Jun 28, 2014 - 1.4.4 Passive Solar Heating and Day Lighting ...... comparatively low temperature range from 50 °C-75 °C and to allow a continuous operation.
DESIGN AND SIMULATION OF SOLAR THERMAL COOLING AND HEATING SYSTEM A DISSERTATION
Submitted in partial fulfillment of credit course work requirement for the award of degree of
MASTER OF TECHNOLOGY in
ALTERNATE HYDRO ENERGY SYSTEMS
By ARVIND GUPTA (12512006)
ALTERNATE HYDRO ENERGY CENTRE INDIAN INSTITUTE OF TECHNOLOGY ROORKE ROORKEE – 247667 (INDIA) JUNE, 2014
DECLARATION
I hereby declare that the work presented in this dissertation entitled “ Design and Simulation of Solar Thermal Cooling and Heating System” submitted in partial fulfilment of requirements for the award of degree of Master of Technology in “Alternate Hydro Energy Systems”, submitted in Alternate Hydro Energy Centre, Indian Institute of Technology Roorkee, is an authentic record of my work carried out under the guidance and supervision of Dr R. P. Saini, Associate Professor, Alternate Hydro Energy Centre, Indian Institute of Technology, Roorkee.
I also declare that I have not submitted the matter embodied in this dissertation for award of any other degree or diploma.
Date: Place: Roorkee
(Arvind Gupta)
CERTIFICATE This is certified that the above statements made by the candidate is correct to the best of my knowledge.
Dr. R. P. Saini Associate Professor Alternate Hydro Energy Centre Indian Institute of Technology Roorkee Roorkee – 247667 ii
ACKNOWLEDGEMENT
I would like to express my deep sense of gratitude and indebtedness to my supervisor Dr R. P. Saini, Associate Professor, Alternate Hydro Energy Centre, Indian institute of Technology, Roorkee for guiding me to undertake this dissertation as well as providing me all the necessary guidance and support throughout this work. He has displayed unique tolerance and understanding at every step of progress, without which this work would not have been in the present shape. I am also thankful to all the staffs of Alternate hydro Energy Centre for their constant support and all my friends, for their help and encouragement at the hours of need.
Date: Place: Roorkee
(Arvind Gupta)
iii
ABSTRACT The cooling and heating systems have become indispensable part of our life. These systems are one of the major contributors to the overall energy consumption. The heating, ventilation and cooling (HVAC) market continues to rise since last few decades. Undoubtedly, heating and cooling systems add enormously to the CO2 emission. Fossil fuel depletion also is an issue associated with the fossil fuel based energy. With the concerns over usage of conventional electricity the solar powered cooling and heating systems have evolved which are driven by thermal input. There are many solar based cooling systems viz. (i) absorption, (ii) adsorption, (iii) desiccant and (iv) ejector systems. The literature review shows that among the various solar refrigeration methods absorption refrigeration has been extensively studied and preferred. Present work focuses on solar powered cooling and heating system based on absorption chiller. A solar thermal absorption system for cooling and heating system has been theoretically designed for Students’ Computer Laboratory, Alternate Hydro Energy Center at IIT Roorkee, India. Among the two popular refrigerant pairs NH3 -water and LiBr-water, LiBr-water has been chosen for the absorption chiller because it can utilize the thermal energy at relatively very low temperatures to produce cooling. The system in consideration is powered by hot water from flat plate collector field. The solar data and meteorological information have been generated with the guidelines from ASHRAE Handbook Fundamentals – 2013. The cooling load has been calculated using the radiant time series (RTS) method which was 8.77 kW at the peak. The absorption cycle was simulated in spreadsheet program after knowing the cooling demand and the heat input required to the system was obtained. It was seen that to meet a peak cooling demand of 8.77 kW, the flat plate solar collector field required was 54 m2 . The optimum volume of hot water storage tank was suggested to be integrated with the system. The system is able to meet the heating requirement in winter by directly supplying the hot water and by heating the air. A COP of 0.82 was achieved in the present study and it was seen that the system required minimal amount of electricity to pump the working fluid.
iv
CONTENTS Title
Page No.
