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Study on the performance of a solar collector with heat collection and storage
Applied Thermal Engineering 147 (2019) 380–389
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Research Paper
Study on the performance of a solar collector with heat collection and storage
T
⁎
Gang Lia, Huilan Huangb, , Junjie Zhangb, Hua Zhangc a
College of Electrical Engineering, Guangxi University, Nanning 530004, China College of Mechanical Engineering, Guangxi University, Nanning 530004, China c Energy and Power Engineering School, University of Shanghai for Science and Technology, Shanghai 200093, China b
A R T I C LE I N FO
A B S T R A C T
Keywords: Solar chimney power plant Solar collector Energy conversion Heat collection Heat storage
The collector in a solar chimney power plant system (SCPPS) is a heat exchanger that converts solar energy into the internal energy of the transport medium. This study proposed a new collector structure that could be used in combination with agricultural production, based on modelling and experimental research. Comparison of the heat collection performance of the new collector structure with the performances of two traditional collector structures, indicated that the new structure, in which the heat absorber plate was located in the middle of the collector, had the best performance. Since the airflow channel was divided into upper and lower channels, the heat transfer area was increased, and the heat transfer in the channel was enhanced. In addition, the perforated heat absorber plate with a certain porosity characteristic had a better heat transfer performance than the full heat absorber plate and could improve the stability and continuous power generation capability of SCPPS.
1. Introduction The solar chimney power plant system (SCPPS) is a promising way to use solar energy for large-scale power generation. As shown in Fig. 1, SCPPS comprises three components, including a solar collector, a chimney situated on the centre of the collector, and a power conversion unit with one or several turbine generators [1,2]. The collectors are mostly circularly distributed and have a certain height relative to the ground. The solar radiation raises the temperature of the fluid inside the collector, while the density of the fluid decreases. Driven by the chimney effect, the fluid flows to the bottom of the chimney to drive the turbine to generate electricity. Numerous research studies have been carried out on the SCPPS. Some researchers have analysed the performance of the SCPPS [3,4], and others have researched its components [5,6]. Among the three components, the solar collector is the key component used to convert solar energy into heat, in accordance to the processes of heat conduction, convection, radiation, and energy storage. The performance of solar collector will have a direct impact on the energy efficiency of the overall system. Therefore, many researchers have focused on improving the performance of the collector. Lodhi [7] studied the characteristics of energy collection and storage in the collector of SCPPS. Schlaich et al. [8] used water-filled black tubes to store energy and analysed the influence of output power. Pasumarthi [9] increased the collector area to ⁎
enhance the air temperature difference and air mass flow rate, thereby increasing the generating capacity of the power generation system. Gannon and von Backström [10] built a simple model to analyse the coupling of the mass flow and temperature rise in the solar collector. Bernardes et al. [11] developed a numerical model and proved that the power output increased by increasing the collector area and the transmittance. Pretorius and Kröger [12] proved that utilising highquality glass could increase the annual plant power output by 3.4%. Frederick et al. [13] found that appreciable electric power (P ≥ 103 W) could be generated by a solar chimney with specific dimensions, thus exhibiting the minimum threshold value of τ = 2.9. Effectively, this value denotes the temperature ratio of the difference between the collector’s surface temperature and the temperature at the turbine (Ts − TH) compared to the difference of the air mass temperature under the roof and the collector’s surface temperature (Tm − Ts). Above the threshold value of, τ, the instantaneous electric power increased exponentially. Zhang et al. [14] established a test model of the collector with a multi-functional air open cycle solar collector and experimentally tested its performance. Previously conducted SCPPS tests mostly focussed on the size of the collector, and experimental heat storage studies are thus still lacking. Therefore, employing different structures of heat collection and storage devices to solve the problem of continuous power generation of SCPPS over 24 h is becoming more and more important [15]. In addition, increasing the temperature difference
Corresponding author. E-mail addresses: [email protected], [email protected] (H. Huang).
https://doi.org/10.1016/j.applthermaleng.2018.10.068 Received 11 December 2017; Received in revised form 13 September 2018; Accepted 17 October 2018 Available online 25 October 2018 1359-4311/ © 2018 Elsevier Ltd. All rights reserved.
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υ ρ λ σ δ ΔT
Nomenclature A cp g H h I0 L P Qv q S T Ub V Gr Nu Pr Ra
area, m2 specific heat at constant pressure, J/(kg·K) gravitational acceleration, m/s2 height, m heat transfer coefficient, W/(m2·K) solar radiation intensity, W/m2 length, m pressure, Pa volume flow, m3/h heat flux, W absorbed solar radiation, W/m2 temperature, K coefficient of heat transfer ground,W/(m2· K) velocity, m/s Grashof number, dimensionless Nusselt number, dimensionless Prandtl number, dimensionless Raleigh number, dimensionless
Subscripts a abs f g gl sl i o q r s w 21 23 1,2,3
Greek symbols β η
kinematic viscosity density, kg/m3 thermal conductivity, W/(m·K) Stefan–Boltzmann constant,W/(m2· K4) thickness, m temperature difference, K
volumetric coefficient of thermal expansion, 1/K efficiency
ambient absorber plate fluid ground glass surface of storage layer inlet outlet heat flux radiation sky wind between surface 2 and surface 1 between surface 2 and surface 3 position 1, position 2, position 3
Hot air
Chimney
Solar radiation
Turbine
Collector
Cold air
Absorber
Storage layer Fig. 1. Schematic of solar chimney power plant.
circumstances. A feasible scheme is associated with the division of the solar collector into two channels. The airflow temperature in the upper channel is higher and the channel is fitted for power generation. In contrast, the lower channel temperature is suitable for agricultural production and thermal storage is installed on the ground to keep the temperature constant over longed time periods. Based on the above purposes, we investigated the performance of the solar collector with a special structure in SCPPS. This includes two steps: a) change of the location of the heat absorber to study the air temperature difference between the inlet and outlet of the collector as well as its efficiency, and b) change of the structure of the absorber to investigate the performance of the collector with the same location in the heat absorber.
between the inner and outer collector parts in the SCPPS can improve the power generation efficiency [16,17]. Therefore, the identification of, measures for the enhancement of the effect of air heat transfer in the collector is very important [18,19]. The solar air collector is usually a single channel with a single-layer glass cover [20,21]. The unsteady sunlight cannot provide radiation energy continually. Correspondingly, the energy absorption and storage issues in the solar collector have become very influential. The SCPPS is difficult to put into practice because the system has a low efficiency and occupies a large area. The more promising way is to develop a composite system. Given that the temperatures of a solar collector in the SCPPS are above the ambient temperature, it is not suitable to allow grazing or plant growth in its interior under the 381
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2. Methodology
and an absorber placed in a middle location is referred to as type III. The heat absorber is made of aluminium with a black coating on the surface. Fig. 2 shows the different absorber plate locations in the solar collector. Each test was repeated three times to ensure data reliability. The relevant parameters, such as the solar radiation intensity, volume flow rate, and the temperature of the measured point and the ambient temperature were recorded, and the temperature difference between inlets and outlets and efficiency were calculated. The type III absorber can be divided into two airflow channels. The heat transfer of the physical model is described below. The energy balance equation for the collector roof is
The SCPPS collector is generally circular with hundreds of meters in diameter. It is too difficult to construct a complete experimental mechanism. Correspondingly, herein, we have used an infinitesimal element that approximated a cuboid as the study object. The studies were carried out on a multi-functional solar collector. It was open at its two ends in the horizontal direction [14]. On the vertical direction, it can be divided into several airflow channels at different heights. In the bottom of the collector, some thermal storage material could be used to enhance energy conversion. The device can be regarded as a micro-unit in the SCPPS collector, which can change the height of the airflow channel, and transform it into different heat collection structures. In the steady state, the performance of the solar collector is evaluated with the comprehensive consideration of the collector’s efficiency and the temperature difference between the collector’s inlet and outlet. The collector efficiency η can be calculated based on the following equation,
η =
h1 × (Tf 1 − Tgl ) + hr ,21 × (Tabs − Tgl ) = h w × (Tgl − Tw ) + hrs × (Tgl − Ts ) (2) The energy balance equation for the airflow between the collector roof and the absorber is
h1 × (Tgl − Tf 1) + h2 × (Tabs − Tf 1) = q1
cp ρQ v ΔT
(3)
Correspondingly, the energy balance equation for the heat absorber is
(1)
I0 A
S1 + h2 × (Tf 1 − Tabs ) = h2′ × (Tabs − Tf 2) + hr ,21 × (Tabs − Tgl ) + hr ,23 × (Tabs − Tsl )
2.1. Physical model and numerical model of heat absorber plate
(4)
Furthermore, the energy balance equation for the airflow between the heat absorber and the storage layer surface is
There were no changes associated with the variation of the operating parameters of the collector in the effort to compare the absorber plate location in the solar collector. The absorber that is located at the bottom part of the collector is referred to as type I, while an absorber located on the top part to replace the glass cover is referred to as type II,
h2′ × (Tabs − Tf 2) + h3 × (Tsl − Tf 2) = q2
(5)
The energy balance equation for the storage layer is
Glass
Absorber
Fluid Fluid
Absorber Storage layer
Storage layer
Ground Ground
(a) Absorber located at a bottom position, type I.
