distributed generation and alternative energy journal

SUMMER 2014

Vol. 29, No. 3

DISTRIBUTED GENERATION AND ALTERNATIVE ENERGY JOURNAL Jorge B. Wong, PhD, PE, CEM Editor-in-Chief Dr. Ing. Jose Ramos Saravia Associate Editor In this issue... 5 From the Editor: Huge Solar Plant in California versus Affordable Solar Appliances for the Poor .

7 Thermodynamic Modeling of the Solar Organic Rankine Cycle with Selected Organic Working Fluids for Cogeneration; Suresh Baral and Kyung Chun Kim 35 Design and Performance Analysis of a Solar Air Heater with High Heat Storage; Abhishek Saxena, Vineet Tirth, and Ghanshyam Srivastava

56 Development of an Empirical Model for Assessment of Solar Air Heater Performance; P.K. Choudhury and D.C. Baruah

76 Guidelines for Graphics

ISSN: 2156-3306 (print) ISSN: 2156-6550 (on-line)

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Distributed Generation and Alternative Energy Journal

Dr. Ing. Jorge Wong Dr. Ing. José Ramos-Saravia Editor-in-Chief Associate Editor WK Engineers, USA University of Zaragoza, Spain. [email protected] [email protected]

Dr. Ing. José Ramón Vega Galaz Associate Editor Monterrey Tech, Mexico [email protected]

AEE EXECUTIVE COMMITTEE 2014 Randy Haines, President, Thomas Jefferson University; Scott C. Dunning, President-Elect, University of Maine; Asit Patel, Secretary; Paul Goodman, Treasurer. Regional Vice Presidents: Region I, David Eberly; Region II, Cecil Jones; Region III, Robert Marolt; Region IV, Jared Higgins; Region V, Lori K. Moen.

DISTRIBUTED GENERATION AND ALTERNATIVE ENERGY JOURNAL, (ONLINE) ISSN 2156-6550, (PRINT) ISSN 2156-3306, was previously published as COGENERATION & DISTRIBUTED GENERATION JOURNAL. It is published quarterly by The Fairmont Press, Inc., 700 Indian Trail Rd. NW, Lilburn, GA 30047-6862. Periodicals postage is paid at Lilburn, GA and additional mailing offices. POSTMASTER: Send address changes to DISTRIBUTED GENERATION AND ALTERNATIVE GENERATION JOURNAL, 700 INDIAN TRAIL RD. NW, LILBURN, GA 30047-6862. Production Office: 700 Indian Trail, Lilburn, GA 30047 (770) 925-9388 Copyright 2014 by The Fairmont Press, 700 Indian Trail, Lilburn, GA 30047. Contributed material expresses the views of the authors, not necessarily those of the Association of Energy Engineers, the editors or publisher. While every attempt is made to assure the integrity of the material, neither the author, Association nor the publisher is accountable for errors or omissions. Individual print subscription: $306; $371 for combined institutional print and online subscriptions; $334 for institutional online only subscriptions. Back issues, when available, are $40 per copy. Special annual subscription fee of $12.50 is included in annual dues for AEE members. AEE MEMBERSHIP CHANGES Please notify Association of Energy Engineers, 4025 Pleasantdale Road, Suite 420, Atlanta, GA 30340. Tel: 770-447-5083, ext. 224, email [email protected]. SUBSCRIBE USA/North America – Taylor & Francis LLC, 325 Chestnut Street, Philadelphia, PA 19106. Tel: 215-625-8900. Fax: 215-625-2940. UK/Europe – Taylor & Francis Customer Services, Sheepen Place, Colchester, Essex CO3 3LP, UK. Tel: +44(0)-20-7017-5198. Fax: +44(0)-20-7017-5198. EDITORIAL

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STATEMENT OF OWNERSHIP, MANAGEMENT AND CIRCULATION Title of Publication: Distributed Generation and Alternative Energy Journal Frequency of Issue: Quarterly No. of Issues Published Annually: 4 Annual Subscription Price: $306 US Location of Known Office of Publication: 700 Indian Trail, Lilburn, GA 30047. Production Office: Distributed Generation and Alternative Energy Journal, 700 Indian Trail, Lilburn, GA 30047. Editorial Office: Jorge Wong, 8 Dawn Meadow Ct., Simpsonville, SC 29680, USA Owner: The Fairmont Press, Inc., 700 Indian Trail, Lilburn, GA 30047 Extent and Nature of Circulation

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From the Editor

Huge Solar Plant in California versus Affordable Solar Appliances for the Poor The largest solar concentrator power plant in the world has been built in Ivanpah, CA. SEE http://www.ivanpahsolar.com/. The total plant output is 392 megawatt, including 173,500 heliostats spread over 3,500 acres (1,420 ha). Power magazine, a trade journal traditionally focused on conventional power plants, has surprised everyone by giving this huge renewable energy installation their Plant of the Year award. In an article in Power magazine, T.W. Overton reports: The largest solar thermal plant in the world, the first large-scale concentrating solar power (CSP) project in the U.S. to employ power tower technology, and the biggest project funded to date by the Department of Energy’s (DOE’s) Loan Projects Office (LPO).

We hope Uncle Sam (US taxpayer) and investors like Google get their money back in this huge project.

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The articles in this issue of your journal, however, discuss challenges and approaches resulting from using solar power on site, at a much smaller and sustainable scale, in developing countries. Three underlying challenges are addressed in this issue from an engineering standpoint: 1.

How to deal with the fact that solar is available only during day time and sun shine hours, rain and cloud cover permitting. In addition, depending on latitude, available solar irradiation varies seasonally.

2.

How to deal with the relatively high cost and complexity (to operate and maintain) solar powered domestic appliances, given users are low income and unsophisticated (women are the main workers in rural households in the developing world).

3.

How to meet two primary needs: clean water for drinking and fuel for cooking. Italian designer Gabriella Diamanti has developed a simple, user friendly and portable solar still that turns sea water into drinking water. Source: http://www. solarbaba.com/urun

Many of the Solar Women of Totogalpa, Nicaragua use their solar box cookers to produce baked goods for sale (photo: Grupo Fenix). Source: http://solarcooking.org/newsletters/scrapr10.htm

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Thermodynamic Modeling of the Solar Organic Rankine Cycle with Selected Organic Working Fluids for Cogeneration Suresh Baral and Kyung Chun Kim ABSTRACT Fifteen (15) organic fluids were thermodynamically modeled to evaluate their fitness and performance as working fluids in an Organic Rankine Cycle (ORC) based cogeneration system. This article presents the exergy efficiency, thermal efficiency, solar power cycle efficiency, cogeneration efficiency, mass flow rate, heat input, required area of the solar collector and hot water production for the evaluated working fluids the low-temperature (90 and medium-temperature (125 solar organic Rankine cycles. Thermodynamic modeling was carried out using a commercial 1 kW scroll expander, two compact heat exchangers, a diaphragm pump and a solar collector. The article also describes the use of solar ORC technology for electricity generation and producing hot water as cogeneration. Commercial software, Engineering Equation Solver (EES), was used to calculate the operating parameters of the solar ORC. Of the 15 fluids investigated, R134a and R245fa were found to be the most appropriate working fluids for low-temperature and medium-temperature solar ORC cogeneration systems, respectively. RC318 and R123 offer attractive performance but require environmental precautions owing to their high ozone depletion potential (ODP) and high global warming potential (GWP). The article also estimates the hot water production from different working fluids for a period of one year in Busan, South Korea. Keywords Thermodynamic modeling, Solar organic Rankine cycle, Working fluids, Cogeneration, Heat source temperature, Hot water Nomenclature Q = heat (kW) W = work done (kW) A = area (m2)

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E I T Cp h s V


= exergy (kJ) = solar insolation (W m-2) = temperature (°C) = constant pressure specific heat capacity (kJ kg-1 K-1) = enthalpy (kJ kg-1) = entropy (kJ kg-1 K-1) = volume of hot water (L)

Greek Symbols η = Efficiency (%) Subscripts i = inlet th = thermal in = inlet out = outlet c = collector t = turbine p = pump hw = hot water sun = hour of sunshine o = ambient exg = exergy spc = solar power cycle cog = cogeneration INTRODUCTION The organic Rankine cycle (ORC) is similar to the cycle of a conventional steam turbine power plant but it uses a high molecular organic fluid as the working fluid. The high molecular mass allows the exploitation of effectively low temperature heat sources to produce electricity in a wide range of power outputs ranging from a few watts to several megawatts in many applications, such as geothermal, solar, desalination, ocean thermal and biomass power plants. Currently, research has been carried out extensively focusing on the full utilization of waste heat from industrial plants. Owing to environmental concerns and the depletion of fossil fuels, this energy utilization concept from low heat sources has been used widely.

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More attention is being paid to organic working fluids, which must be risk less in terms of: human health, ozone depletion, global warming contribution. This way, ORC fluids can enhance their sustainable potential as key components of distributed generation (DG) plants capable of utilizing low and medium quality heat sources (80 to 150. The selection of suitable organic fluids in the ORC plants should have the following desirable characteristics, as described by appropriate low critical temperature and pressure: small specific volume, low viscosity and surface tension, high thermal conductivity, suitable thermal stability, non-corrosive, non-toxic and compatible with engine material and lubricating oil [1,2]. Solar energy is available everywhere and is completely renewable. In remote locations, the conversion of solar energy to electricity could be an important option for enhancing the development of rural communities. The low energy density and the fluctuations of the source availability are the main obstacles in solar energy applications. Photovoltaic modules, which are commercially available technology for converting solar energy directly into electricity, have very low efficiency and require batteries for storage, which limits its application in developing countries. Few studies have reported the experimental data from an operational solar ORC system. A performance and design optimization study by Quoilin et al. [3] based on low cost solar ORCs provided the modeling results for solar thermal electricity generation in the remote off-grid areas of developing countries, and revealed an overall thermal efficiency of 8%. Saitoh et al. [4] reported a solar ORC efficiency of 7% and with the working fluid, R113, and an expander efficiency of 63%. Jing et al. [5] analyzed a combination of ORC with compound parabolic concentrators (CPCs) for electricity generation with the working fluid, HCFC-123. They suggested that heat exchangers with two thermal oil cycles can improve the collector efficiency by 8.1-20.9% with a maximum evaporating temperature of the working fluid of 120°C. He et al. [6] simulated a parabolic trough solar energy generation system, and examined the optimal feeding fluid temperature, suitable number of heat recovery series, thermal storage capacity, and system performance with three different working fluids, R113, R123 and pentane. Torres and Rodriguez [7] provided a detailed analysis of low power solar-driven Rankine cycles with working fluids, such as toluene, octamethylcyclotetrasiloxane (D4) and hexamethyldisiloxane (MM) in direct solar vapor generation configuration of solar ORC applied in the desalination process, and have a global efficiency of 17.3%, 15.7% and 15%, respectively when the condensation temperature was set to

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35°C. Gang et al. [8] examined the regenerative and without regenerative solar ORC cycle, and reported an efficiency of approximately 8.6% and 4.9% respectively for irradiance 750 W/m2. Wang et al. [9], using R245fa as the working fluid and a flat-plate solar collector, reported a maximum system efficiency of approximately 3%, where the time-averaged Rankine cycle efficiency was approximately 2.8% in the morning and 5.3% in the afternoon. Twomey et al. [10] simulated a solar ORC and revealed an efficiency of 3.4% with a maximum expander isentropic efficiency of 59%. Manolakos et al. [11] provided the design and experimental results from a 2 kWe solar ORC system using evacuated tube collectors, scroll expanders and R134a as the working fluid, and reported an efficiency of 4% applied in a reverse osmosis process in desalination. Pedro et al. [12] examined the exhaust waste heat recovery potential of a micro-turbine using an ORC, where different dry organic fluids were considered to find the exergy efficiency. Currently, 1.6 billion people all around the world still have no access to electricity. Huge communities in underdeveloped countries do not have a centralized grid connected to their main lines of electricity. Therefore, conversion techniques from low heat source to electricity in the underdeveloped countries, which require rural electrification, are needed. Several groups including Solar Turbine Group (STG) have tried to implement small scale solar thermal technology using medium temperature collectors aimed at replacing or complementing diesel generators in the off grid areas in underdeveloped countries, thereby generating clean and green power at very low installation and maintenance cost. The present work aims to select most suitable working fluids for lowtemperature (90 and medium-temperature (125 solar ORC systems to produce both electricity and hot water. Characteristics of 15 potential working fluids are evaluated and compared for a 1 kW micro-power system. SYSTEM MODELING The present solar ORC system consisted of a working fluid pump (diaphragm type), two compact heat exchangers and a commercial scroll expander, as shown in Figure 1. The working principle is explained as the hot water obtained by the solar collector is passed through the evaporator. The ORC working fluid is pumped and passed through the evaporator, where it changes its phase. The working fluid is expanded into the scroll expander

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to generate electricity. The working fluid at the turbine outlet is condensed in the condenser by cold water supplied from a tap and flows back into the circulation pump to begin a repeated cycle. The condensing temperature determines the temperature of hot water production as cogeneration. A simulation and thermodynamic modeling of the present system was carried out to predict the performance characteristics of the solar organic Rankine cycle under a range of conditions and parameters. The graphical representation of computation modules organized by direction of energy flow can be shown in Figure 2. The system equations were derived easily from the energy and mass balance for a control volume. The exergy of the ORC system was calculated using the enthalpy and entropy of the evaporator inlet and the dead state (To = 25°C). The system equations obtained can be expressed as Evaporator:

Qin = h4 – h1 (1)

Condenser:

Qout = h3 – h2 (2)

Turbine:

(3)

Pump:

(4)

Power output:

(5)

Wout Thermal efficiency: ηORC = ——— Qin Solar power cycle efficiency:

(6) (7)

Cogeneration efficiency:

(8)

Exergy state point:

(9)

Exergy efficiency:

(10)

Area of collector: Ac =

(11)

Hot water production:

(12)

Figure 2. Flow chart of simulation with input parameters and output quantities.

