAH.

2) For Nusselt number, heat transfer coefficient and the recommended range of air mass flow rate which gives an appropriate outlet air temperature and thermal efficiency is 40 m³/h–80 m³/h.

3) The rise of the outlet air temperature from the air heater above the ambient air temperature is in the range of 15°C–22°C during the testing runs.

4) Evidently, the mean highest thermal efficiency (η = 58%) was obtained at a solar intensity of I= 937 W/m² by fixing the baffle under the absorber plate at a volume flow rate of 80 m³/h and a tilt angle of 34.8° at 12:15.

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