(b) Absorber located at a top position, type II.
hw Tgl Fluid Tabs
Tsl
h1 h2 h2’ h3
Tw
S1
Glass (surface1)
Tf1
hr,21
Absorber (surface2) S2
Tf2
hr,23
Surface of storage layer (surface3)
Tg
Storage layer
Ub
Ground
(c) Absorber located in the middle, type III. Fig. 2. Location of heat absorber plate in the solar collector. 382
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S2 + h3 × (Tf 2 − Tsl ) = hr ,23 × (Tsl − Tabs ) + Ub × (Tsl − Tg )
(6)
Gri =
g × βi × Δti × L3 νi2
(i = 1, 2, 3)
(16)
where S2 represents the solar radiation absorbed by the heat storage layer. The relevant heat transfer coefficient are given as follows, [22]
Under the test conditions, the Grashof number (Gr) or Prandtl number (Pr) are given by
h w = 2.8 + 3.0 × Vw [22]
(7)
Gr1 = 1.263 × 108 × (Tf 1 − Tgl ) × L3 , Pr1 = 0.7010, h1
hrs = σ × εgl × (Tgl + Ts ) × (Tgl2 + Ts2)
(8)
hr ,21 =
hr ,23 =
hi =
= 1.904 × (Tf 1 − Tgl )1/3 Gr2 = 1.010 × 108 × (Tabs − Tf 1) × L3 , Pr2 = 0.6985, h2
2 σ × (Tgl + Tabs ) × (Tgl2 + Tabs )
(1/ εgl + 1/ εabs − 1)
= 1.854 × (Tabs − Tf 1 )1/3
(9)
2 σ × (Tabs + Tg ) × (Tabs + Tsl2 )
L
(i = 1, 2, 3)
= 1.872 × (Tabs − Tf 2 )1/3
= 1.887 × (Tsl − Tf 2 )1/3 (12)
(13)
The required similarity criterion number can be calculated in accordance to [24]
Rai = Pri × Gri
(i = 1, 2, 3)
(14) (15)
(i = 1, 2, 3)
(20)
A one-dimensional radiation model is adopted as the numerical model used in this study. Considering the small energy loss at the transverse boundary of the heat absorbing plate and the storage layer, the transverse heat conduction between adjacent meshes is much weaker than the convection heat transfer on its surface. So the heat transfer between adjacent meshes in absorber plate or storage layer is ignored, This calculation model way is used in the literature [25], and assumed that the airflow had been in a steady state throughout the entire process when it passed through the collector. The calculated area was discretised into 50 grids with a grid distance of 0.04 m, and the heat transfer was calculated based on the one-dimensional equal-pitch
where Ts is the sky temperature used when calculating radiation, and is given as follows [23],
Ts = 0.0552 × Tw1.5
(19)
Gr3 = 1.172 × 108 × (Tsl − Tf 2) × L3 , Pr3 = 0.7000, h3
(11)
Ub = λ g / δg
Nui = 0.16 × Rai1/3 [24]
(18)
Gr2′ = 1.089 × 108 × (Tabs − Tf 2) × L3 , Pr′2 = 0.6991, h2′ (10)
(1/ εabs + 1/ εsl − 1)
Nui × kf
(17)
Start Divide the entire the flow path into 50 one‐dimensional grids Assume that the ambient temperature is the initial temperature of the inlet , the glass plate, the absorber, and the outlet in the first grid Assume that the glass plate, the absorber and the outlet temperature increase in 0.1K steps. Calculate the coefficients of the Eqs. (7)‐(20) in the grid Calculate the outlet temperature of the grid according to this model and compare it with the assume temperature. Choose the result with the smaller error as the outlet temperature Consider that the outlet, the glass plate and the absorber temperature of the current grid is the initial temperature of the inlet, the glass plate, and the absorber of the subsequent grid
No Is this the last grid? Yes End Fig. 3. Numerical simulation flow chart. 383
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2.3. Test rig
grid. By assuming the temperature values at the glass plate and the absorber of the grid, and the outlet temperature of each grid, we can obtain the coefficients of the Eqs. (7)–(20) and begin the calculation. Finally, the outlet temperature of a grid is that which corresponds to the minimum calculated deviation. The inlet temperature of the first grid is the ambient temperature, and the outlet temperature of the previous grid is the inlet temperature of the subsequent grid. By analogy, the calculated airflow mean temperature of the every grid was obtained. The numerical simulation flow chart is as Fig. 3.
The multi-functional solar collector test rig used for the study is shown in Fig. 5. It consisted of a collector box with dimensions of 2000 mm × 1000 mm × 500 mm, a cover with a single glass with a thickness of 5 mm, an emissivity of 0.94, a solar radiation simulation device, a radiation intensity meter, a differential pressure gauge to measure the airflow rate, various sensors/meters to measure the air collector inlet, outlet, or ambient temperatures, and a data acquisition system. A fan was mounted on a windpipe at the collector outlet. It was regarded as the chimney that pumped the air to the upper channels of the atmosphere at its outlet. The airflow rates of the collector’s outlet can be changed by changing the speed of the fan so it can emulate the heat air changes in the collector of an actual solar chimney power system at various operating conditions. In fact, the airflow in the heat collector of SCPPS is a type of natural convection airflow driven by the chimney effect which is due to the temperature and density difference in the heat collector. The airflow rate of the natural convection was constant and in accordance to the specified ambient temperature, radiation intensity, ground absorption condition, and collector size. The experimental method used, whereby a fan suctions the air, does not violate the actual working condition, and it operates based on the same mechanism such that the heated air in the collector is suctioned through the solar chimney and expelled to the ambient environment. The designed collector is only a micro-unit of the overall system, and the size of the overall system is changeable. Given that different sizes of the collectors are associated with various airflow rates, even under the same ambient and radiation intensities, the appropriate experimental method was selected that changed the airflow rates in a gradual manner. The main instruments/equipment used during the tests are listed in Table 1. Solar radiation energy concentrates mainly in the wavelength range of 0.2–2 μm and the zone of visible light occupies most of this range. Several types of lamps have been used as light sources for solar simulators [26]. Herein, 21 calix lamps HW200-250 were used to simulate the sun radiation, and its irradiance can be adjusted in value within the range of 200–1000 W/m2 by a voltage regulator. A TBQ-2 radiometer (accuracy of 5% or less) was used to measure the irradiance of the infrared radiation lamp. Thirty chrome–constantan thermocouples were used to measure the air temperatures in the collector that had been
2.2. Structure of heat absorber plate In the storage layer of the collector, the gravel was regarded as an ideal material for heat absorption and heat storage, as shown in Fig. 4. The heat absorber is a black coating aluminium plate that is placed in the middle of the collector. The new structure proposed in this study should be able to adjust the distribution of radiation to realise heat storage or to change the lower channel temperature to ensure its suitability for agricultural production. Therefore, the absorber is divided into the perforated plate and the full plate for comparison. Fig. 4(a) shows a diagrammatic sketch of the collector with a perforated heat absorber plate. For the comparison to the absorber plate structure played in the same location in the solar collector, the procedures were realised in two steps: (i) a full plate absorber without holes was placed in the middle location of the collector, and (ii) a full plate absorber with holes was placed in the middle location of the collector. The absorber plate was perforated by a periodic hole pattern (with a separation distance t) and with a set perforation rate, as shown in Fig. 4(b). As the hole-spacings became smaller, more holes existed, and the effective area of the collector was correspondingly reduced. In general, t = (1.5–5)d, where d is the hole diameter. The hole spacing was determined in accordance to the perforation rate requirements, that is, the ratio of the area that included the holes to the total area of the entire plate. In this study, the holes-spacing distance was 30 mm. The holes were arranged in an equilateral triangle. The dimensions of the perforated aluminium plate had a length of 2000 mm, a width of 1000 mm, and a thickness of 2 mm. The diameters of the holes were 20 mm. The emissivity of the absorber plate was 0.8.