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The present work compared several ORC working fluids for system efficiencies, appropriate area of the solar collector and production of hot water as cogeneration for 1 kW power output. Several authors [13-18] worked on analyzing the characteristics of the different working fluids for the conversion of low grade heat. For the simulation and analysis of ORC working fluids as potential candidates in solar applications, the critical temperature of fluids, ranging from 95°C to 240°C, were selected. For low-temperature solar ORC using a flat-plate solar collector, the critical temperature of the working fluids ranged from 95°C to 135°C, whereas the critical temperature for medium-temperature solar ORC using a vacuum tube type solar collector ranged from 150°C to 240°C. These critical temperature ranges were taken as a fundamental basis because of its thermophysical behavior at various heat source temperatures. For the simulation, the desired thermodynamic properties of the ORC working fluid at different state points should be known. The thermodynamic properties of an ORC working fluid can be described best by the energy equation and calculated using software known as an Energy Equation Solver (EES). Therefore, all the thermodynamics properties of ORC working fluids in this study were obtained using commercial software EES. For simplicity, several reasonable assumptions were implemented by thermodynamic modeling. The assumptions are listed as follows: a. The electric power at the expander is fixed at 1 kW. b. The evaporating temperature for low and medium temperature solar ORC are supposed to be 85°C and 120°C, respectively, with a pinch point temperature difference of 5°C. c. The condensation temperature is fixed to 45°C. d. The efficiency of the expander is 70%. e. The efficiency of the pump is assumed to be 70%. f. The solar collector efficiency is presumed to be 70%. g. The internal irreversibility is ignored. h. The pressure drops in the components other than the expander are ignored. For low-temperature and medium-temperature solar ORCs, the heat source is expected to be 90°C and 125°C, respectively.

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RESULTS AND DISCUSSION In total, 15 pure organic working fluids were selected as potential candidates, as listed in Table 1. Only one criterion was considered at this first step, a critical temperature above 90°C. Hot water was provided from the solar collectors. The condenser was cooled with tap water at 25°C. In this study, the vapor at the turbine inlet was saturated. A re-heater was not introduced because it could add too much expense to the system. Table 2 lists the system performance results for a design of 1kW power output. This study examined the ORC system located in Busan, South Korea (latitude and longitude of 35.15°N and 129.06°E, respectively). Efficiencies The ORC efficiency is the most important index used to evaluate the performance of the system. Nine and six working fluids emerged as potential candidates for the low-temperature and medium-temperature solar ORC respectively. The exergy efficiency is an indicator of how close the thermal efficiency is to the highest value permissible depending on the temperature range (Carnot efficiency). The system exergy efficiency ranged from 29.29% to 50.03%. The maximum exergy efficiency was obtained using RC318 and R123, as shown in Table 2. This is because these Table 1. Properties of the organic fluids along with the safety and environmental parameters

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fluids rise at low working pressures as the enthalpy difference in the expander increases. Lower exergy efficiency means higher system irreversibility. Normally, in an ORC system, the evaporator makes the largest contribution to the overall irreversibility followed by the expander. R500, the lowest exergy efficiency fluid, requires the largest heat input due to the large irreversibility in the evaporator. The thermal efficiency and solar power cycle efficiency ranged from 5.38% to 12.58% and 3.67% to 8.8%, respectively, as shown in Figures 3 and 4. The thermal efficiency was based on the first law of thermodynamics and is the typical efficiency of an ORC. The solar power cycle efficiency considers the efficiency of the solar collectors. Therefore, the solar power cycle efficiency is always 2 ~ 3% lower than the thermal efficiency. This efficiency can be compared directly with photovoltaic devices in terms of electric power generation. On the other hand, the system efficiency of a solar ORC with cogeneration should be considered using hot water from the condenser. The cogeneration efficiency ranged 73.11% to 77.92% for the low-temperature ORC whereas 80.24% to 82.24% for the medium-temperature solar ORC. A cogeneration unit should utilize the high thermal efficiency of solar thermal collectors to provide water heating, while generating electricity during the off-peak periods when the sun’s potential would otherwise be wasted. Table 2. Comparison of the performances of different working fluids for a 1 kW power output.

Figure 4. System exergy, thermal and solar power cycle efficiency for medium-temperature solar ORC working fluids at a condensing temperature of 45°C.

Figure 3. System exergy, thermal and solar power cycle efficiency for low-temperature solar ORC working fluids at condensing temperature of 45°C.


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Variations of the exergy efficiencies of the low-temperature and the medium-temperature solar ORC with respect to the turbine inlet temperature are shown in Figures 5 and 6. The exergy efficiencies are increased with increasing turbine inlet temperature. The pinch point temperature was maintained as a constant and the vapor at the expander inlet became saturated. In the case of low-temperature solar ORC, RC318 shows the maximum exergy efficiency up to 47% at 90°C of the turbine inlet temperature followed by R600a and R717. Other six fluids show similar exergy efficiency in the range of 27 to 33% at 90°C of the turbine inlet temperature. In the case of medium-temperature solar ORC, R123 shows the maximum exergy efficiency. However, the difference is not pronounced compared with the low-temperature case. The exergy efficiency of R245fa shows a unique pattern. The maximum exergy efficiency is saturated when the turbine inlet temperature exceeds 120°C. This nature suggests the existence of an optimal operating condition. Effects of the turbine inlet pressure on the exergy efficiency can be seen in Figures 7 and 8. The system exergy efficiency increased with increasing turbine inlet pressure. In the case of low-temperature solar ORC,

Figure 5. Variation of the exergy efficiency for low-temperature solar ORC working fluids with the turbine inlet temperature at the condensing temperature of 45°C.

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Figure 6. Variation of the exergy efficiency for medium-temperature solar ORC working fluids with the turbine inlet temperature at a condensing temperature of 45°C.

RC318 shows the maximum exergy efficiency 38 to 55% for the range of turbine inlet pressure from 1000 to 2500 kPa. The high exergy efficiency in the low-temperature ORC with RC318 attributes to high molecular mass compared to other working fluids. Other eight working fluids show a similar range of exergy efficiency with different turbine inlet pressure ranges as shown in Figure 7. In the case of medium-temperature solar ORC, R123 shows the best performance for the range of 500 to 1800 kPa turbine inlet pressure. Interestingly, the exergy efficiencies of R245fa are gradually increased with increasing turbine inlet pressure for the wide range of pressure variation, 870 to 2900 kPa as shown in Figure 8. The exergy efficiencies of methanol and ethanol increase in a short range 0.2 to 1.0 MPa. Pressure Ratio The cycle pressure ratio determines the type of expander size and number in the ORC system. A high pressure ratio also increases the efficiency of the system. On the other hand, a higher pressure ratio accounts for the installation of a number of expanders arranged in series or parallel

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Figure 7. Exergy efficiency versus turbine inlet pressure for low-temperature solar ORC working fluids at a condensing temperature of 45°C

Figure 8. Exergy efficiency versus turbine inlet pressure for medium-temperature solar ORC working fluids at a condensing temperature of 45°C.

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in an ORC system. An excessively high pressure in the evaporator and an excessively low pressure in the condenser should be avoided due to structural problem and leakage. According to Tchanche et al. [18], good pressure values were in the range of 0.1 – 2.5 MPa, and a pressure ratio of approximately 3.5 is reasonable for a single stage expander. From Table 2, R227ea, RC318 and R600a, R123, methanol and ethanol have low condensing pressure, meaning that they have a higher pressure ratio. R22, R290, R500, and R717 have higher pressure above 3 MPa in the evaporator. A high turbine inlet temperature yields a high pressure ratio, which can be observed from Figures 9 and 10 when the temperature is varied from 65°C to 140°C. The ORC fluids with good condensing and evaporating pressures are R134a, R227ea, RC318, and R600a for the low-temperature solar ORC system, and R245fa, R123, R11, and R141b for the mediumtemperature solar ORC system. Flow Rates For economic aspects, the turbine outlet volume flow rate plays an important role because it determines the size and cost. As shown in Table 2, R12, R500, R600a, methanol and ethanol exhibited a high volume flow rate. Fluids with a low volume flow rate are preferable for decreasing the pump size and friction loss in a pipe system. Among these are R245fa,

Figure 9. Variation of the pressure ratio for low-temperature solar ORC working fluids with turbine inlet temperature at a condensing temperature of 45°C.

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Figure 10. Variation of the pressure ratio for medium-temperature solar ORC working fluids with turbine inlet temperature at a condensing temperature of 45°C.

R717, R123, R22, RC318, R227ea, and R134a. In general, the turbine outlet volume flow rate is inversely proportional to the turbine inlet temperature, which can be illustrated in Figures 11 and 12. For the low-temperature solar ORC, R717 shows the best performance followed by R22, RC318 and R134a. For the medium-temperature solar ORC, R245fa shows the lowest turbine inlet volume flow rate followed by R123 and R11. The proper introduction of a working fluid mass flow rate is necessary for the expansion process. The pump selection is also another fundamental aspect for appropriate mass flow rate pumping. From Table 2, R290, RC318, R717 and R600a yielded a low mass flow rate for a lowtemperature solar ORC. This is useful because they also require a low heat input. For medium-temperature solar ORC, methanol and ethanol exhibited a similar low mass flow rate because of their similar properties. From Table 2, methanol and ethanol yielded the lowest maximum pressures and highest enthalpy heat of evaporation. For the economic condition, R290, RC318, R600a, methanol and ethanol are feasible for a large capacity ORC system.

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Figure 11. Variation of turbine outlet volume flow rate for low-temperature solar ORC working fluids with a turbine inlet temperature at a condensing temperature of 45°C.

Figure 12. Variation of turbine outlet volume flow rate for medium-temperature solar ORC working fluids with a turbine inlet temperature at a condensing temperature of 45°C.

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Solar Collectors

The heat input to the system is significant in a solar ORC that determines the size of the collector and establishes a major part of the system cost. Therefore, solar applications will be more viable with fluids for which the amount of heat required is small. From Table 2, the heat required for a 1 kW power output falls in the range between 12-20 kW and 6-19 kW for the low-temperature and medium-temperature solar ORC system, respectively. In addition, at a higher turbine inlet temperature, the system requires less heat input, which can be explained by Figures 13 and 14. High temperature saturated vapors reduce the heat input. When designing a solar ORC, one could choose between a system with a larger collector area-low temperature and a system with a small collector areahigh operating temperature depending on the application. The area of solar collector in ORC system is determined by the heat input, net work done and solar insolation of the location for installation. From Figures 13 and 14, a high turbine inlet temperature requires less area for the solar collector for the system. For a 1 kW power output, the required area of the collector is in the range, 23-34 m2, for the low-

Figure 13. Variation of the area of collector for low-temperature solar ORC working fluids with the turbine inlet temperature at a condensing temperature of 45°C.

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Figure 14. Variation of the area of collector for medium-temperature solar ORC working fluids with the turbine inlet temperature at a condensing temperature of 45°C.

temperature, whereas for the medium-temperature solar ORC, the range is between 14-17 m2, as listed in Table 2. The collector area also constitutes a major part of the system cost. Figure 13 shows that the lowest required area of solar collector is in order RC318, R600a, and R717 in the lowtemperature ORC. The lowest heat input fluid is R123 followed by R245fa, methanol, ethanol, R141b, and R11 for the medium-temperature solar ORC system as shown in Figure 14. Hot Water Production For the production of hot water, the area of the collector plays a vital role. The other factors include solar radiation, ambient temperature and solar collector efficiency. Hot water can be provided with a condensing temperature from the condenser of an ORC system. This concept of cogeneration can be applied widely when there is plenty of water in the located site of the ORC. In the present study, the obtained hot water temperature was fixed to 45°C. Table 3 lists estimates of the average daily hot water production at different months of the year from the required area of solar collector for working fluids. The meteorological conditions data of Busan,

Table 3. Hot water production for different working fluids in a year in Busan, South Korea.

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South Korea were obtained from Korea Meteorological Administration [19]. From this, a larger collector area and higher solar insolation produces more hot water. The production reached a maximum at April, May, June, July and August. The highest production is for R500, which has a collector area of 30.05 m2 and produces 4087 liters per day of hot water followed by R290, R22 and R134a in a low-temperature solar ORC, whereas in the medium-temperature solar ORC, the highest production was obtained using R11, which is 1949 liters per day followed in order by R245fa, ethanol, methanol and R123. The hot water produced is stored in the insulated storage tank. The application of hot water in rural areas of developing countries are washing utensils, bathing, washing clothes as well as in small community-based hospitals, health posts and clinics. Verification and Validation For the verification and validation of simulation model the results were compared with the experimented data. The experimental data were obtained from a small scale ORC system using a scroll expander developed in our laboratory [20]. The ORC system used R245fa as the working fluid. The experiment was carried out with the evaporating pressure of 2520 kPa and condensing temperature of 45°C. The experiment was carried out with the setting mass flow at different rates. The heat source was hot water generated by an electric heater whose temperature was 130°C [20]. Figure 15 depicts a comparison of the simulated results and the experimental data. In general, both results agree well within the uncertainty range of experiment. However, the simulated efficiency is slightly higher than that of experiment. It can be explained that the internal losses of the real piping system and heat exchanger pressure losses are not taken into accounted in the system while carrying out the simulation. The other reason attributed to the expander’s efficiency being fixed for whole calculation. In the real case, the expander’s efficiency is not always constant. Additionally, there occurs irreversibility in the system. The trend of the simulated ORC system thermal efficiency versus working fluid mass flow rate shows a good agreement with the experimental data. The average deviation between experiment and simulation is 3.5% while the maximum deviation is 4.5% in correspondence of the higher values of working fluid mass flow rate. Moreover, it can be observed that ORC system efficiency values are in the range 5%-8%, in agreement with literature data for ORC systems characterized by a net electric power output in order of few kW. The experimental and simulated thermal efficiency follow the same be-

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havior, even though the simulation model has a few simplifying assumptions as discussed earlier. Furthermore, it was revealed that the increase in working fluid mass flow rate resulted in higher thermal efficiency and shaft power output.

Figure 15. Comparison of thermal efficiency predicted by simulation and experimental data obtained in a small scale ORC system [20] for the validation of mathematical models.

Working Fluids Selection According to the operating conditions and various parameters for the estimation of a 1 kW solar power output ORC, most of the working fluids were not accepted, even though the simulated results yielded high thermal efficiencies and pressure ratio. This is because of safety, toxicity, flammability and environmental characteristics of the working fluids. The safety criteria cannot be neglected. The terms of safety according to the ASHRAE are A1 (non-flammable and non- toxic), A2 (lower flammability and non-toxic), A3 (flammable and non-toxic), B1 (non-flammable but toxic), B2 (toxic and low flammability) and B3 (high toxic and high flammability). In Table 4, the decision rating criteria for working fluids have

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been recommended according to the manufacturer’s operating conditions of a 1 kW power scroll expander, average value of exergy, thermal, solar cycle efficiencies, area of the solar collector, hot water production along with the minimal safety and environmental considerations. For the value of the parameter preferring the working fluid, it is marked as “√,” and “x” vice versa. According the decision criteria rating in Table 4, the following working fluids were rejected, which are listed in Table 5: high ozone depletion potential (ODP) (R22, R500, R123, R11, R141b), high global warming potential (GWP) (R227ea, RC318), low safety (R290), high pressure ratio (Methanol, Ethanol), large collector area (R12), and low hot water production (R717,R600a). Only two working fluids (R134a and R245fa) were finally acceptable for operating in a low-temperature and medium-temperature solar ORC system, whose heat source temperature are 90°C and 125°C, respectively. CONCLUSIONS A 1 kW solar organic Rankine cycle system was studied and modeled thermodynamically with 15 selected organic working fluids for low-temperature and medium-temperature heat sources based on its critical temperature, ranging from 150°C to 240°C. The analysis involved comparing various parameters, such as exergy, thermal, solar power cycle efficiencies in addition to the pressure ratio, mass flow rate, heat input, turbine inlet volume flow rate, the required area of solar collector and hot water production. The maximum exergy efficiency was 48.19% and 50.03% for RC318 and R123, respectively. The maximum turbine inlet pressure was 4037 kPa and 1929 kPa for R22 and R245fa, respectively. The maximum pressure ratio accounts for RC318 and ethanol. RC318 and R123 had the minimum heat inputs required. Ideally, a high efficiency, high pressure ratio, high hot water production, low mass flow rate and small area of collector are appropriate. On the other hand, many working fluids are rejected because of other factors, such as flammability, toxicity, and environmental conditions. Therefore, the most suitable working fluids are R134a and R245fa for low-temperature and medium-temperature solar ORCs, respectively. These recommended working fluids require an appropriate collector area size and the production of hot water is reasonable. This solar ORC can be mostly suitable for remote places of developing countries that lack electricity. Therefore, this technology is expected to be

HIGH PERFORMANCE BUILDINGS: A GUIDE FOR OWNERS & MANAGERS

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High Performance Buildings: A Guide for Owners and Managers, is a template - a blueprint for action for those making decisions about how to improve the energy efficiency and performance of new or existing buildings. It is designed to have broad appeal, both for the seasoned veteran facility or energy manager and for the new manager alike, but can also be utilized as a practical desk reference by professionals such as architects, engineers, and construction managers. The book provides an objective and orderly approach to what is often a complex, costly and time-consuming process. The full spectrum of topics relevant to achieving optimum building performance is addressed, including analysis of overall building energy use and performance, building commissioning, applicable codes, standards and rating systems, building envelope, onsite power generating options, optimizing performance of building mechanical and electrical equipment, and importance of effective building operation and maintenance practices. It is thorough in its topical scope, technically accurate, yet concise. Fundamental principles are discussed and illustrated with case studies.

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ISBN: 0-88173-646-5

1 - Inputs 2 - Definitions 3 - High Performance - A Real Estate Perspective 4 - Building Energy Analysis 5 - Commissioning 6 - Standards, Codes & Rating Systems 7 - Envelope 8 - On-site Distributed Power (DEG) 9 - Mechanical & Electrical Equipment 10 - Operations & Maintenance 11 - Outputs

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CREATING A STRATEGIC ENERGY REDUCTION PLAN

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Scott Offermann

This book provides a simple, easily followed process for auditing building operations in order to identify and reduce energy consumption. The crucial steps of this process involve assessing the facility’s current conditions, understanding and analyzing the operational and cost-based opportunities, reporting the findings, and then documenting the plan. The book discusses the full scope of building components and systems, including how each impacts energy efficiency, and then goes on to describe the operational energy efficiencies that can be gained by implementing no-cost changes or alternative maintenance activities already funded. Capital improvement opportunities, along with evaluating return on investment and life cycle replacement of equipment are also covered. The four-step process described in this book will serve as a valuable tool for every building operator seeking to improve facility energy performance.


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Table 4: Decision rating criteria for the selection of working fluids

Vol. 29, No. 3 2014 31

Table 5: Decision criteria table for the selection of working fluids


Vol. 29, No. 3

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33

used for small distributed power generations systems and producing hot water with the same unit. Acknowledgements This study was supported by the National Research Foundation of Korea (NRF) grant funded by the Korea government (MSIP) through GCRC-SOP (No. 2011-0030663) and by the Korea Institute of Energy Technology Evaluation and Planning (KETEP) grant funded by MSIP (No. 2011201010030-11-2-300). References

1. Chen H, Goswami DY, Stefanakos EK. A review of thermodynamic cycles and working fluids for the conversion of low-grade heat. Renewable and Sustainable Energy Reviews 2010; 14:3059-3067. 2. Tchanche BF, Lambrinos G, Frangoudakis A, Papadakis G. Low-grade heat conversion into power using organic Rankine cycles -A review of various applications. Renewable and Sustainable Energy Reviews 2011; 15:3963-3979. 3. Quoilin S, Orosz M, Hemond H, Lemort V. Performance and design optimization of a low-cost solar organic Rankine cycle for remote power generation. Solar Energy 2011; 85:955-966. 4. Saitoh T, Yamada N, Wakashima SI. Solar Rankine cycle system using scroll expander. Journal of Environment and Engineering 2007; 2:708-719. 5. Jing L, Gang P, Jie J. Optimization of low temperature solar thermal electric generation with Organic Rankine Cycle in different areas. Applied Energy 2010; 87:33553365. 6. He YL, Mei DH, Tao WQ, Yang WW, Liu HL. Simulation of the parabolic trough solar energy generation system with Organic Rankine Cycle. Applied Energy 2012; 97:630-641. 7. Delgado-Torres AM, García-Rodríguez L. Preliminary assessment of solar organic Rankine cycles for driving a desalination system. Desalination 2007; 216:252-275. 8. Pei G, Li J, Ji J. Analysis of low temperature solar thermal electric generation using regenerative Organic Rankine Cycle. Applied Thermal Engineering; 30: 998-1004. 9. Wang XD, Zhao L, Wang, JL. Experimental investigation on the low-temperature solar Rankine cycle system using R245fa. Energy Conversion and Management 2011; 52:946-952. 10. Twomey B, Jacobs PA, Gurgenci H. Dynamic performance estimation of small-scale solar cogeneration with an organic Rankine cycle using a scroll expander. Applied Thermal Engineering 2013; 51:1307-1316. 11. Manolakos D, Papadakis G, Kyritsis S, Bouzianas K. Experimental evaluation of an autonomous low-temperature solar Rankine cycle system for reverse osmosis desalination. Desalination 2007; 203:366-374. 12. Mago PJ, Luck R. Energetic and exergetic analysis of waste heat recovery from a microturbine using organic Rankine cycles. International Journal of Energy Research 2013; 30:926-938. 13. Bao J, Zhao L. A review of working fluid and expander selections for organic Rankine cycle. Renewable and Sustainable Energy Reviews 2013;24:325-342. 14. RayeganR, Tao YX. A procedure to select working fluids for Solar Organic Rankine Cycles (ORCs). Renewable Energy 2011; 36:659-670.

34

Distributed Generation and Alternative Energy Journal 15. Gao H, Liu C, He C, Xu X, Wu S, Li. Performance Analysis and Working Fluid Selection of a Supercritical Organic Rankine Cycle for Low Grade Waste Heat Recovery. Energies 2012; 5:3233-3247. 16. Saleh B, Koglbauer G, Wendland M, Fischer J. Working fluids for low-temperature organic Rankine cycles, Energy 2007; 32:1210-1221. 17. Maizza V, Maizza A. Unconventional working fluids in organic Rankine-cycles for waste energy recovery systems. Applied Thermal Engineering 2001; 21:381-390. 18. Tchanche BF, Papadakis G, Lambrinos G, Frangoudakis A. Fluid selection for a low-temperature solar organic Rankine cycle, Applied Thermal Engineering 2009; 29:2468-2476. 19. Korea Meteorological Administration. Monthly and seasonal climate summary, 2013. Available online: http://www. web.kma.go.kr (accessed 08.13.13). 20. Yun E, Kim HD, Yoon SY, Kim KC. Development and characterization of smallscale ORC system using scroll expander. Applied Mechanics and Materials 2013; 291294;1627-1630.