Glass Collector Perforated heat absorber plate
S1 S2
Gravel
d
(a) Diagrammatic sketch of collector with perforated heat absorber plate.
t (b) Perforated heat absorber plate. Fig. 4. Collector with perforated heat absorber plate. 384
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Fig. 5. Test rig of solar heat collector.
system, there were several ADAM–4018 A/D modules and ADAM–4015 A/D modules that respectively obtained a signal from the thermocouples and the Platinum resistance thermometers, while two ADAM–4017 A/D modules acquired the signal from the pressure difference transmitter and airflow rate. A host computer acquired various experimental parameters data from the Datataker.
Table 1 Instruments and equipment used in the test. Name
Model and specification
Solar radiation source
Twenty one lamps with HW200–250 infrared radiation CECC–6, capacitive pressure transmitter Pt-100 thermal resistance, thermocouple TBQ–2, radiometer ADAM A/D modules
Differential pressure gauge Thermometers Radiation intensity meter Data acquisition units
3. Results and discussions According to the thermal model of the collector mentioned in Section 2, a numerical simulation was performed to calculate air temperature distribution in the collector. The results from the experimental tests and numerical calculations are shown in the following figures.
calibrated with an accuracy of ± 0.1 K. Two grade-A precision Pt-100 platinum resistance thermometers were used to measure the inlet and outlet air temperatures of the collector. The specific measurement method used in the experiment guaranteed that the external environment had no effect on the measured results elicited by the thermocouple. A Venturi tube that was placed into a windpipe was used to measure the airflow rate (accuracy of ± 1%). In the data acquisition
3.1. Air temperature distribution in the direction of airflow through the type I collector The experimental study and numerical calculations were carried out 385
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of the collector.
at the ambient temperature of 306 K (Ta) and at an airflow rate of 105 m3/h (Qv). Considering that the outcomes of the one-dimensional radiation model denote the average temperature, the temperature values at the measured locations were calculated as linear functions of the average temperature. This temperature calculation values at different collector distances and temperatures measured in the experiments are shown in Fig. 6: Because the calculation was the mean value of the longitudinal region of each mesh, the air-flow temperature of each mesh near absorber plate was calculated by linear fitting. From Fig. 6, it can be observed that the air temperature increases along the direction of airflow in the type I collector and that the radiation intensity influenced the air temperature at the outlet of the collector. The experimental results showed that the temperature at 1.2 m increased to 338.6 K under the condition where the radiation intensity was 773 W/m2, which was effectively 19 K higher than the increase that corresponded to an intensity of 325 W/m2. At the same time, we found that the temperature of the measured points at 1.6 m was lower than 1.2 m. The reason is attributed to the fact that the 1.6 m position was near the contraction of the channel at the outlet, which enhanced the mixed intensity of air and influenced the rising trend of the original temperature. Fig. 6 also shows that there was some difference between the numerical and the experimental results. There are three main reasons for this:
3.3. Experimental temperature in type I and type III collectors with solar radiation intensity variations The heat collection effects of the type I and type III collectors were experimentally compared at an ambient temperature of 303 K and an airflow rate of 122 m3/h. Fig. 8 shows the temperature distribution of air in the direction of airflow near the surface of the two absorbers under the three radiation intensities. The heating effect of type III on air was generally higher than type I under the same radiation intensity. In fact, the measured temperatures of type III at 1.6 m were 9.3 K, 10.2 K, and 9.8 K higher than type I at the respective radiation intensities of 267 W/m2, 538 W/m2 and 773 W/m2. Additionally, the temperature differences at the inlet of type III and the outlet reached 13 K, 15.7 K, and 16.5 K, which were higher than those for type I. This was owing to the fact that the type III absorber was not in direct contact with the ground and thus had smaller heat loss, the absorber temperature was higher, and the heating effect on the air was better than that for type I, thus establishing higher type III air temperatures than those for type I. 3.4. Performance comparisons of the absorber plate placed in three locations of the solar collector
(1) The numerical temperature along the measurement point was calculated based on a linear dependence of the average temperature, but the measured data originated from the fixed-point temperature measurements. For example, the temperature of the measured point at 1.6 m was lower than that at 1.2 m, and the point at 1.6 m was near the contraction of the channel at the outlet, which enhanced the mixed intensity of air, and influenced the rising trend of the original temperature. (2) We assumed that the airflow was always steady and ignored the transverse heat conduction between adjacent meshes in the theoretical calculation, but these existed during the experimental process. (3) The viscous force of fluid from the collector was not considered in the numerical calculation model. The viscous drag may slow down the movement of hot air that affected the heat transfer. As a result, the air temperature of collectors was increased.
The temperature differences between the collector inlets and outlets at three different positions of the heat absorber plate were investigated by changing the parameters of airflow conditions but at the same radiation intensity levels. The radiation intensity I0 was 494 W/m2 and the ambient temperature was 308 K. The airflow rates through the collector were changed from low to high. The experimental results of the inlet and outlet temperature differences ΔT and the efficiency η of the three types of the collector at different locations are presented in Figs. 9 and 10. Figs. 9 and 10 show that the efficiency η as well as the inlet and outlet temperature difference ΔT of the collector in the case where the aluminium plate was placed on the top, and with a volume flow rate Qv < 175 m3/h, were lower than that the outcomes elicited by an aluminium plate placed at the bottom or centre of the collectors. When Qv > 175 m3/h, both ΔT and η increased compared to the case where the collector was located at the bottom, but they were still lower than the case where the collector was in the middle location. It is obvious that the collector with the aluminum plate placed at the top had no advantage except in the cases associated with higher airflow rates. This
3.2. Effect of the airflow channel length on the type I collector The numerical results obtained from the condition of the airflow rate of 105 m3/h were taken as an example. Fig. 7 shows the theoretically calculated results of the air temperature affected by the length of the airflow channel in the collector based on a radiation intensity of 1000 W/m2 and an ambient temperature of 306 K. Under the same conditions, there were no changes in other parameters except the airflow channel length, and the air temperature of the collector outlet was increased with increases of the airflow channel length. However, after a certain length, the increasing trend of outlet air temperature slowed down. It can be observed that when the channel length was 30 m, the collector outlet temperature was 361.1 K, which was just 0.9 K higher than the case where the channel length was 25 m and at a temperature at 2 K higher than that for a channel length of 20 m. The reason is attributed to the fact that as the air flowed toward the centre of the collector, the airflow velocity increased and the surface area of the collector decreased. Therefore, the same airflow received less solar radiation, and the temperature tended to be stable. Thus, increasing the length of the channel was not an effective way to increase the temperature of the outlet air. The use of double glazing or a film on the collector cap can reduce the heat loss by increasing the thermal resistance. In addition, the heat absorber plate can be located in the middle of the collector (Fig. 2(c)) to improve the heat-transfer capacity
360
Temperature (K)
350
340
773 W/m2 (exp) 674 W/m2 (exp) 494 W/m2 (exp) 325 W/m2 (exp) 773 W/m2 (cal) 674 W/m2 (cal) 494 W/m2 (cal) 325 W/m2 (cal)
330
320 Ta = 306 K
310
300 0.0
Qv = 105 m3/h
0.2
0.4
0.6
0.8
1.0
1.2
1.4
1.6
1.8
2.0
Distance to the collector inlet (m)
Fig. 6. Measured temperatures (points) and calculated temperatures (lines) in type I collector with solar radiation intensity variations. 386
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370
20
360
Temperature (K)
350
L = 30 m L = 25 m L = 20 m L = 15 m L = 10 m L=8m L=4m L=2m
Type I (linear fitting) Type II (linear fitting) Type III (linear fitting)
18 Temperature difference ΔT (K)
I0 = 1000 W/m2 Ta = 306 K Qv = 105 m3/h
340 330 320
16 14
Type III (middle)
12 10
Type I (bottom)
8 6 4
Type II (top)
310
Ta = 308 K
2 300 0
5
10
15
20
25
30
0 60
35
I0 = 494 W/m2
80
100
120
Distance to collector outlet (m)
Fig. 7. Influence of airflow channel length on the calculated temperature in the type I collector.