———————————————————————————————— ABOUT THE AUTHORS Suresh Baral is a PhD student in the School of Mechanical Engineering, Pusan National University, South Korea. His advisor is Prof. KC Kim. Mr. Baral received his bachelor and master degree in Mechanical Engineering from Institute of Engineering, Pulchowk Campus and Kathmandu University, Nepal respectively. He is permanent faculty member in Pokhara University, Nepal as a Lecturer. His areas of interests are renewable energy, energy conversion technology and Finite Element Methods. He can be reached at [email protected] Kyung Chun Kim (corresponding author) is a Professor in the School of Mechanical Engineering of Pusan National University in Korea. He obtained the Ph.D. degree from the Korea Advanced Institute of Science and Technology (KAIST), Korea, in 1987. He was selected as a member of the National Academy of Engineering of Korea in 2004. His research interests include 3D3C Micro-PIV, Bio-MEMS, turbulent flow measurements based on PIV/LIF, biomedical engineering, POCT development, wind turbines, and organic Rankine cycle system. He can be reached at [email protected]

Vol. 29, No. 3

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Design and Performance Analysis of a Solar Air Heater With High Heat Storage Abhishek Saxena, Vineet Tirth, and Ghanshyam Srivastava

ABSTRACT The aims of this work is to enhance the efficiency of a simple designed solar air heater for crop drying and space heating. ‘Desert sand’, has been introduced as a heat absorbing media inside the solar heater. The experimental testing has been carried out on four different configurations by operating it on natural and forced convection in the climatic conditions of Moradabad, India. The thermal behavior of the system is evaluated by operating it on auxiliary power by placing a halogen lamp tube (300W) inside the inlet and outlet ducts. Because of using halogen lights the system is feasible to perform in poor ambient conditions. The thermal performance of all new configurations of the modified system was found better over a similarly designed conventional solar air heater. Keywords: air heater, desert sand, heat storage, solar energy performance, thermal energy storage INTRODUCTION Since solar radiation is inherently time-of-day dependent, storage of energy is essential if solar is to meet energy needs at night or during daytime periods of cloud cover and make a significant contribution to total energy needs. Since radiant energy can be converted into a variety of forms, energy may be stored as thermal, chemical, kinetic, and potential energy. Generally, the choice of the storage media is related to the end use of the energy and the process, storage as thermal energy is often cost effective. The optimum capacity of the storage device for a given solar process depends on the time dependence of the solar availability,

36


the nature of the load, the cost of auxiliary energy, and the prices of the process components. These factors must all be weighed carefully for a particular application to arrive at the system design which minimizes the final cost of delivering energy [1-3]. The system basically consists of a flat plate collector (FPC) tilted at some angle from the horizontal with a blower to transfer heat from the collector to storage. All the solar energy provides goes through the storage unit, may be a tank of water or a bin of dry crushed rock. The stored energy could be used for either building heating or water heating [4]. All materials absorb, store heat release heat as they become heat up or cool down. Materials like water or stone will absorb more heat for a fixed temperature rise than straw or wood. Heavy materials can store large quantities of heat without becoming too hot. When temperatures around them drop, the stored heat is released and the materials themselves cool down. This heat storage capacity of various materials can be used to store the sun’s heat for later use. Solar energy (a viable source for

Vol. 29, No. 3

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cooking, drying and heating) penetrates through walls or glazing to the interior of a body or system. This solar heat is absorbed in the air and inside surrounding materials. The air in the system is likely to heat up first. It then distributes this heat to the surrounding materials via convection [5]. Most collectors employ a black-painted, flat absorber plate with heat-transfer passages built within, above or below it and with one or more glass covers on the top. The heat transfer in solar collector takes place by simultaneous radiation, convection, and conduction [6]. In view of rising energy prices and an increasing share of power generated by renewable energy sources, the importance of energy storage is growing. In the framework of this work, a thermal energy storage concept for solar air heater is being developed, in which desert sand serves as a storage medium. The desert is suitable due to its properties such as high thermal stability, specific heat capacity, and low-cost availability. Compared with storages based on ceramic bodies, the use of sand promises to reduce costs of energy storage and thus to reduce the costs of heating. A solar air heater (SAH) is a specific type of heat exchanger which transfers heat to air, which is obtained from absorbing solar radiation by an absorber. During solar air heating heat transfer occurs from an energy source which spreads radiation in the air. It consists of an absorber plate (an Aluminium or Al made sheet that converts the solar energy into heat energy, fixed in such a way as to form an air duct through which the air is circulated and heated), supportive walls, ducts or channels for fluid flow (G.I or Al sheets of which one ‘end’ is used to carry the fresh air from the outside to the heat collector. The other ‘end’ is of large area carrying hot air from the heat collector to the required place), glazing (a transparent float glass that minimizes convective and radiative losses to the atmosphere and obtains solar radiation to stay between absorber and glazing, and to be absorbed by blackened absorber), air blower or fans (to supply air at high velocity for forced convection), and insulation (to minimize heat losses to the environment). Almost parts of solar air heaters are thermally well insulated to reduce thermal heat losses. These air ducts (with two open ends) are made of SAHS are generally used to dry agricultural products, to dry fabrics and space heating. Drying grains, fruits, vegetables, tea, and building heating are a few examples of this [7-8]. Since solar radiant energy can be converted into a variety of forms of energy storage as thermal, chemical, kinetic or potential energy.

38


Commonly, the choice of the storage media is related to the end use of the energy and the process employed to meet that application. In the thermal conversion process, stored as thermal energy, it is often most cost effective [9]. Thermal Energy Storage (TES) can be classified as sensible heat in hot liquids and solids, as latent heat in melts and vapor, as thermochemical heat appearing in chemical reactions, and as sorption heat in adsorption processes. Thermal energy is stored by raising the temperature of a solid or a liquid medium by using its heat capacity. Apart of this, LHS uses the latent heat of the material to store thermal energy. Latent heat is the amount of heat absorbed or released during the change of the material from one phase to another. The amount of thermal energy stored in the form of sensible heat or latent heat in the material can be calculated by reference [10]. In SAHS, it is essential to provide a suitable thermal storage in adding together to the solar collector to enhance the heat transfer and storage capacity. Both latent and sensible heat storages use as a heat transfer enhancing media to solar energy collectors. Phase change materials are used widely as a TES in many solar applications. It has been originated by studying the available literature that the rock bed, pecked bed and paraffin wax are commonly used in solar air heaters as TES. In several previous works, a few good improvements are found in the thermal behavior of SAHS by using of these heat storing materials (table 1). An attempt is made here for an optimal design of SAH that is simple in design, user friendly, easy to maintain, can be performed in poor ambient conditions, inexpensive and have improved efficiency over than other previous designs. All the works and experiments have been carried out previously [11-23] are typical in design and costly in comparison of our model especially in the case of SAH with heat storages. Besides this almost available materials used as thermal heat storage used in solar energy equipments are much costlier to purchase and higher production cost [10]. In the present model, the desert sand is used as a TES, which is purchased from the local market @ of 0.1 $/kg. Near about 3 kg of the desert sand was used in the system and the cost was 0.3 $- approximately, which is very less than the cost of any of other thermal heat storage materials used for solar air heaters.

Table 1. Improvements in the performance of solar air heaters by using different heat storing materials

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EXPERIMENTAL SET UP Two solar air heaters (H1 and H2) of same specific dimensions (Figure 1a) have been designed and fabricated to supply hot air for crop drying and space heating purposely. Both the air heaters were experimentally tested individually for their thermal performance on the 4 different configurations. Plywood of 1 cm thickness was used for the fabrication of both the heaters. The precise area of absorber tray was 151 x 53 cm2 and made of 0.5 mm thick 22 SWG Al sheet. The absorber plate was painted dull black to store the maximum amount of solar energy for H1. To reduce the heat losses, a 2 cm thick layer of glass-wool (insulator) was inserted between the absorber tray and outer cabinet. A single pane transparent float glass of 0.3 cm thickness and 151 x 70 cm2 was used as glazing to allow the solar radiation inside the solar heaters. The single glazing was considered especially for the maintenance of the SAHS. The efficiency effecting elements of SAH such as; halogen lights, inside wall of ducts (for good reflection), and principally the absorber plate are required to be very clean while performing (a float glass, beneath which the desert sand is spread to be high SHS) as well for efficient working in poor ambient conditions. Dusty transparent glass over storage media and dusty reflective walls will be resulted in lower efficiency of the solar energy systems [24]. Double glazing makes the system a little bit complicated in comparison of discussing system and more maintenance and attention will be requisite. The distance between glazing and absorber plate was 10 cm for both the heaters. A layer of desert (Figure 1 b) of 1 mm was spread on the absorber plate and sealed by 2 mm float glass for H2 (Figure 2). The side walls of SAH were tilted at 115o angle because of receiving good amount of solar radiation through its exposed area. For the supply of the air 2 similar fine velocity table fans (40W) of ‘CINNI- 20 CR’ trademark were used to supply the air to the system. The speed of air supplying was set at 2 m/s for forced convection. This speed of the fan can be controlled by a knob which was fixed on 3 inbuilt controls viz; low speed -1750 rpm, mid speed-1980 rpm, and high speed-2200 rpm. Even this speed can also be varied up to 2500 rpm through a high voltage of current supply. A cylindrical tapered shaped vessel was used to connect the fan directly to the inlet duct of SAH to pass the fresh air {thermal conductivity- 0.26 (W/mK), specific heat-1005 (J/kgK), density-1.2 (kg/m3)} inside the SAH. This vessel was fabricated of an Al bucket (0.5 mm thick) with

Vol. 29, No. 3

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two ‘Ends’ in which round shaped end was fixed with air supplying fan while rectangular end was fixed to the inlet duct of SAH. Other two ducts (inlet and outlet) of same dimensions were fitted to the SAH (fabricated with the same grade of Al and 0.5 mm thick) for air supply and for exhausting. Both ducts were also painted dull black from outside. The ducts were remained as it is (with no painted walls) inside which results in increasing the temperature of the duct because of the high reflectivity produced by using halogen lights (placed inside the ducts) while operating. A rigid frame of galvanized iron was used to support the SAH at the ground level. The system was placed towards the southward at an angle of 43o from horizontal. For a HHTS a thin layer of 1 mm desert was spread above on the blackened surface i.e. absorber tray. The desert sand is taken because of good thermal heat storage {Density (kg/m3)→ 1450, thermal conductivity (W/m-K)→ 0.26, specific heat (kJ/kg-K)→ 0.80, thermal diffusivity (m2/s/106)→ 0.35, heat capacity (J/m3k x10-6)→ 1.28, emissivity→ 0.91 and absorbitivity→ 0.95, porosity→ 26%} for high heat absorber and performs as SHS [25]. The test desert sand was sieved to 20 x 50 (US Sieve) mesh, yielding a particle size range from 0.25 to 0.60 mm. It has been observed the desert gets the surface temperature up to 50 to 60°C in 30 to 35°C of Tamb [26]. In the experimental set up the desert was fully sealed with a transparent float glass with an adhesive of ‘M-Seal’, trademark. To enhance the ηther of new designed SAH, 2 halogen lights of ‘Phillips’ trademark (300 W each- MFG Model #: 415711) were placed in the inlet and outlet ducts of SAH at the distance of 65 cm from the open channels and the halogens can be performed on auxiliary power. The key advantage of the system was that it can be performed well in poor ambient conditions. All the experiments have been conducted at the M.I.T, Moradabad-244001, (latitude-28°58’N and Longitude-78°47’E). There were 4

Figure 1a. Desert spread on the absorber plate and sealed by float glass

42


Figure 1b. Desert spread on the absorber plate

different configurations for which the H1 and H2 have experimentally tested on both natural and forced convection. The variation in temperature of both the SAHS and Tamb was measured by using a 6 wire K-type thermocouple meter with an accuracy of ± 1°C. The solar insolation (W/m2) on the horizontal surface was directly measured by a standard solarimeter ‘Surya-maapi’ (CEL-201) with an accuracy of 1 W/m2. The velocity of the air was monitored using an anemometer with 1% accuracy located nearby the experimental set up while the air temperature was measured using mercury in-glass type thermometer. The measured variables were recorded at time intervals of 15 min (and discussed on an hourly basis with an average value of 04 actual reading values) included; solar radiation, inlet and outlet temperatures of the air circulating through the systems, ambient temperature, absorber plate temperatures at three selected points, and wind speed. The experiments were repeated for different mass air flow rates (0.019 kg/s for natural convection and 0.12 kg/s on forced convection). All the experiments have been started at 10:00 AM and ended at 19:00 PM to observe the behavior of heat storage media. RESULT AND DISCUSSION The thermal efficiency of a solar collector is the major requirement to predict the thermal performance of the complete solar system of which the solar air collector is a part. It is, therefore, important that the ηtherm information is in a form which is useful for existing and projected future solar energy system design methods. To evaluate the thermal

Figure 2. Schematic view of experimental set-up of solar air heater

Vol. 29, No. 3 2014 43

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performance of solar air heaters different methods are developed earlier [27]. Among them steady-state procedure is the most common, the , procedure involves simultaneous and accurate measurements of m the inlet and outlet temperatures of the collector fluid, and the ambient conditions. Here, after a change of flow rate, both the solar heaters ‘H1’ and ‘H2’, was allowed to attain a steady state before experimentation. Following are the parameters, used to evaluate the thermal performance of both solar heaters ‘H1’ and ‘H2’ and enhancement by using TES. Mass flow rate

[1]

Equivalent hydraulic diameter de = 2LH/L+H

[2]

Reynolds number

[3]

Friction factor

[4]

Pressure drop ΔP = 2f (ρ.V2) L/de Overall heat transfer coefficient

[5]

Overall heat loss coefficient

[6] [7]

Equation 7 can be solved by using following relations;

hw = 5.7 + 3.8vw

Ct = 520.[1 – 0.0000513

[8] •

b2]

ft = (1 + 0.089 hw + 0.1166 hwep)(1 + 0.0786N) •

•

[9] [10]

Vol. 29, No. 3

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Ub = ki/ti

[11]

UL = Ut + (ki/t)

[12]

[13]

The value obtained for hw is 9.5 W/m2k, for Ct is 473.25 W/m2k, for ft is 3.37, for Ub is 1.75, for Ut is 10.09, for UL is 11.84 W/m2k, N = 1 because of single glazing, σ = 5.67 x 10-8 W/m2k4 and the emissivity of the absorbing plate and glass cover are e p = 0.91 , and e = 0.81 respecg tively. Pressure drop was observed 0.271 (at velocity of air = 0.30 m/s and f = 0.01) and 0.163 (at velocity of air = 0.95 m/s and f = 0.006) for natural and forced convection respectively for full length 3.19 m, de = 0.254, ρ = 1.1 kg/m3.