180
200
220
240
260
90 773 W/m2 (type III, exp) 773 W/m2 (type III, linear fitting) 538 W/m2 (type III, exp) 538 W/m2 (type III, linear fitting) 267 W/m2 (type III, exp) 267 W/m2 (type III, linear fitting)
325
Ta = 303 K
773 W/m2 (type I, exp) 773 W/m2 (type I, linear fitting) 538 W/m2 (type I, exp) 538 W/m2 (type I, linear fitting) 267 W/m2 (type I, exp) 267 W/m2 (type I, linear fitting)
80
Ta = 308 K I0 = 494 W/m2
70
Qv = 122 m3/h
Efficiency η (%)
Temperature (K)
330
160
Fig. 9. Temperature difference between the air inlet and outlet of the collector at different positions of the heat absorber plate.
340 335
140
Volume flow rate Qv (m3/h)
320 315
Type III (middle)
60 50 Type I (bottom)
40 30
310
Type II (top)
20 305
Type I (second-order fitting) Type II (second-order fitting) Type III (second-order fitting)
10 300 0.0
0.5
1.0
1.5
0 60
2.0
Distance to the collector inlet (m)
80
100
120
140
160
180
200
220
240
260
Volume flow rate Qv (m /h) 3
Fig. 8. Experimental temperature in type I and type III collectors with solar radiation intensity variations.
Fig. 10. Efficiency of the collector at different locations of the heat absorber plate.
is because the aluminium plate placed on the top replaced the glass cover. Without glass, the aluminium plate surface temperature (the measured value was 346 K) increased more than that at the bottom (measured value of 337.5 K). The hot face of the aluminium plate slowed down the airflow in the collector, i.e. the air near the hot surface did not flow easily at low velocities. Therefore, ΔT and η of the collector increased slowly. With the increase of the airflow rate, the heat transfer increased rapidly on the hot surface of the aluminium plate located on top, followed by abrupt ΔT and η increases. Figs. 9 and 10 indicate that the collector with an intermediate absorber plate exhibits an excellent performance regarding η and ΔT, compared to the corresponding values elicited at other locations of aluminium plates. Considering that both sides of the absorber are surrounded by fluid, which can reduce heat loss owing to the greenhouse effect, the surface temperature of the aluminium plate could rise to 340 K. The airflow was heated after it flowed through the top or the back of the heat absorber plate, and the area of the heat exchange surface inside the collector was doubled. This ultimately resulted in the enhancement of ΔT and η.
3.5. Effect of the different structures with and without absorber plate holes on the solar collector As mentioned above, the ΔT and η values of the collector increased when the aluminium plate was placed in the middle of the collector as a heat absorber plate. Combination of the two methods mentioned above allowed the conduct of experimental tests by changing the form of the heat absorber in order to examine the performance of the collector. Temperature changes in the full- and perforated-plate collectors were investigated at the same radiation intensity of 674 W/m2, ambient temperature of 291 K, and changing airflow rate. The experimental results of the temperature difference ΔT of the two types of collector structures are shown in Fig. 11. The perforated absorber shown in Fig. 4(b) was used for experiments. The experimental results of the efficiency η of the two types of collector structures are shown in Fig. 12. It can be seen from Figs. 11 and 12 that the ΔT and η of the perforated plate absorber with a certain porosity are higher than those of the full plate. Thus, the perforated plate has a better heat collecting performance than the full plate. This is because the air can only 387
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solar heat collector. The inlet and outlet temperature difference ΔT and efficiency η were higher than that of the solar collector with the full plate absorber. (3) The air temperature at the outlet of the solar collector increased as the airflow channel length increased, and the rising trend slowed down. Therefore, in the SCPPS, increasing the radius of the collector was not an effective way to improve the outlet temperature of the collector. (4) The airflow channel was subdivided into two channels, the upper air temperature was higher and suitable for power generation, the lower air temperature was moderate for agricultural production, and the heat storage material could be arranged on the ground to maintain its temperature for a long time.
25 Ta = 291 K
Temperature difference ΔT (K)
I0 = 674 W/m2
20 Perforated plate 15
10
Full plate
5 Perforated plate (linear fitting) Full plate (linear fitting)
0 80
100
120
140
160
180
200
220
240
The above results can be used for the optimal design of heat collection and thermal storage structures of large SCPPS. In an actual SCPPS, if the perforated plate absorber and heat storage layer are applied, the performance of the solar collector will be enhanced without occupying any other areas.
260
Volume flow rate Qv (m3/h)
Fig. 11. The temperature difference between air inlet and outlet of the collector with different structural absorber versus volume flow rate.
Acknowledgements This research was financially supported by the National Natural Science Foundation of China (No. 51466001, 51266001), the Natural Science Foundation of Guangxi (No. 2017GXNSFDA198042), Guangxi Key Laboratory of Manufacturing System & Advanced Manufacturing Technology Foundation (16-380-12S003). The authors would like to thank Elsevier for improving the language quality of the revised manuscript.
65 60
Ta = 291 K I0 = 674 W/m2
Perforated plate
Efficiency η (%)
55 50 Full plate
45
Appendix A. Supplementary material
40 35
Supplementary data to this article can be found online at https:// doi.org/10.1016/j.applthermaleng.2018.10.068.
30
References Perforated plate (second-order fitting) Full plate (second-order fitting)
25 20 80
100
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[1] A.B. Kasaeian, Sh. Molana, K. Rahmani, D. Wen, A review on solar chimney systems, Renew. Sustain. Energy Rev. 67 (2017) 954–987. [2] P.H. Guo, Y.F. Wang, J.Y. Li, Y. Wang, Thermodynamic analysis of a solar chimney power plant system with soil heat storage, Appl. Therm. Eng. 100 (2016) 1076–1084. [3] C.O. Okoye, O. Taylan, Performance analysis of a solar chimney power plant for rural areas in Nigeria, Renewable Energy 104 (2017) 96–108. [4] A. Koonsrisuk, T. Chitsomboon, Mathematical modeling of solar chimney power plants, Energy 51 (2013) 314–322. [5] Y.J. Choi, D.H. Kam, Y.W. Park, Y.H. Jeong, Development of analytical model for solar chimney power plant with and without water storage system, Energy 112 (2016) 200–207. [6] X.P. Zhou, M.A.D.S. Bernardes, R.M. Ochieng, Influence of atmospheric cross flow on solar updraft tower inflow, Energy 42 (2012) 393–400. [7] M.A.K. Lodhi, Thermal collection and storage of solar energies, in: A. Mufti, T.N. Veziroglu, Proc. Int. Symp. Workshop on Silicon Technology Development, 1987. [8] J. Schlaich, R. Bergermann, W. Schiel, et al., Design of commercial solar updraft tower systems-utilization of solar induced convective flows for power generation, ASME J. Sol. Energy Eng. 127 (2005) 117–124. [9] N. Pasumarthi, S.A. Sherif, Experimental and theoretical performance of a demonstration solar chimney model-Part I: mathematical model development, Int. J. Energy Res. 22 (1998) 277–288. [10] A.J. Gannon, T.W. Von Backström, Solar chimney cycle analysis with system loss and solar collector performance, ASME J. Sol. Energy Eng. 122 (2000) 133–137. [11] M.A.D.S. Bernardes, A. Voß, G. Weinrebe, Thermal and technical analyses of solar chimneys, Sol. Energy 75 (2003) 511–524. [12] J.P. Pretorius, D.G. Kröger, Critical evaluation of solar chimney power plant performance, Sol. Energy 80 (2006) 535–544. [13] F.N. Onyango, R.M. Ochieng, The potential of solar chimney for application in rural areas of developing countries, Fuel 85 (2006) 2561–2566. [14] H. Zhang, J.M. Zhang, H.L. Huang, et al., Experimental Device Setup for Solar Chimney Collector, J. University of Shanghai for Science and Technology, 2007, pp. 383–385. [15] M.I. Hussain, H. Glee, Thermal performance evaluation of a conical solar water heater integrated with a thermal storage system, Energy Convers. Manage. 87 (2014) 267–273. [16] G. Rodono, R. Voples, Heat transfer calculation in a free convection air solar
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Volume flow rate Qv (m /h) 3
Fig. 12. The efficiency of the collector with different structural absorber versus volume flow rate.
exchange heat on the surface of the full plate absorber. Instead, the perforated structure allowed air to pass through holes to create turbulent flow, which is conductive to heat transfer. At the same time, air between the two channels can be mixed to increase the heat efficiency of the absorber. 4. Conclusions A new collector structure was proposed with the perforated plate absorber arranged at the middle position. A multi-functional solar collector test rig was established and related experiments and numerical simulations were carried out. The conclusions are as follows: (1) The position of the absorber has an important influence on the performance of the solar collector. When the absorber was in the middle position, the airflow was divided into two channels, the heat exchange area between the absorber and airflow increased, and the efficiency of the solar collector improved. (2) The perforated plate absorber with certain porosity enhanced the air heat transfer and improved the heat storage performance of the 388
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1209–1227. [22] J.A. Duffie, W.A. Beckman, Solar Engineering of Thermal Processes, second ed., Wiley Interscience, NewYork, 1991. [23] Y.J. Luo, S.N. He, C.G. Wang, Solar Energy Utilization Technology, Chemical Industry Press, Beijing, 2014. [24] S.M. Yang, W.X. Tao, Heat Transfer, fourth ed., Higher Education Press, Beijing, 2006. [25] J.Y. Li, P.H. Guo, Y. Wang, Effects of collector radius and chimney height on power output of a solar chimney power plant with turbines, Renewable Energy 47 (2012) 21–28. [26] P.H. Guo, Y. Wang, Q.L. Meng, J.Y. Li, Experimental study on an indoor scale solar chimney setup in an artificial environment simulation laboratory, Appl. Therm. Eng. 107 (2016) 818–826.
collector, J. Energy Build. 27 (1998) 21–27. [17] E. Bilgen, B.J.D. Bakeka, Solar collector systems to provide hot air in rural applications, Renewable Energy 33 (2008) 1461–1468. [18] H.L. Huang, H. Zhang, E.K. An, X.P. Yao, Z.M. Wu, Influence of thermal storage materials on thermal performance of heat collector in solar chimney power generations, J. Power Eng. 28 (4) (2008) 647–650. [19] M.A.D.S. Bernardes, T.W. Von Backström, D.G. Kröger, Analysis of some available heat transfer coefficients applicable to solar chimney power plant collectors, Sol. Energy 83 (2009) 264–275. [20] J.Q. Lin, Numerical simulate of solar air collector I type, Acta Energiae Solaris Sin. 3 (1999) 38–43. [21] F.K. Forson, M.A.A. Nazha, H. Rajakaruna, Experimental and simulation studies on a single pass, double duct solar heater, Energy Convers. Manage. 44 (2003)
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Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng
Research Paper
Study on the performance of a solar collector with heat collection and storage
T
⁎
Gang Lia, Huilan Huangb, , Junjie Zhangb, Hua Zhangc a
College of Electrical Engineering, Guangxi University, Nanning 530004, China College of Mechanical Engineering, Guangxi University, Nanning 530004, China c Energy and Power Engineering School, University of Shanghai for Science and Technology, Shanghai 200093, China b
A R T I C LE I N FO
A B S T R A C T
Keywords: Solar chimney power plant Solar collector Energy conversion Heat collection Heat storage
The collector in a solar chimney power plant system (SCPPS) is a heat exchanger that converts solar energy into the internal energy of the transport medium. This study proposed a new collector structure that could be used in combination with agricultural production, based on modelling and experimental research. Comparison of the heat collection performance of the new collector structure with the performances of two traditional collector structures, indicated that the new structure, in which the heat absorber plate was located in the middle of the collector, had the best performance. Since the airflow channel was divided into upper and lower channels, the heat transfer area was increased, and the heat transfer in the channel was enhanced. In addition, the perforated heat absorber plate with a certain porosity characteristic had a better heat transfer performance than the full heat absorber plate and could improve the stability and continuous power generation capability of SCPPS.
1. Introduction The solar chimney power plant system (SCPPS) is a promising way to use solar energy for large-scale power generation. As shown in Fig. 1, SCPPS comprises three components, including a solar collector, a chimney situated on the centre of the collector, and a power conversion unit with one or several turbine generators [1,2]. The collectors are mostly circularly distributed and have a certain height relative to the ground. The solar radiation raises the temperature of the fluid inside the collector, while the density of the fluid decreases. Driven by the chimney effect, the fluid flows to the bottom of the chimney to drive the turbine to generate electricity. Numerous research studies have been carried out on the SCPPS. Some researchers have analysed the performance of the SCPPS [3,4], and others have researched its components [5,6]. Among the three components, the solar collector is the key component used to convert solar energy into heat, in accordance to the processes of heat conduction, convection, radiation, and energy storage. The performance of solar collector will have a direct impact on the energy efficiency of the overall system. Therefore, many researchers have focused on improving the performance of the collector. Lodhi [7] studied the characteristics of energy collection and storage in the collector of SCPPS. Schlaich et al. [8] used water-filled black tubes to store energy and analysed the influence of output power. Pasumarthi [9] increased the collector area to ⁎
enhance the air temperature difference and air mass flow rate, thereby increasing the generating capacity of the power generation system. Gannon and von Backström [10] built a simple model to analyse the coupling of the mass flow and temperature rise in the solar collector. Bernardes et al. [11] developed a numerical model and proved that the power output increased by increasing the collector area and the transmittance. Pretorius and Kröger [12] proved that utilising highquality glass could increase the annual plant power output by 3.4%. Frederick et al. [13] found that appreciable electric power (P ≥ 103 W) could be generated by a solar chimney with specific dimensions, thus exhibiting the minimum threshold value of τ = 2.9. Effectively, this value denotes the temperature ratio of the difference between the collector’s surface temperature and the temperature at the turbine (Ts − TH) compared to the difference of the air mass temperature under the roof and the collector’s surface temperature (Tm − Ts). Above the threshold value of, τ, the instantaneous electric power increased exponentially. Zhang et al. [14] established a test model of the collector with a multi-functional air open cycle solar collector and experimentally tested its performance. Previously conducted SCPPS tests mostly focussed on the size of the collector, and experimental heat storage studies are thus still lacking. Therefore, employing different structures of heat collection and storage devices to solve the problem of continuous power generation of SCPPS over 24 h is becoming more and more important [15]. In addition, increasing the temperature difference
Corresponding author. E-mail addresses: [email protected], [email protected] (H. Huang).
https://doi.org/10.1016/j.applthermaleng.2018.10.068 Received 11 December 2017; Received in revised form 13 September 2018; Accepted 17 October 2018 Available online 25 October 2018 1359-4311/ © 2018 Elsevier Ltd. All rights reserved.