Figure 3. Heat transfer mechanism of SAH

To evaluate the thermal performance of both the models ‘H1’ and ‘H2’ at different 4 configurations on natural and forced convection, all the experiments have been conducted in the month of April 2012 (10.04.2012 to 17.04.2012) from 10:00AM to 19:00 PM because of a drawn out sunshine hours. Both the systems were placed at southward direction at an angle of 43o from horizontal, exposed to the Sun and both the systems were operated on natural convection by supplying of fresh air

46


to the inlet duct. It is notable that the cylindrical-conical duct was not used in any of the natural convection process. While in the forced convection it is considered as an important element to supply the forced air to the inlet rectangular duct of the SAH by an additional circular duct at constant velocity. The velocity speed at the inlet rectangular duct from circular duct was measured 2 m/s and kept constant for all forced convection experiments. The experiments were conducted in a subsequent manner from point 1 to 8, for both the models simultaneously to observe the variations in performance (For heat transfer mechanism see figure 3)

i). On 10.04.2012 both the models ‘H1’ and ‘H2’ were operated on natural convection.

ii). On 11.04.2012 both the models ‘H1’ and ‘H2’ were operated on forced convection.

iii). On 12.04.2012 both the models ‘H1’ and ‘H2’ were operated on natural convection by placing a 300 W halogen light inside the inlet duct at 65 cm from opening end.

iv). On 13.04.2012 both the models were ‘H1’ and ‘H2’ operated on forced convection by placing a 300 W halogen light inside the inlet duct at 65 cm from opening end.

v). On 14.04.2012 both the models were ‘H1’ and ‘H2’ operated on natural convection by placing a 300 W halogen light inside the outlet duct at 65 cm from opening end.

vi). On 15.04.2012 both the models were ‘H1’ and ‘H2’ operated on forced convection by placing a 300 W halogen light inside the outlet duct at 65 cm from opening end.

vii). On 16.04.2012 both the models were ‘H1’ and ‘H2’ operated on natural convection by placing two separate 300 W halogen lights inside the inlet and outlet duct at 65 cm from opening ends. viii). On 17.04.2012 both the models were ‘H1’ and ‘H2’ operated on forced convection by placing two separate 300 W halogen lights inside the inlet and outlet duct at 65 cm from opening ends. On the first day (10.04.2012) both the solar heaters ‘H1’ and ‘H2’ were operated on natural convection simultaneously. At the starting (10:00 Hrs) of experiments the Tamb was observed 29°C with 580 W/m2

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of solar radiation and ended with 29°C with 490 W/m2 at 19:00 hrs. The Tp was observed maximum 59°C and 62°C at 15:00 Hrs at noon respectively for ‘H1’ and ‘H2’ during the performance testing. The maximum ηtherm was found 14.1% and 15.98% at 15:00 Hrs and 17:00 Hrs, while the minimum ηtherm was measured 8.38% and 10.18% at 10:00 Hrs respectively for ‘H1’ and ‘H2’ (Figure 4). The average output hot air velocities were measured 0.31 m/s and 0.33 m/s with average output temperatures 44.50°C and 47.30oC, respectively for ‘H1’ and ‘H2’. On the next day (11.04.12) both the SAHS were operated on forced convection at a constant air velocity (2 m/s), and in this testing both the systems were found to be taken much time to get hot in comparison of natural convection. At the starting (10:00 Hrs) of experiments the Tamb was observed 30°C with 590 W/m2 of solar radiation and ended with 29oC with 485 W/m2 at 19:00 hrs. The Tp was noticed maximum 47oC and 49°C at 14:55 Hrs and 14:05 Hrs respectively for ‘H1’ and ‘H2’ at noon while testing. The maximum ηtherm was found 41.44% and 47.97% at 14:50 Hrs and 13:00 Hrs, while the minimum ηtherm was observed 33.45% and 29.73% at 10:00 Hrs respectively for ‘H1’ and ‘H2’ (Figure 5). The average output hot air velocities were 0.93 m/s and 0.95 m/s while the average output temperatures were measured 35°C and 36.7°C respectively for ‘H1’ and ‘H2’. On 12.04.12, the experiments were conducted on natural convection by placing a halogen light of 300 W at the inlet ducts of both the solar heaters at the distance of 65 cm from the ends open for intake of fresh air of ducts. After the performance evaluation on this configuration of both the systems, the results were found satisfactorily with an improve-

Figure 4. Performance of H1 and H2 on plain absorber plates on natural convection

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Figure 5. Performance of H1 and H2 on plain absorber plates forced convection

Figure 6. Performance of H1 and H2 by placing halogen light inside the inlet duct on natural convection

Figure 7. Performance of H1 and H2 by placing halogen light inside the inlet duct on forced convection

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ment in output air velocity, outlet temperature and ηtherm. Readings were taken in the same way with an interval of 15 minutes during 10:00 hrs to 19:00 hrs. Tp was noticed maximum 62°C and 66°C at Hrs and 14:55 Hrs at noon for ‘H1’ and ‘H2’ respectively during the performance evaluation. The maximum ηtherm was observed 15.19% and 17.36% at 15:00 Hrs respectively while the minimum ηtherm was evaluated 8.98% and 9.58% at 9:55 Hrs respectively for ‘H1’ and ‘H2’. The average output hot air velocities respectively for ‘H1’ and ‘H2’ were observed 0.32 m/s and 0.35 m/s with an average output temperatures 46°C and 48.5°C while testing (Figure 6). On 13.04.12 both the SAHS were operated on forced convection (as previous). This time Tp was noticed maximum 47°C and 50°C at 14:55 Hrs and 14:05 Hrs at noon for ‘H1’ and ‘H2’ respectively. The maximum ηtherm was found 46.73% and 54.39% at 15:50 Hrs and 14:10 Hrs respectively while the minimum ηtherm was found 33.73% and 29.99% at 10:00 Hrs respectively for ‘H1’ and ‘H2’ (Figure 7) during the experiments. The average output hot air velocities were 0.93 m/s and 0.94 m/s for both ‘H1’ and ‘H2’ while the average output temperatures were measured 36.9°C and 37.7°C respectively for ‘H1’ and ‘H2’. On the next day 14.04.2012, the experiments were conducted by placing halogen lights of 300 W inside outlet ducts of both the solar heaters at the distance of 65 cm from the opening ends of ducts (open to the environment for heating). In this configuration a good improvement was observed in output air’s temperature while ηtherm was found for an agreed improvement for both the air heaters. The maximum value of Tp was noticed 66°C and 68°C at 13:55 Hrs at noon for ‘H1’ and ‘H2’ respectively. The maximum ηtherm was observed 16.56% and 17.63% at

Figure 8. Performance of H1 and H2 by placing halogen light inside the outlet duct on natural convection

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Figure 9. Performance of H1 and H2 by placing halogen light inside the outlet duct on forced convection

14:00 Hrs, while the minimum ηtherm was observed 10.18% and 10.77% at 10:00 Hrs respectively for ‘H1’ and ‘H2’ during the testing. The average output hot air velocities were 0.32 m/s and 0.34 m/s and the average output temperatures were measured 48.1°C and 50°C respectively for ‘H1’ and ‘H2’ (Figure 8). On 15.04.12 both the SAHS were operated on forced convection. In this case Tp was notified maximum 49°C and 51°C at 13:05 Hrs and 13.50 Hrs at noon for ‘H1’ and ‘H2’ respectively. The maximum ηtherm was noticed 51.39% and 54.82% at 14:00 Hrs respectively for ‘H1’ and ‘H2’ and the minimum ηtherm was found 36.85% and 33.17% at 10:00 Hrs respectively for ‘H1’ and ‘H2’ (Figure 9) while conducting the experiments. The average output hot air velocities were 0.93 m/s and 0.95 m/s for ‘H1’ and ‘H2’ respectively while the average output temperatures were measured 38°C and 39.10°C respectively for ‘H1’ and ‘H2’. On 16.04.2012, the performance evaluation of both solar heaters was carried out by placing halogen lights of 300 W inside both the ducts (inlet and outlet) of both SAHS at same position as in previous configurations. This configuration was found to provide a better output in terms of air’s temperature and ηtherm of both the heaters on both natural and forced convection heat transfer with respect of all previous configurations. Tp was noticed maximum 65°C and 67°C at 14:55 Hrs at noon for ‘H1’ and ‘H2’ respectively. During the testing, the maximum ηtherm was observed 17.50% and 18.74% at 14:50 Hrs and 14: 05, while the minimum ηtherm was evaluated 11.87% and 12.46% at 10:00 Hrs respectively for ‘H1’ and ‘H2’ (Figure 10). The average output hot air velocities were 0.32

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Figure 10. Performance of H1 and H2 by placing halogen lights inside the inlet and the outlet ducts on natural convection

Figure 11. Performance of H1 and H2 by placing halogen lights inside the inlet and the outlet ducts on forced convection

m/s and 0.34 m/s with an average output temperatures were measured 49.3°C and 51.3°C respectively for ‘H1’ and ‘H2’. On the next day while operating both the air heaters Tp was notified maximum 51°C and 54°C at 13.55 Hrs at noon. The maximum ηtherm was found 60.12% and 69.06 % at 13:00 Hrs and at 13:55 Hrs while the minimum ηtherm was found 37.31% and 41.59% at 10:10 Hrs respectively for ‘H1’ and ‘H2’ (Figure 11). The average output hot air velocities were 0.93 m/s and 0.95 m/s for ‘H1’ and ‘H2’ while the average output temperatures were measured 39.9°C and 41.95°C respectively for ‘H1’ and ‘H2’. After the completion of all the experiment testing for the thermal performance evaluation of both the solar heaters successfully, the heat transfer rate was observed moderately improved while operating both

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the systems on forced convection in comparison of operating on natural convection. The Qu ranges from 242.51 W to 372.5 W for H1 model while 295 W to 372.35 W for model H2 respectively in the case of natural convection. In the case of forced convection Qu ranges from 373.49 W to 1036.13 W for H1 model while 578.30 W to 1373.47 W for model H2 respectively. Apart this, the heat transfer coefficient was found to be increased for each new configuration (from configuration 1 to 4) on both operating conditions (i.e., natural and forced convection) for both the models. The value of “h” was observed in a range from 42.69 W/m2k to 57.29 W/m2k and 45.67 W/m2k to 72.73 W/m2k for ‘H1’ and ‘H2’ respectively on natural convection while 64.84 W/m2k to 392.47 W/m2k and 86.06 W/m2k to 429.21 W/m2k for ‘H1’ and ‘H2’ respectively on forced convection respectively. In both the solar heaters model “H2” was found better than model “H1” in temperature of the exhaust air, while the output air velocities range from 0.31 m/s to 0.34 m/s for natural convection and 0.91 m/s to 0.93 m/s for both the solar heaters. The minimum and maximum value of Tos1 was 26.7°C and 33.2°C while 30.1°C and 37.4°C for Tos2. This temperature “37.4oC” of model H2 is good enough for heating or drying in poor ambient conditions and makes the models H2 better in comparison of model H1. It is also notable that all the modifications made to both air heaters in a similar manner but model H2 is found better for thermal performance than H2. In order to perform the uncertainty analysis of experiments conducted from 10.04.2012 to 17.04.2012, all the experiments have been repeated in the same sequence from 20.04.2012 to 27.04.2012. The uncertainty of each parameter is shown in table 2. It can be concluded that the uncertainty at 96% confidence level is ± 0.98.5% for thermal efficiency. Table 2. Uncertainty analysis of performance parameters