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υ ρ λ σ δ ΔT
Nomenclature A cp g H h I0 L P Qv q S T Ub V Gr Nu Pr Ra
area, m2 specific heat at constant pressure, J/(kg·K) gravitational acceleration, m/s2 height, m heat transfer coefficient, W/(m2·K) solar radiation intensity, W/m2 length, m pressure, Pa volume flow, m3/h heat flux, W absorbed solar radiation, W/m2 temperature, K coefficient of heat transfer ground,W/(m2· K) velocity, m/s Grashof number, dimensionless Nusselt number, dimensionless Prandtl number, dimensionless Raleigh number, dimensionless
Subscripts a abs f g gl sl i o q r s w 21 23 1,2,3
Greek symbols β η
kinematic viscosity density, kg/m3 thermal conductivity, W/(m·K) Stefan–Boltzmann constant,W/(m2· K4) thickness, m temperature difference, K
volumetric coefficient of thermal expansion, 1/K efficiency
ambient absorber plate fluid ground glass surface of storage layer inlet outlet heat flux radiation sky wind between surface 2 and surface 1 between surface 2 and surface 3 position 1, position 2, position 3
Hot air
Chimney
Solar radiation
Turbine
Collector
Cold air
Absorber
Storage layer Fig. 1. Schematic of solar chimney power plant.
circumstances. A feasible scheme is associated with the division of the solar collector into two channels. The airflow temperature in the upper channel is higher and the channel is fitted for power generation. In contrast, the lower channel temperature is suitable for agricultural production and thermal storage is installed on the ground to keep the temperature constant over longed time periods. Based on the above purposes, we investigated the performance of the solar collector with a special structure in SCPPS. This includes two steps: a) change of the location of the heat absorber to study the air temperature difference between the inlet and outlet of the collector as well as its efficiency, and b) change of the structure of the absorber to investigate the performance of the collector with the same location in the heat absorber.
between the inner and outer collector parts in the SCPPS can improve the power generation efficiency [16,17]. Therefore, the identification of, measures for the enhancement of the effect of air heat transfer in the collector is very important [18,19]. The solar air collector is usually a single channel with a single-layer glass cover [20,21]. The unsteady sunlight cannot provide radiation energy continually. Correspondingly, the energy absorption and storage issues in the solar collector have become very influential. The SCPPS is difficult to put into practice because the system has a low efficiency and occupies a large area. The more promising way is to develop a composite system. Given that the temperatures of a solar collector in the SCPPS are above the ambient temperature, it is not suitable to allow grazing or plant growth in its interior under the 381
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2. Methodology
and an absorber placed in a middle location is referred to as type III. The heat absorber is made of aluminium with a black coating on the surface. Fig. 2 shows the different absorber plate locations in the solar collector. Each test was repeated three times to ensure data reliability. The relevant parameters, such as the solar radiation intensity, volume flow rate, and the temperature of the measured point and the ambient temperature were recorded, and the temperature difference between inlets and outlets and efficiency were calculated. The type III absorber can be divided into two airflow channels. The heat transfer of the physical model is described below. The energy balance equation for the collector roof is
The SCPPS collector is generally circular with hundreds of meters in diameter. It is too difficult to construct a complete experimental mechanism. Correspondingly, herein, we have used an infinitesimal element that approximated a cuboid as the study object. The studies were carried out on a multi-functional solar collector. It was open at its two ends in the horizontal direction [14]. On the vertical direction, it can be divided into several airflow channels at different heights. In the bottom of the collector, some thermal storage material could be used to enhance energy conversion. The device can be regarded as a micro-unit in the SCPPS collector, which can change the height of the airflow channel, and transform it into different heat collection structures. In the steady state, the performance of the solar collector is evaluated with the comprehensive consideration of the collector’s efficiency and the temperature difference between the collector’s inlet and outlet. The collector efficiency η can be calculated based on the following equation,
η =
h1 × (Tf 1 − Tgl ) + hr ,21 × (Tabs − Tgl ) = h w × (Tgl − Tw ) + hrs × (Tgl − Ts ) (2) The energy balance equation for the airflow between the collector roof and the absorber is
h1 × (Tgl − Tf 1) + h2 × (Tabs − Tf 1) = q1
cp ρQ v ΔT
(3)
Correspondingly, the energy balance equation for the heat absorber is
(1)
I0 A
S1 + h2 × (Tf 1 − Tabs ) = h2′ × (Tabs − Tf 2) + hr ,21 × (Tabs − Tgl ) + hr ,23 × (Tabs − Tsl )
2.1. Physical model and numerical model of heat absorber plate
(4)
Furthermore, the energy balance equation for the airflow between the heat absorber and the storage layer surface is
There were no changes associated with the variation of the operating parameters of the collector in the effort to compare the absorber plate location in the solar collector. The absorber that is located at the bottom part of the collector is referred to as type I, while an absorber located on the top part to replace the glass cover is referred to as type II,
h2′ × (Tabs − Tf 2) + h3 × (Tsl − Tf 2) = q2
(5)
The energy balance equation for the storage layer is
Glass
Absorber
Fluid Fluid
Absorber Storage layer
Storage layer
Ground Ground
(a) Absorber located at a bottom position, type I.
(b) Absorber located at a top position, type II.
hw Tgl Fluid Tabs
Tsl
h1 h2 h2’ h3
Tw
S1
Glass (surface1)
Tf1
hr,21
Absorber (surface2) S2
Tf2
hr,23
Surface of storage layer (surface3)
Tg
Storage layer
Ub
Ground
(c) Absorber located in the middle, type III. Fig. 2. Location of heat absorber plate in the solar collector. 382
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S2 + h3 × (Tf 2 − Tsl ) = hr ,23 × (Tsl − Tabs ) + Ub × (Tsl − Tg )
(6)
Gri =
g × βi × Δti × L3 νi2
(i = 1, 2, 3)
(16)
where S2 represents the solar radiation absorbed by the heat storage layer. The relevant heat transfer coefficient are given as follows, [22]
Under the test conditions, the Grashof number (Gr) or Prandtl number (Pr) are given by
h w = 2.8 + 3.0 × Vw [22]
(7)
Gr1 = 1.263 × 108 × (Tf 1 − Tgl ) × L3 , Pr1 = 0.7010, h1
hrs = σ × εgl × (Tgl + Ts ) × (Tgl2 + Ts2)
(8)
hr ,21 =
hr ,23 =
hi =
= 1.904 × (Tf 1 − Tgl )1/3 Gr2 = 1.010 × 108 × (Tabs − Tf 1) × L3 , Pr2 = 0.6985, h2
2 σ × (Tgl + Tabs ) × (Tgl2 + Tabs )
(1/ εgl + 1/ εabs − 1)
= 1.854 × (Tabs − Tf 1 )1/3
(9)
2 σ × (Tabs + Tg ) × (Tabs + Tsl2 )
L
(i = 1, 2, 3)
= 1.872 × (Tabs − Tf 2 )1/3
= 1.887 × (Tsl − Tf 2 )1/3 (12)
(13)
The required similarity criterion number can be calculated in accordance to [24]
Rai = Pri × Gri
(i = 1, 2, 3)
(14) (15)
(i = 1, 2, 3)
(20)
A one-dimensional radiation model is adopted as the numerical model used in this study. Considering the small energy loss at the transverse boundary of the heat absorbing plate and the storage layer, the transverse heat conduction between adjacent meshes is much weaker than the convection heat transfer on its surface. So the heat transfer between adjacent meshes in absorber plate or storage layer is ignored, This calculation model way is used in the literature [25], and assumed that the airflow had been in a steady state throughout the entire process when it passed through the collector. The calculated area was discretised into 50 grids with a grid distance of 0.04 m, and the heat transfer was calculated based on the one-dimensional equal-pitch
where Ts is the sky temperature used when calculating radiation, and is given as follows [23],
Ts = 0.0552 × Tw1.5
(19)
Gr3 = 1.172 × 108 × (Tsl − Tf 2) × L3 , Pr3 = 0.7000, h3
(11)
Ub = λ g / δg
Nui = 0.16 × Rai1/3 [24]
(18)
Gr2′ = 1.089 × 108 × (Tabs − Tf 2) × L3 , Pr′2 = 0.6991, h2′ (10)
(1/ εabs + 1/ εsl − 1)
Nui × kf
(17)
Start Divide the entire the flow path into 50 one‐dimensional grids Assume that the ambient temperature is the initial temperature of the inlet , the glass plate, the absorber, and the outlet in the first grid Assume that the glass plate, the absorber and the outlet temperature increase in 0.1K steps. Calculate the coefficients of the Eqs. (7)‐(20) in the grid Calculate the outlet temperature of the grid according to this model and compare it with the assume temperature. Choose the result with the smaller error as the outlet temperature Consider that the outlet, the glass plate and the absorber temperature of the current grid is the initial temperature of the inlet, the glass plate, and the absorber of the subsequent grid
No Is this the last grid? Yes End Fig. 3. Numerical simulation flow chart. 383
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2.3. Test rig
grid. By assuming the temperature values at the glass plate and the absorber of the grid, and the outlet temperature of each grid, we can obtain the coefficients of the Eqs. (7)–(20) and begin the calculation. Finally, the outlet temperature of a grid is that which corresponds to the minimum calculated deviation. The inlet temperature of the first grid is the ambient temperature, and the outlet temperature of the previous grid is the inlet temperature of the subsequent grid. By analogy, the calculated airflow mean temperature of the every grid was obtained. The numerical simulation flow chart is as Fig. 3.