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CONCLUSION A SAH “H2” has been designed, fabricated and evaluated for thermal performance on 4 different configurations. The results were compared with a similarly designed SAH (H1) consisting an Al made dull blackened absorber tray. All the experiments were carried out on both ‘H1’ and ‘H2’ in the same ambient conditions simultaneously. It was observed that the “H2” got the maximum temperature in very short span of time in comparison of “H1.” A good improvement was found in the thermal performance of both the heaters while performing on all new configurations on both natural and forced convection. The results show that there was a significant effect on “H1” and “H2” with ambient conditions. The maximum ηtherm 14.1%, 15.19%, 16.56%, and 17.5% were observed for “H1” on natural convection for four different configurations while for “H2,” the maximum ηtherm were obtained as; 15.73%, 17.36%, 17.63%, and 18.74% on same configurations on natural convection. There was a good improvement by operating both heaters on forced convection and the maximum ηtherm were obtained as; 41.44%, 46.73%, 51.39%, and 60.12% for “H1” while the maximum ηtherm for “H2” were calculated as; 47.97%, 54.39%, 54.82%, and 69.06%. Beside this a good stability in Tp, output air velocity, outlet temperature, and ηtherm were found in “H2” in comparison of “H1” while performing. The model “H2” was also found as well high heat storage SAH and suitable for drying and space heating than conventional SAHS. Model “H2” was also found economical by eliminating a blower of high power consumption [28], which minimizes the system’s operating and maintaining cost. The use of halogen lights makes systems feasible to perform in bad climatic conditions. This model “H2” has the feasibility to increase the exhaust temperature by increasing the number of halogen lights or accordingly (in the case of requiring high heat temperature of the air) in poor ambient conditions. References [1] K.W. Kauffman,

Some choices of materials for thermal energy storage, Solar Engineering, 1976, 19-20. [2] P.C. Auh, Solar energy: its technologies and applications, KSEA Symposia’78: Seoul, Korea. July, 1978. [3] Charles Lee, Lawrence Taylor, John De Vries, and Stephen Heibein, Solar application of thermal energy storage, final report, H-C0199-79-753F, U.S. Department of Energy, Hitman associates Inc., Columbia, Maryland- 21045, 1979. [4] J.A. Duffie, W.A. Beckman, Solar Engineering of Thermal Processes, Wiley & Sons,

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Distributed Generation and Alternative Energy Journal New York, 1991. [5] Bruce Anderson and Michael Riordan, The New Solar Home Book, by R.A.K. Publishing Co., U.S.A. 1987. [6] C.L. Gupta and H.P. Garg, Performance studies on solar air heaters, Solar Energy, 1967, 11, 25-31. [7] S.A. Kalogirou, Solar thermal collectors and applications, Progress in Energy and Combustion Science, 2004, 30, 231-295. [8] G.N. Tiwari, Solar Energy, Narosa Publishing house, New Delhi, 2002. [9] W.A. Stanley and E.W. Charles, Energy storage, Economics of Solar Energy and Conservation Systems, 1979, 207-250. [10] I. Dincer and M. A. Rosen, Thermal Energy Storage: Systems and Applications, Published by John Wiley & Sons Ltd, United Kingdom, 2011. [11] I. Abbud, G.O.G. Lof, and D.C. Hittle, Simulation of solar air heating at constant temperature, Proceedings of SOLAR 1993, The American Solar Energy Society Annual Conference, Washington DC, (April-1993) 22-28. [12] P.M. Chauhan, C. Choudhary and H.P. Garg, Comparative performance of coriander dryer coupled to solar air heater and solar air heater cum rock-bed storage, Applied Thermal Engineering, 1996, 16, 475-486. [13] S. Aboul-Enein, A.A. El-Sebaii, M.R.I. Ramadan, H.G. El-Gohary, Parametric study of a solar air heater with and without thermal storage for solar drying applications, Renewable Energy, 2000, 21, 505-522. [14] S.O. Enibe, Performance of a natural circulation solar air heating system with phase change material energy storage, Renewable Energy, 2002, 27, 69-86. [15] N.S. Thakur, J.S. Saini, and S.C. Solanki, Heat transfer and friction factor correlations for packed solar air heater for a low porosity system, Solar Energy, 2003, 74, 319-329. [16] P. Naphon, Effect of porous media on the performance of the double-pass flat plate solar air heater, International Communications in Heat and Mass Transfer, 2005, 32, 140-150. [17] S.B. Prasad, J.S. Saini, K.M. Singh, Investigation of heat transfer and friction characteristics of packed bed solar air heater using wire mesh as packing material, Solar Energy, 2009, 83, 773-783. [18] P.T. Saravanakumar and K. Mayilsamy, Forced convection flat plate solar air heaters with and without thermal storage, Journal of Scientific & Industrial Research, 2010, 69, 966-968. [19] M.M. Alkilani, K. Sopian and S. Mat, Fabrication and experimental investigation of PCM capsules integrated in solar air heater, American Journal of Environmental Sciences, 2011, 7, 542-546. [20] S. Karthikeyan and R. Velraj, Numerical and experimental investigation of the charging and discharging processes in a packed bed PCM based storage unit for air heating applications, European Journal of Scientific Research, 2011, 66, 105-119. [21] V.V. Tyagi, A.K. Pandey, S.C. Kaushik, S.K. Tyagi, Thermal performance evaluation of a solar air heater with and without thermal energy storage, J Therm Anal Calorim, 2011, 1-8. [22] Walid Aissa, Mostafa El-Sallak, and Ahmed Elhakem, An experimental investigation of forced convection flat plate solar air heater with thermal storage material, Thermal Science, 2012, 16, 1105-1116. [23] S. S. Krishnananth and K. K. Murugavel, Experimental study on double pass solar air heater with thermal energy storage, Journal of King Saud University—Engineering Sciences, 2012, http://dx.doi.org/10.1016/j.jksues.2012.05.004. [24] H.K. Elminir, A.E. Ghitas, R.H. Hamid, F.E. Hussainy, M.M. Beheary, K.M. Abdel-

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[25]

[26]

[27]

[28]

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Moneim, Eﬀect of dust on the transparent cover of solar collectors, Energy Conversion and Management, 2006, 47, 3192–3203. T.T. Warner, Desert meteorology, Published by Cambridge University Press, New York, 2005. C.D. Kern and Capt., USAF, Desert soil temperature and infrared radiation received by TIROS III, Journal of The Atmospheric Sciences, 1963, 175-176. Ram Chandra and M.S. Sodha, Testing procedures for solar air heaters: A review, Energy Conversion Management, 1991, 32, 11-33. Mahmud M. Alkilani, K. Sopian, M.A. Alghoul, M. Sohif, M.H. Ruslan, Review of solar air collectors with thermal storage units, Renewable and Sustainable Energy Reviews, 2011, 15, 1476-1490.

———————————————————————————————— ABOUT THE AUTHORS Abhishek Saxena is an Assistant Professor in the Department of Mechanical Engineering in Moradabad Institute of Technology, Moradabad-244001, India. He is also a member of five international societies of renewable energy and serving various international research publication houses as an editorial member and reviewer. Currently, he is working on his area of interest, i.e., alternative energy fuels and solar hybrid energy systems. Dr. Vineet Truth is a professor in the department of Mechanical Engineering, at M.I.T, Moradabad. He is an active member of various reputed universities of India as an expertise of Material science, energy materials. His area of interest is materials science and energy systems. Dr. Ghanshyam Srivastava is the Director of Eshan College of Engineering, Mathura-281001, India. He is an active member of various reputed universities of India as an expertise of thermal engineering especially in gas power turbines. His area of interest is thermal engineering and renewable energy.

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Development of an Empirical Model For Assessment of Solar Air Heater Performance P.K. Choudhury and D.C. Baruah ABSTRACT Solar flat plate collectors are used for meeting the hot air requirement in a number of applications ranging from domestic to industrial sectors. However, the varying degree of uncertainties of available solar radiation along with varying weather conditions often acts as restriction to their wider use. The varying input conditions like solar radiation, ambient temperature and relative humidity have diverse effect on the output of the flat plat collector used for hot air generation and poses difficulty in terms of reliability in providing the output as per users’ requirement. This study refers to a method of predicting the output of flat plate collectors under varying working conditions. The model suggested here defines a relationship between the output in terms of instantaneous efficiency and the inputs, namely, solar radiation, ambient temperature, fluid inlet temperature and the mass flow rate. The relationship is based on the behavior of a given flat plate collector under certain sets of input parameters and can be used to estimate the efficiency of the collector at various mass flow rates. Further, this model can be suitably extended to estimate the mass flow rate for providing a particular output under specific set of input parameters. Key words: Flat plate collector, mass flow rate, efficiency, solar air heater INTRODUCTION In the context of present energy crisis, use of renewable energy sources plays an important role in energy conservation efforts. Solar energy, being a major source among all renewable energy sources, has enough potential for thermal applications. In practice, various types of

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devices are being utilized to trap and use solar thermal energy. However, it is difficult to operate such devices to provide constant output continuously due to varying operating conditions or region specific uncertainties associated with solar radiation. The most common examples of using solar thermal energy are the solar water heating systems and the solar air heating systems. Water heating systems utilize solar radiations to heat up water whereas the air heating systems use the solar radiation to heat up the air. Hot water is required in various purposes such as domestic, industrial, commercial establishments, educational institutes etc. These requirements can be met by using solar water heating systems [5,6,10]. Different types of commercial systems such as flat-plate type, evacuated tube type etc are available for such applications and experimental study comparing their performances are also conducted by researchers [21]. Similarly, solar air heaters are also used for many purposes. Uses of solar air heaters are reported for drying of fruits, vegetables and other agricultural and marine products [2,3,8,16,18]. It may also be suitably used for providing industrial process heat requirements, room heating and other similar applications. The performance of either a solar water heating system or an air heating system is governed by the behavior of the collector used to trap solar thermal energy. Among different types of collectors used in solar thermal applications, flat plat collector is the most common, which is featured by its constructional simplicity and economy. A flat plat solar air heater is a flat plate collector used to generate hot air. As shown in Figure 1, it comprises of an absorber plate, a transparent cover at the top and a parallel plat at the bottom which forms the air passage. The absorber plate is painted black with selective coating and the transparent cover is usually formed by one or more numbers of glass cover. The system is properly insulated to reduce the heat loss from the sides as well as from the bottom. Solar radiation falling on the absorber plate through the glass cover is utilized to heat up the air flowing in the air passage. The circulation of air is usually forced by a fan or a blower and a temperature gradient exists in the air stream flowing from inlet to the outlet. With the advancement of technology, research and development works related to use of solar thermal energy is primarily oriented towards improvement of efficiency. Various research works are reported in the areas of improving the performance of solar air heaters. This in-

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Figure 1. Schematic diagram of a typical solar air heater.

cludes using different types of collectors in terms of geometry, material and configuration [3,11,19]. Similarly works have also been reported in the field of efficiency improvement of solar flat plat collectors [1,4,7,16]. The effect of changing glass cover, single and double pass circulation, incorporation of barriers dividing the air channel on the collector performance, efficiency and exergy analysis are also reported [12,14,20]. Economics of some industrial applications of solar air heater as well as case studies of technical feasibilities are also reported by researchers [8,15]. There are still certain issues concerning the operation of solar air heater as per users’ suitability.