The multi-functional solar collector test rig used for the study is shown in Fig. 5. It consisted of a collector box with dimensions of 2000 mm × 1000 mm × 500 mm, a cover with a single glass with a thickness of 5 mm, an emissivity of 0.94, a solar radiation simulation device, a radiation intensity meter, a differential pressure gauge to measure the airflow rate, various sensors/meters to measure the air collector inlet, outlet, or ambient temperatures, and a data acquisition system. A fan was mounted on a windpipe at the collector outlet. It was regarded as the chimney that pumped the air to the upper channels of the atmosphere at its outlet. The airflow rates of the collector’s outlet can be changed by changing the speed of the fan so it can emulate the heat air changes in the collector of an actual solar chimney power system at various operating conditions. In fact, the airflow in the heat collector of SCPPS is a type of natural convection airflow driven by the chimney effect which is due to the temperature and density difference in the heat collector. The airflow rate of the natural convection was constant and in accordance to the specified ambient temperature, radiation intensity, ground absorption condition, and collector size. The experimental method used, whereby a fan suctions the air, does not violate the actual working condition, and it operates based on the same mechanism such that the heated air in the collector is suctioned through the solar chimney and expelled to the ambient environment. The designed collector is only a micro-unit of the overall system, and the size of the overall system is changeable. Given that different sizes of the collectors are associated with various airflow rates, even under the same ambient and radiation intensities, the appropriate experimental method was selected that changed the airflow rates in a gradual manner. The main instruments/equipment used during the tests are listed in Table 1. Solar radiation energy concentrates mainly in the wavelength range of 0.2–2 μm and the zone of visible light occupies most of this range. Several types of lamps have been used as light sources for solar simulators [26]. Herein, 21 calix lamps HW200-250 were used to simulate the sun radiation, and its irradiance can be adjusted in value within the range of 200–1000 W/m2 by a voltage regulator. A TBQ-2 radiometer (accuracy of 5% or less) was used to measure the irradiance of the infrared radiation lamp. Thirty chrome–constantan thermocouples were used to measure the air temperatures in the collector that had been
2.2. Structure of heat absorber plate In the storage layer of the collector, the gravel was regarded as an ideal material for heat absorption and heat storage, as shown in Fig. 4. The heat absorber is a black coating aluminium plate that is placed in the middle of the collector. The new structure proposed in this study should be able to adjust the distribution of radiation to realise heat storage or to change the lower channel temperature to ensure its suitability for agricultural production. Therefore, the absorber is divided into the perforated plate and the full plate for comparison. Fig. 4(a) shows a diagrammatic sketch of the collector with a perforated heat absorber plate. For the comparison to the absorber plate structure played in the same location in the solar collector, the procedures were realised in two steps: (i) a full plate absorber without holes was placed in the middle location of the collector, and (ii) a full plate absorber with holes was placed in the middle location of the collector. The absorber plate was perforated by a periodic hole pattern (with a separation distance t) and with a set perforation rate, as shown in Fig. 4(b). As the hole-spacings became smaller, more holes existed, and the effective area of the collector was correspondingly reduced. In general, t = (1.5–5)d, where d is the hole diameter. The hole spacing was determined in accordance to the perforation rate requirements, that is, the ratio of the area that included the holes to the total area of the entire plate. In this study, the holes-spacing distance was 30 mm. The holes were arranged in an equilateral triangle. The dimensions of the perforated aluminium plate had a length of 2000 mm, a width of 1000 mm, and a thickness of 2 mm. The diameters of the holes were 20 mm. The emissivity of the absorber plate was 0.8.
Glass Collector Perforated heat absorber plate
S1 S2
Gravel
d
(a) Diagrammatic sketch of collector with perforated heat absorber plate.
t (b) Perforated heat absorber plate. Fig. 4. Collector with perforated heat absorber plate. 384
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Fig. 5. Test rig of solar heat collector.
system, there were several ADAM–4018 A/D modules and ADAM–4015 A/D modules that respectively obtained a signal from the thermocouples and the Platinum resistance thermometers, while two ADAM–4017 A/D modules acquired the signal from the pressure difference transmitter and airflow rate. A host computer acquired various experimental parameters data from the Datataker.
Table 1 Instruments and equipment used in the test. Name
Model and specification
Solar radiation source
Twenty one lamps with HW200–250 infrared radiation CECC–6, capacitive pressure transmitter Pt-100 thermal resistance, thermocouple TBQ–2, radiometer ADAM A/D modules
Differential pressure gauge Thermometers Radiation intensity meter Data acquisition units
3. Results and discussions According to the thermal model of the collector mentioned in Section 2, a numerical simulation was performed to calculate air temperature distribution in the collector. The results from the experimental tests and numerical calculations are shown in the following figures.
calibrated with an accuracy of ± 0.1 K. Two grade-A precision Pt-100 platinum resistance thermometers were used to measure the inlet and outlet air temperatures of the collector. The specific measurement method used in the experiment guaranteed that the external environment had no effect on the measured results elicited by the thermocouple. A Venturi tube that was placed into a windpipe was used to measure the airflow rate (accuracy of ± 1%). In the data acquisition
3.1. Air temperature distribution in the direction of airflow through the type I collector The experimental study and numerical calculations were carried out 385
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of the collector.
at the ambient temperature of 306 K (Ta) and at an airflow rate of 105 m3/h (Qv). Considering that the outcomes of the one-dimensional radiation model denote the average temperature, the temperature values at the measured locations were calculated as linear functions of the average temperature. This temperature calculation values at different collector distances and temperatures measured in the experiments are shown in Fig. 6: Because the calculation was the mean value of the longitudinal region of each mesh, the air-flow temperature of each mesh near absorber plate was calculated by linear fitting. From Fig. 6, it can be observed that the air temperature increases along the direction of airflow in the type I collector and that the radiation intensity influenced the air temperature at the outlet of the collector. The experimental results showed that the temperature at 1.2 m increased to 338.6 K under the condition where the radiation intensity was 773 W/m2, which was effectively 19 K higher than the increase that corresponded to an intensity of 325 W/m2. At the same time, we found that the temperature of the measured points at 1.6 m was lower than 1.2 m. The reason is attributed to the fact that the 1.6 m position was near the contraction of the channel at the outlet, which enhanced the mixed intensity of air and influenced the rising trend of the original temperature. Fig. 6 also shows that there was some difference between the numerical and the experimental results. There are three main reasons for this:
3.3. Experimental temperature in type I and type III collectors with solar radiation intensity variations The heat collection effects of the type I and type III collectors were experimentally compared at an ambient temperature of 303 K and an airflow rate of 122 m3/h. Fig. 8 shows the temperature distribution of air in the direction of airflow near the surface of the two absorbers under the three radiation intensities. The heating effect of type III on air was generally higher than type I under the same radiation intensity. In fact, the measured temperatures of type III at 1.6 m were 9.3 K, 10.2 K, and 9.8 K higher than type I at the respective radiation intensities of 267 W/m2, 538 W/m2 and 773 W/m2. Additionally, the temperature differences at the inlet of type III and the outlet reached 13 K, 15.7 K, and 16.5 K, which were higher than those for type I. This was owing to the fact that the type III absorber was not in direct contact with the ground and thus had smaller heat loss, the absorber temperature was higher, and the heating effect on the air was better than that for type I, thus establishing higher type III air temperatures than those for type I. 3.4. Performance comparisons of the absorber plate placed in three locations of the solar collector
(1) The numerical temperature along the measurement point was calculated based on a linear dependence of the average temperature, but the measured data originated from the fixed-point temperature measurements. For example, the temperature of the measured point at 1.6 m was lower than that at 1.2 m, and the point at 1.6 m was near the contraction of the channel at the outlet, which enhanced the mixed intensity of air, and influenced the rising trend of the original temperature. (2) We assumed that the airflow was always steady and ignored the transverse heat conduction between adjacent meshes in the theoretical calculation, but these existed during the experimental process. (3) The viscous force of fluid from the collector was not considered in the numerical calculation model. The viscous drag may slow down the movement of hot air that affected the heat transfer. As a result, the air temperature of collectors was increased.