60


From users’ perspective, two distinct operating modes of flat plat solar collector may be visualized, viz., (i) hot air required at constant temperature and (ii) hot air required at maximum efficiency of the system. For example, supply of hot air at constant temperature is necessary for drying of crops and vegetables whereas operation of the collector at the best efficiency is required in case of hybrid system or integrated renewable energy systems. Ideally, for the first case, the operation may lead to the reduction of collector efficiency while the second case of operation may lead to the variation of the air temperature at the outlet far beyond the required or acceptable range. It is difficult to operate a solar flat plat collector naturally in either mode continuously due to the varying (temporal and spatial) nature of input solar radiation. The situation thus demands for a mechanism or a method which will predict the behavior of the flat plate collector under varying input conditions so that some measures can be appropriately taken for getting the intended output. Table 1. Major input parameters for solar flat plate collector

The temperature of air at the outlet as well as the efficiency of a collector depends on number of parameters. These parameters can be classified as (a) design parameters, (b) environmental parameters and (c) user controlled parameters. The performance of the flat plat collector varies with the changes in either one or all of these parameters (Table 1). As mentioned earlier, a flat plate solar collector in the forced circulation mode consists of a device (blower or fan) to facilitate the flow of work-

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ing fluid from inlet to the outlet. The rate of flow has an effect on the output temperature as well as on performance of the flat plate collector. This flow rate can be termed as user controlled parameter as it can be adjusted through blower. Solar radiation and ambient temperature, which are controlled by nature, are treated as environmental parameters. The environmental parameters cannot be manipulated as per users’ choice while the design parameters, such as, gross area of the collector and area of the absorber plate, remain fixed for a particular collector. Thus, for getting collector output as per users’ choice, possible scope lies in user controlled parameters and suitable ways should be explored for manipulating these parameters in order to compensate the variations of the nature controlled environmental parameters. A precise relationship among the input and output parameters reflecting their physical dependency will be very much helpful for decision making process in output specific applications. Keeping in view of the above, an attempt has been made to develop a mathematical relationship to predict the performance of a solar air heater. MODELING APPROACH In the present study, a relationship among the input and output parameters is explored with the background knowledge of collector performance for a wide range of operating conditions. The input parameters consist of design parameters, environmental parameters and user controlled parameters. Flat plate collectors having known design parameters such as gross collector area, area of the air flow duct, area and type of the absorber plate are considered for this investigation. The environmental parameters considered are the solar radiation, ambient temperature and fluid inlet temperature in open loop mode. Similarly, user controlled parameter, i.e., the mass flow rate of air is considered. The output parameter is the instantaneous efficiency (ηi) which depends on both environmental and user controlled parameters. A collective set of environmental parameters (Tfi, Ta and IT), user controlled parameter (ṁ) and a fixed set of design parameters (such as Ap, Ac, τ, α, ep, ec, M, L) decides the behavior of the collector. For a fixed set of design parameters, the instantaneous efficiencies are experimentally observed under given operating conditions (environmental and user controlled parameters). A model is formulated with the available theory on heat

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transfer along with available experimental data to predict performance of solar air heater. The theoretical relationship is compared with the trend of experimental investigation to identify important parameters in terms of slope and intercept of corresponding plots. Further, these slopes and intercepts are studied independently at varying flow rates and fitted into a power curve to obtain two characteristic power equations, namely, characteristic slope equation and characteristic intercept equation. Characteristic slope equations and characteristic intercept equations are collector design specific and are useful for predicting the performance. DEVELOPMENT OF THE MODEL As discussed earlier, the model under the present study deals with the development of a relationship between the instantaneous efficiency, ηi, and the temperature parameter [(Tfi –Ta)/IT] with respect to mass flow rate, ṁ. Experimental data for a collector will be prerequisite for developing the relationship. The relationship then may be used to estimate the behavior of the collector at various required mass flow rates. The procedure is briefly highlighted below. The instantaneous efficiency of a collector can be expressed following the Hottel-Whillier-Bliss procedure [17,20] as,

(1)

The collector heat removal factor FR refers to the thermal resistance faced by solar radiation in reaching the collector fluid and is given by the ratio of the actual heat gain rate to the gain if the fluid temperature through the absorber plate were at Tfi everywhere. Typical values of FR fall in the range of 0 to 1 [17]. FR can be expressed as, where F́ is the collector efficiency factor.

(2)

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Eq. (2) implies that FR is a function of mass flow rate, specific heat, overall loss coefficient, absorber plate area and collector efficiency factor. The variation of FR with respect to these parameters can be investigated. It is observed from the Eq. (2) that for a constant mass flow rate, FR is dependent on specific heat, overall loss coefficient, and absorber plate area and collector efficiency factor. For a particular collector, its geometry remains the same, i.e., it will have constant absorber plate area as well as gross collector area. Again, the specific heat of air varies with temperature. For the temperature range from 10 to 93.30 C, cp varies from 1.0048 to 1.0090 kJ/kg-K and may be considered as constant at an average value of 1.0069 kJ/kg-K [20]. The collector efficiency factor F́ can be expressed in terms of effective heat transfer coefficient (he) between the absorber plate and air stream and overall heat loss coefficient (Ul) as

(3)

where and

(4) (5)

Ub, Us and Ut in Eq. (5) represent the bottom loss, side loss and top loss coefficients respectively. The bottom and side loss coefficients can be expressed as

(6)

(7)

where 𝓀i is the thermal conductivity of insulation (W/m-K), δb thickness of the bottom insulation (m), L1 and L2 length (m) and width (m) of the absorber plate, L3 height of the collector casing (m), δs thickness of the side insulation (m). The top loss coefficient Ut can be expressed by the correlation suggested by Malhotra et al [13,17] as

64


(8)

where, M L h w Tpm Ta σ β e p eco

(9)

(10) = number of glass covers = spacing between the covers or absorber plate and cover (m) = wind heat transfer coefficient (W/m2-K) = mean absorber surface temperature (K) = ambient temperature (K) = Stefan – Boltzmann constant (W/m2-K4) = slope or tilt (degrees) = emissivity of absorber surface for long wave length radiation = emissivity of cover for long wave length radiation

From Eq. (8) it can be seen that Ut depends primarily on the mean absorber surface temperature, ambient temperature, wind heat transfer coefficient and flow conditions. The convective heat transfer coefficients (hfp and hfb) and the radiative heat transfer coefficient hr are also dependent on the flow condition as well as on the temperature of the absorber and bottom plate. Considering these variations during the observation period as negligible FR, (τα)av, Ul and F́ can be treated as constant for the period of observation. Modeling Instantaneous Efficiencies of a Given Collector for Specified Mass Flow Rate Under the above assumptions, for a particular mass flow rate ṁi, the Eq. (1) can be written as below.

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(11)

where Ai represents the negative slope and Bi represents the Y –axis intercept for ith mass flow rate. Expressions for Xi, Ai and Bi are provided below.

(12)

(13)

(14)

Here the subscript i is used to indicate varying mass flow rate. Modeling Instantaneous Efficiencies of a Given Collector for Varying Mass Flow Rates By performing experiments at different mass flow rates different sets of values of Xi and ηi are obtained. Plotting the values of ηi against Xi, scatter plots are obtained for particular mass flow rate ṁi which can be fitted into a straight line represented by the Eq. (11). From this equation, the values of Ai and Bi corresponding to the ith mass flow rate can be determined. Experimental results in various literatures have shown the variations in the slope and intercepts according to the mass flow rates [9, 18]. It is observed that by increasing mass flow rate the slope as well as the intercept increases. These variations against mass flow rates is best fitted into a power curve defined by the following characteristic equations

(15)

(16)

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where DA, EA, DB and EB are constants. Eq. (15) is termed as characteristic slope equation and Eq. (16) is termed as characteristic intercept equation. The constants DA, EA, DB and EB are characteristic model parameters for the collector. Once the characteristic equations (Eq. (15) and Eq. (16)) are defined and characteristic model parameters (DA, EA, DB and EB) are known, A and B for any mass flow rate ṁ can be determined and the corresponding efficiency may be represented by the following equation

(17)

DATA SOURCE FOR MODEL TESTING In order to test the model, four flat plate solar thermal collectors (STC1, STC2, STC3 and STC4) were considered. The necessary data (ηi, Xi and respective ṁ) have been taken from published literature [9,18]. The specification of the collectors taken from literature is provided in Table 2. Table 2. Different collectors used for testing the model

The flow rates considered for finding out the characteristic constant for the model (Eqs. (16) to (21)) are given in the Table 3. RESULTS AND DISCUSSION The values of A and B are determined from the efficiency curve for each flow rate (Table 3) and for each collector (Table 2). These values are

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7 - Auditing, Planning and Retrofitting 8 - Work Order Development by the Auditor 9 - Heating and Cooling 10 - Baseload and How to Improve It 11 - New Construction Energy Evaluations 12 - Building Professional Training and Certification Appendices, Glossary, Index

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PUMP USER’S HANDBOOK: LIFE EXTENSION, FOURTH EDITION


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Heinz P. Bloch and Allan R. Budris

Just published in its updated fourth edition, this highly regarded text explains in clear terms how and why the best-of-class pump users are consistently achieving superior run lengths, low maintenance expenditures, and unexcelled safety and reliability. Written by practicing engineers whose working careers were marked by involvement in all facets of pumping technology, operation, assessment, upgrading and cost management, this book endeavors to describe in detail how you, too, can accomplish optimum pump performance and low life cycle cost. A new chapter on breaking the cycle of pump repairs examines the cost of failures and the defined operating range of pumps. The authors also explore mechanical issues, deviations from best available technology, and preventing problems with oil rings and constant level lubricators. Additional topics include bearing housing protector seals, best lube application practices, lubrication and bearing distress, and paying for value.


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1 Pump System Life Cycle Cost Reduction 2 How to Buy & Ship a Better Pump 3 Piping, Baseplate, Installation, & Foundation Issues 4 Operating Efficiency Improvement Considerations 5 Improved Pump Hydraulic Selection Extends Pump Life 6 Improvements Leading to Pump Mechanical Maintenance Cost Reduction 7 Bearings in Centrifugal Pumps 8 Mechanical Seal Selection & Application

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9 Improved Lubrication & Lubricant Application 10 Oil Mist Lubrication & Storage Protection 11 Coupling Selection Guidelines 12 Pump Condition Monitoring Guidelines 13 Pump Types & Materials 14 Pump Failure Analysis & Troubleshooting 15 Shop Repair, Spare Parts Availability & Procurement 16 Failure Statistics & Component Uptime Improvement Summary 17 Breaking the Cycle of Pump Repairs Appendices, References, Index

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69

Table 3. Flow rates used for estimation of characteristic constants in the model

fitted against ṁ into power curves using spreadsheet (Microsoft EXEL). The corresponding regression equations in the form of Eq. (15) and (21) are shown in the Table 4 and Table 5. Table 4. Characteristic slope equations derived from experimental data of four different collectors

Table 5. Characteristic intercept equations derived from experimental data of four different collectors

The equations as obtained in Table 4 and Table 5 are used to predict the efficiency at some selected flow rates (0.0317, 0.0248, 0.0438 and 0.055 kg/m2-s for STC1, STC2, STC3 and STC4 respectively). The efficiency equations for respective collectors are shown in Table 6. Using the above equations, efficiencies for some representative Xvalues are estimated and compared with actual experimental values for each collector. The comparison can be seen in Figures 2-5.

70


Table 6. Efficiency equations at selected flow rates for four different collectors

Figure 2. Observed and estimated efficiencies of STC1 at 0.0317 kg/m2-s flow rate.

It can be seen from the plots that the model predictions are closer to the experimental values almost for all the four collectors. However, in case of STC4 the deviation of the predictive efficiency from the observed value increases for higher values of X parameter. This may be due the fact that only three sets of data were considered for determining DA, DB, EA and EB in STC4. Also it is observed that the experimental data for X parameter which were used to determine the efficiency curve at the flow rate of 0.070 kg/m2-s, falls mostly in the range of 0.015 to 0.020 C/W/m2

Vol. 29, No. 3

2014

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with a span of about 0.050 C/W/m2. So it may be expected that for better estimation by this model, observations with at least 4 to 5 mass flow rates may be done and the X parameter should be distributed uniformly over a wide range. The effect of relative humidity on collector performance is not considered in this model. MODEL VERIFICATION AND VALIDATION As discussed in Section 2, the model is developed using certain sets of experimental observations and takes into account of four different types of collectors. In general, it is seen that the model predictions reasonably agrees with the observed values. The marginal differences between observed and predicted values might be due to (i) errors in assuming some parameters (cp, Ut, hfp, hfb and hr) as temperature independent and (ii) errors in neglecting the effects of wind velocity and

72



relative humidity of air. The required of model prediction is also verified through standard statistical technique (SPSS 16.0) which is incorporated in the Figures 2-5. CONCLUSIONS The domestic and industrial applications of flat plate solar air heater have remained limited mainly due to the difficulty associated with getting constant thermal output. The variations of either efficiency or output temperature are caused by related uncertainties. The uncertainties of output will have adverse effect on the quantity or quality of the product in typical applications, viz., drying applications. The model in this study is expected to predict the behavior of the collector in terms of efficiency with respect to varying conditions and mass flow rates. The

Vol. 29, No. 3

2014

73

Figure 5. Observed and estimated efficiencies STC4 at 0.055 kg/m2-s flow rate.

collector heat removal factor and the overall loss coefficient are the characteristics of a collector working under a given condition. The present study correlates these two parameters with measurable characteristic parameters such as slope and intercept of the performance equation. Now, the behavior of the device could be examined using the derived model parameters (DA, DB, EA and EB). Further, this model would also be useful to estimate the required mass flow rate corresponding to the desired efficiency or output temperature. This will enable to take appropriate decision for a given sets of specific temperature or performance efficiency as per user’s choice. Thus, this model would serve as a decision making tool for an automatic control mechanism for providing desired output. In addition to the capability of predicting the performance of a given collector, this empirical model is also useful to compare different flat plat collectors. The ever increasing demand for energy has become a major con-

74


cern in India. Continued efforts are being made to exploit renewable energy sources for various applications to reduce the gap between energy demand and supply. Assam, a north eastern state of India, has many tea industries requiring huge amount of thermal energy for different processes. These process heat requirements are generally fulfilled by using conventional energy sources and a suitable solar hot air generator can substantially reduce the consumption of conventional energy. Experimental setups are being prepared to investigate the validity of the model under local operating conditions of Assam in order to explore the possibility of using solar hot air generator in output specific applications. References 1. Amer, B.M.A., Hossain, M.A. and Gottschalk, K., 2010, Design and performance

2.

3.

4.

5.

6.

7.

8.

9.

10.

11.

12.

13.

14.

15.

16.

evaluation of a new hybrid solar dryer for banana, Energy Conversion and Management, 51, 813–820. Bala, B.K. and Mondol, M.R.A., 2001, Experimental investigation on solar drying of fish using solar tunnel dryer, Drying Technology, 19, 427–436. Fudholi, A., Sopian, K., Ruslan, M.H., Alghoul, M.A. and Sulaiman, M.Y., 2010, Review of solar dryers for agricultural and marine products, Renewable and Sustainable Energy Reviews, 14, 1–30. Fuller, R.J. and Charters, W.W.S., 1997, Performance of a solar tunnel dryer with microcomputer control, Solar Energy, 59, 151-154. Healey, H.M., 1997, Cost-effective solar applications for commercial and industrial facilities, Energy Engineering, 94, 34-49. Henkel, E.T., 2005, New solar thermal energy applications for commercial, industrial, and government facilities, Energy Engineering, 102, 39-58. Hossain, M.A. and Bala, B.K., 2007, Drying of hot chili using solar tunnel drier, Solar Energy, 81, 85–92. Jairaj, K.S., Singh, S.P. and Srikant, K., 2009, A review of solar dryers developed for grape drying, Solar Energy, 83, 1698–1712. Karim, M.A. and Hawlader, M.N.A., 2004, Development of solar air collectors for drying applications, Energy Conversion and Management, 45, 329–344. Kulatunga, A., 1999, Flat plate solar collector for a tropical climate: determining the solar energy contribution, Energy Engineering, 96, 35 – 45. Kumar, P., Shanmugam, S. and Veerappan, A.R., 2011, An experimental study on drying of non-parboiled paddy grains using an oscillating bed solar dryer, Energy Engineering, 108, 69-80. Kurtbas, I. and Durmus, A., 2004, Efficiency and exergy analysis of a new solar air heater, Renewable Energy, 29, 1489–1501. Malhotra, A., Garg, H.P. and Palit, A., 1981, Heat loss calculation of flatplate solar collectors, J. Thermal Energy, 2, 2. Ong, K.S., 1995, Thermal performance of solar air heaters: mathematical model and solution procedure, Solar Energy, 55, 93-109. Palaniappan, C. and Subramanian, S.V., 1998, Economics of solar air preheating in south Indian tea factories: a case study, Solar Energy, 63, 31–37. Sacilik, K., 2007, Effect of drying methods on thin-layer drying characteristics of hull-less seed pumpkin (cucurbita pepo l.), Journal of Food Engineering, 79, 23–30.

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17. Sukhatme, S.P., Solar energy principles of thermal collection and storage, Tata McGraw Hill, Second edition. 18. Tiris, C., Tirist, M. and Dincer, I., 1995, Investigation of the thermal efficiencies of a solar dryer, Energy Convers. Mgmt, 36, 205-212. 19. Tiwari, A., Sodha, M.S., Chandra, A. and Joshi, J.C., 2006, Performance evaluation of photovoltaic thermal solar air collector for composite climate of India, Solar Energy Materials & Solar Cells, 90, 175–189. 20. Yeht, H. and Lin, T., 1996, Efficiency improvement of flat-plate solar air heaters, Energy, 21, 435-443. 21. Yohanis, Y.G., Popel, O.S., Frid, S.E. and Kolomiets, Y.G., 2012, Detailed comparison of the performance of flat-plate and vacuum tube solar collectors for domestic hot water heating, International Journal of Sustainable Energy, 31, 347-364.

—————————————————————————————— ABOUT THE AUTHORS P.K. Choudhury, corresponding author, is a research scholar in the Department of Energy, Tezpur University, Tezpur-784028, Assam, India. He can be reached at e-mail: [email protected]. D.C. Baruah is a professor in the Department of Energy, Tezpur University, Tezpur-784028, Assam, India. He can be reached at email: [email protected].

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Guidelines for Graphics GENERAL The Distributed Cogeneration and Distributed Generation Journal is published in black and white only. No color graphics will appear. Therefore, it is important to have all graphics in black and white only. The EASIEST way to provide graphics is to simply submit hard copies of them. These copies should be of the highest possible quality. Print your graphics using a high quality printer. Almost any printer today is 300 dpi minimum, so this will be fine. What your printout looks like is what it will look like at best in the Journal. So if the image is sharp, clear and readable when you print it, then it should be fine for the Journal. But if your printout is fuzzy, blurry, and you can’t read some of the text or data, then the people who read the Journal will not be able to read it either. For tips on improving the quality of line drawings in Microsoft Word, see Section 5. If a graphic depends on its color, i.e., “red line indicates kW usage,” or “green line represents hours/day,” then you will need to change that. If you have a high-resolution computer screen shot or graphic that is only available to you in color, we will put it through a software graphics program that converts it to black and white. This will reduce the resolution and sharpness, so please remove all the color that you can. Do not use any color for text. In addition, minimize the use of shading. If you do use shading, please use a very light grey that will not obscure the text, data, or graphs. GRAPHIC RESOLUTION, SIZE AND FORMAT Provide each graphic or table in a separate file. For GRAYSCALE figures, the format should be TIF with 400 dpi resolution or density. For LINE DRAWINGS, the format should be TIF with 600 dpi resolution. Each graphic should be sized no smaller than the physical size it should be in the Journal. It is okay to submit a graphic a bit larger than desired finished size, but no larger than an 8-1/2 x 11 sheet of paper.

Vol. 29, No. 3

2014

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It is critical that both the 400 (or 600) dpi resolution and size be set for the graphic. Without this, the publishing program cannot produce a quality image. HOW DO I DO THIS? A common program that will accomplish these graphic requirements is Photoshop. This software will allow you to physically size your graphic, specify that the image is to be 400 (or 600) dpi density, and then save it as a TIF file. You can bring a JPEG or many other file formats into Photoshop, and then prepare them for saving to the required specifications. If you are familiar with using Photoshop, you can also do some nice graphic manipulations and improvements. You might want to try this for some of your screen shots. In Photoshop, open the image to be edited. If you need to crop the image, use the “Rectangular Marquee Tool” (the dashed square icon) and select the area of the image you want to save. Then select the Image, Crop function to crop the image. Adjust the size by selecting the IMAGE pull down, then IMAGE SIZE. Adjust the size using the DOUMENT Size dialog and remember to set the Resolution to 400 Pixels/Inch. Next, select the IMAGE pull down and then the MODE option and select GRAYSCALE. This will remove the color from the image. Save the file using the FILE dropdown and SAVE AS option. Select the TIF format, and then select the IBM PC option and LZW Compression. IMPROVING CHARTS AND GRAPHS FROM MICROSOFT EXCEL OR POWERPOINT Cutting and pasting a graph from Excel or an image from PowerPoint into Photoshop does not produce the highest quality image. A better solution, if you are using Microsoft Office 2003, is the following: Select the Image or Slide to output, then select File, Print. Select the “Microsoft Office Document Image Writer” to “print” to. Next select

78


the properties button that is to the right of the printer selection pull-down. Select the Advanced tab and the TIFF radio button. Select the “Superfine (300 DPI)” option in the adjacent pull-down. Fill in the file location to save the TIFF image to. You can then bring the TIFF image into Adobe Photoshop and crop and re-size it. IMPROVING LINE DRAWINGS FROM WORD If you have drawn line diagrams in Word that do not print out well, or give you the required density, you can click on the lines or shapes and get a menu that allows you to thicken the lines. This makes the lines appear thicker, bolder and blacker, and results in higher resolution diagrams. SCREEN SHOTS Screen shots are one of the biggest problems in getting high quality graphics for the journal. The color in the screenshots is a big problem in many cases. If there is any way to reduce the appearance of color in the screen shot, this will help the most. If the screenshot is a graph or plot from data from the EIS/ECS, one possibility is to take the original data and export it to Excel, (see note above about printing from Excel) and make an Excel graph out of it. This can then be imported to Photoshop to get the required density and size, and saved as a TIF. One important consideration in screenshots is to be sure your computer monitor is set to the highest resolution setting that is allowed. This captures the best quality screen shot. Another trick is to get the screen shot from a large or two-page computer monitor. This can then be resized to get a very good graphic. If you are familiar with using Photoshop, you can also do some nice graphic manipulations and improvements. You might want to try this for some of your screen shots.

Vol. 29, No. 3

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FINAL CHECK ON HOW YOUR GRAPHICS WILL LOOK The final check on how your graphics are going to look in the Journal is to print them using your printer. What your printout looks like is what it will look like at best in the Journal. So if it is sharp, clear and readable when you print it, then it should be fine for the Journal. But if your printout is fuzzy, blurry, and you can’t read some of the text or data, then the people who read the Journal will not be able to read it either. It is to the advantage of the overall Journal and to each author’s articles to be seen as a quality piece of work with quality graphics. We appreciate your effort in making sure that this occurs. THE EDITORS For any questions, please contact the editors as follows: Authors from the Americas, Asia and Australia contact: Dr. Jorge B. Wong Editor-in-Chief Distributed Generation and Alternative Energy Journal 8 Dawn Meadow Ct. Simpsonville, SC 29680 [email protected]

Authors from Africa, Europe and the Middle East contact: Dr. Ing. Jose Ramos-Saravia Associated Editor Distributed Generation and Alternative Energy Journal Dept. of Mechanical Engineering University of Zaragoza [email protected]

Dictionary of 21st Century Energy Technologies


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Gene Beck

This unique and comprehensive desktop reference addresses the diverse terms and acronyms that form the backbone of the 21st century energy technologies, applications of those technologies, and the emerging sustainability sector of the U.S. economy. The convergence of these disciplines has resulted in an explosion of specialized terms, acronyms, and technical lingo. The merging interactions encompass a wide range of legacy as well as emerging renewable energy technologies, including their integral finance and sustainability business components. The volume’s over 8,000 entries makes it the largest dictionary ever compiled on these specific subjects. Intended to help novices and experts alike cut through the confusion and better understand the vocabulary of this fast growing field, this comprehensive body of knowledge concisely explains both the technologies and the thousands of related new technical terms and acronyms. The result is a practical tool that should find a central place on the desk of anyone involved in energy, management and development of sustainability issues anywhere in the world. ISBN: 0-88173-736-4 8-1/2 x 11, 418 pp. Illustrated Hardcover $145 Order Code 0698

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IN THIS ISSUE: 5 From the Editor: Huge Solar Plant in California versus Affordable Solar Appliances for the Poor .

7 Thermodynamic Modeling of the Solar Organic Rankine Cycle with Selected Organic Working Fluids for Cogeneration; Suresh Baral and Kyung Chun Kim 35 Design and Performance Analysis of a Solar Air Heater with High Heat Storage; Abhishek Saxena, Vineet Tirth, and Ghanshyam Srivastava

56 Development of an Empirical Model for Assessment of Solar Air Heater Performance; P.K. Choudhury and D.C. Baruah

76 Guidelines for Graphics

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