The temperature differences between the collector inlets and outlets at three different positions of the heat absorber plate were investigated by changing the parameters of airflow conditions but at the same radiation intensity levels. The radiation intensity I0 was 494 W/m2 and the ambient temperature was 308 K. The airflow rates through the collector were changed from low to high. The experimental results of the inlet and outlet temperature differences ΔT and the efficiency η of the three types of the collector at different locations are presented in Figs. 9 and 10. Figs. 9 and 10 show that the efficiency η as well as the inlet and outlet temperature difference ΔT of the collector in the case where the aluminium plate was placed on the top, and with a volume flow rate Qv < 175 m3/h, were lower than that the outcomes elicited by an aluminium plate placed at the bottom or centre of the collectors. When Qv > 175 m3/h, both ΔT and η increased compared to the case where the collector was located at the bottom, but they were still lower than the case where the collector was in the middle location. It is obvious that the collector with the aluminum plate placed at the top had no advantage except in the cases associated with higher airflow rates. This
3.2. Effect of the airflow channel length on the type I collector The numerical results obtained from the condition of the airflow rate of 105 m3/h were taken as an example. Fig. 7 shows the theoretically calculated results of the air temperature affected by the length of the airflow channel in the collector based on a radiation intensity of 1000 W/m2 and an ambient temperature of 306 K. Under the same conditions, there were no changes in other parameters except the airflow channel length, and the air temperature of the collector outlet was increased with increases of the airflow channel length. However, after a certain length, the increasing trend of outlet air temperature slowed down. It can be observed that when the channel length was 30 m, the collector outlet temperature was 361.1 K, which was just 0.9 K higher than the case where the channel length was 25 m and at a temperature at 2 K higher than that for a channel length of 20 m. The reason is attributed to the fact that as the air flowed toward the centre of the collector, the airflow velocity increased and the surface area of the collector decreased. Therefore, the same airflow received less solar radiation, and the temperature tended to be stable. Thus, increasing the length of the channel was not an effective way to increase the temperature of the outlet air. The use of double glazing or a film on the collector cap can reduce the heat loss by increasing the thermal resistance. In addition, the heat absorber plate can be located in the middle of the collector (Fig. 2(c)) to improve the heat-transfer capacity
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Fig. 6. Measured temperatures (points) and calculated temperatures (lines) in type I collector with solar radiation intensity variations. 386
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90 773 W/m2 (type III, exp) 773 W/m2 (type III, linear fitting) 538 W/m2 (type III, exp) 538 W/m2 (type III, linear fitting) 267 W/m2 (type III, exp) 267 W/m2 (type III, linear fitting)
325
Ta = 303 K
773 W/m2 (type I, exp) 773 W/m2 (type I, linear fitting) 538 W/m2 (type I, exp) 538 W/m2 (type I, linear fitting) 267 W/m2 (type I, exp) 267 W/m2 (type I, linear fitting)
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Fig. 9. Temperature difference between the air inlet and outlet of the collector at different positions of the heat absorber plate.
340 335
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Fig. 8. Experimental temperature in type I and type III collectors with solar radiation intensity variations.
Fig. 10. Efficiency of the collector at different locations of the heat absorber plate.
is because the aluminium plate placed on the top replaced the glass cover. Without glass, the aluminium plate surface temperature (the measured value was 346 K) increased more than that at the bottom (measured value of 337.5 K). The hot face of the aluminium plate slowed down the airflow in the collector, i.e. the air near the hot surface did not flow easily at low velocities. Therefore, ΔT and η of the collector increased slowly. With the increase of the airflow rate, the heat transfer increased rapidly on the hot surface of the aluminium plate located on top, followed by abrupt ΔT and η increases. Figs. 9 and 10 indicate that the collector with an intermediate absorber plate exhibits an excellent performance regarding η and ΔT, compared to the corresponding values elicited at other locations of aluminium plates. Considering that both sides of the absorber are surrounded by fluid, which can reduce heat loss owing to the greenhouse effect, the surface temperature of the aluminium plate could rise to 340 K. The airflow was heated after it flowed through the top or the back of the heat absorber plate, and the area of the heat exchange surface inside the collector was doubled. This ultimately resulted in the enhancement of ΔT and η.
3.5. Effect of the different structures with and without absorber plate holes on the solar collector As mentioned above, the ΔT and η values of the collector increased when the aluminium plate was placed in the middle of the collector as a heat absorber plate. Combination of the two methods mentioned above allowed the conduct of experimental tests by changing the form of the heat absorber in order to examine the performance of the collector. Temperature changes in the full- and perforated-plate collectors were investigated at the same radiation intensity of 674 W/m2, ambient temperature of 291 K, and changing airflow rate. The experimental results of the temperature difference ΔT of the two types of collector structures are shown in Fig. 11. The perforated absorber shown in Fig. 4(b) was used for experiments. The experimental results of the efficiency η of the two types of collector structures are shown in Fig. 12. It can be seen from Figs. 11 and 12 that the ΔT and η of the perforated plate absorber with a certain porosity are higher than those of the full plate. Thus, the perforated plate has a better heat collecting performance than the full plate. This is because the air can only 387
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solar heat collector. The inlet and outlet temperature difference ΔT and efficiency η were higher than that of the solar collector with the full plate absorber. (3) The air temperature at the outlet of the solar collector increased as the airflow channel length increased, and the rising trend slowed down. Therefore, in the SCPPS, increasing the radius of the collector was not an effective way to improve the outlet temperature of the collector. (4) The airflow channel was subdivided into two channels, the upper air temperature was higher and suitable for power generation, the lower air temperature was moderate for agricultural production, and the heat storage material could be arranged on the ground to maintain its temperature for a long time.
25 Ta = 291 K
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The above results can be used for the optimal design of heat collection and thermal storage structures of large SCPPS. In an actual SCPPS, if the perforated plate absorber and heat storage layer are applied, the performance of the solar collector will be enhanced without occupying any other areas.
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Volume flow rate Qv (m3/h)
Fig. 11. The temperature difference between air inlet and outlet of the collector with different structural absorber versus volume flow rate.
Acknowledgements This research was financially supported by the National Natural Science Foundation of China (No. 51466001, 51266001), the Natural Science Foundation of Guangxi (No. 2017GXNSFDA198042), Guangxi Key Laboratory of Manufacturing System & Advanced Manufacturing Technology Foundation (16-380-12S003). The authors would like to thank Elsevier for improving the language quality of the revised manuscript.
65 60
Ta = 291 K I0 = 674 W/m2
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Efficiency η (%)
55 50 Full plate
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Appendix A. Supplementary material
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Supplementary data to this article can be found online at https:// doi.org/10.1016/j.applthermaleng.2018.10.068.
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References Perforated plate (second-order fitting) Full plate (second-order fitting)
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