Declaration
ii
Certificate
ii
Acknowledgement
iii
Abstract
iv
Contents
v
List of tables
viii
List of figures
ix
CHAPTER 1. INTRODUCTION 1.1 Energy Scenario
1
1.2 Renewable Sources of Energy
3
1.3 Solar Energy
5
1.4 Utilization of Solar Energy
5
1.4.1 Photovoltaic System
5
1.4.2 Solar Hot Water
5
1.4.3 Solar Electricity
6
1.4.4 Passive Solar Heating and Day Lighting
7
1.4.5 Solar Process Space Heating And Cooling
7
1.5 Solar Photovoltaic versus Solar thermal
8
1.6 Cooling Demand Trend
9
1.7 Basics of cooling/refrigeration
9
1.7.1 Types of Cooling Processes
10
1.7.1.1 Vapour Compression System
10
1.7.1.2 Vapour Sorption Systems
10
1.7.1.3 Gas Cycle Refrigeration System
12
1.7.2 Coefficient of Performance
13
1.8 Solar Cooling and Heating
14
1.8.1 General
14
1.8.2 Solar Electric Refrigeration and Cooling
15
v
1.8.2.1 Vapour Compression Systems
15
1.8.2.2 Thermoelectric Refrigeration
15
1.8.2.3 Stirling Refrigeration
16
1.8.3 Solar Thermal Refrigeration Systems 1.8.3.1 Absorption Cooling
16 16
1.8.3.1.1 Single-effect Absorption System
16
1.8.3.1.2 Half-effect Absorption System
17
1.8.3.1.3 Double-effect Absorption System
17
1.8.3.2 Adsorption System
18
1.8.3.3 Desiccant System
18
1.8.3.4 Ejector Refrigeration (Thermo-mechanical) System
20
1.9 CLOSURE
20 CHAPTER 2. LITERATURE REVIEW
2.1 Theoretical Studies
21
2.2 Experimental Studies
23
2.3 Summary of Review
25
2.4 CONCLUSION
36
2.5 OBJECTIVE
36
CHAPTER 3. SOLAR ABSORPTION COOLING AND HEATING SYSTEM 3.1 System description
37
3.1.1 Working of the system
38
3.1.2 Components of the system
38
3.1.2.1 Solar Collector
38
3.1.2.1.1 Stationary Collector
39
3.1.2.1.2 Sun Tracking concentrating Collector
42
3.1.2.2 Hot Water Storage Tank
44
3.1.2.3 Absorption Chiller
44
3.1.2.4 Fan Coil Unit
46
3.1.2.5 Cooling Tower
47
3.2 CLOSURE
47
CHAPTER 4. METEOROLOGICAL DATA, SOLAR DATA AND LOAD CALCULATION 4.1 Site
48
vi
4.2 Site details
50
4.3 Load calculation
51
4.3.1 Meteorological data for Roorkee
51
4.3.2 Solar radiation calculation over Roorkee
53
4.4 Cooling load
72
4.4.1 Heat gain through exterior surface
73
4.4.1.1 Wall Cooling Load
76
4.4.1.2 Roof Cooling Load
78
4.4.1.3 Window Cooling Load
78
4.4.2 Lighting/Appliances Cooling Load
81
4.4.3 Internal/Occupants Cooling Load
83
4.4.4 Total Cooling Load
84
4.5 Heating load
86
4.6 CLOSURE
87
CHAPTER 5. DESIGN OF SOLAR BASED LiBr-WATER ABSORPTION COOLING SYSTEM 5.1 Design of absorption chiller
88
5.1.1 Mathematical Modelling of Different Chiller Components
90
5.2 Parameters calculations
91
5.3 Solar Collector area
95
5.4 Flow rate of water
96
5.5 Volume of Hot Water Storage Tank
97
5.6 RESULTS AND DISCUSSION
97
CHAPTER 6. CONCLUSION AND RECOMMENDATION 6.1 Conclusion
98
6.2 Recommendations
99
List of Publications
100
References
101
vii
LIST OF TABLES No.
Caption
Page
1
Summary of literature review
26
2
Solar energy collectors’ summary
43
3
Meteorological Data of Roorkee, India
51
4
Surface Orientations and azimuths, measured from south
57
5
Solar radiation calculation data table over Student computer Lab,
60
AHEC, IIT Roorkee (on June 21) 6
Beam, diffuse and reflected components of radiation on Computer
63
Lab’s wall & roof (June 21) 7
Solar radiation calculation data table over Student computer Lab,
66
AHEC, IIT Roorkee (January 21) 8
Beam, diffuse and reflected components of radiation on Computer
69
Lab’s wall & roof (January 21) 9
Sol-air temperature calculation for the wall and the roof on June 21
75
10
Hourly heat input profile for the vertical wall using the sol-air
77
temperature with indoor temperature 25⁰C 11
Hourly heat input profile for the roof using the sol-air temperature
79
with indoor temperature 25⁰C 12
Window hourly cooling load calculation
80
13
Heat gain from lighting and appliances
81
14
Lighting/Appliances heat gain profile based on occupancy schedule
82
and hourly load 15
Hourly lighting/appliances heat gain and cooling load
83
16
Total cooling load hourly profile for 24hrs on June 21
85
17
Design parameters for absorption chiller
89
18
All calculated parameters of the absorption chiller
94
19
Energy Flows in the absorption chiller
95
20
Radiation components over the collector at an angle of 40 ⁰
96
viii
LIST OF FIGURES No.
Figure Caption
Page
1.
Reserve to production ratio of Coal in years(R/P ratio)
1
2.
Reserve to production ratio of Oil in years (R/P ratio)
2
3.
Reserve to production ratio of Natural Gas in years (R/P ratio)
2
4.
World CO2 emission by sector in 2010
3
5.
Solar photovoltaic panels
6
6.
Flat plate thermal system for water heating
6
7.
Parabolic-trough collector
6
8.
Dish/engine collector
6
9.
A Solar tower
7
10.
Solar spectrum used in a PV system
8
11.
Basic vapour-compression refrigeration system
10
12.
Vapour absorption refrigeration system
11
13.
Schematic of a closed reverse Brayton cycle
12
14.
A solar assisted single-effect absorption cycle
16
15.
Schematic of solar assisted double-effect water/LiBr absorption
17
system 16.
Basic adsorption cycle
18
17.
Schematic drawing of a desiccant cooling system
19
18.
Ejector refrigeration system
20
19.
Solar hybrid tri-generation plant
21
20.
NH3-water absorption chillers with concentrating collector
25
21.
Solar absorption cooling and heating system layout
37
22.
Pictorial view of a Flat plate collector
40
23.
Exploded view of a flat plate collector
40
24.
Schematic diagram of compound parabolic collector
41
25.
Schematic diagram of an evacuated tube collector
42
26.
Sun tracking concentrating collectors
43
27
An absorption cooling system
45
28
A fan coil unit
46
29
Flow chart for the design process
48
30
AHEC, IIT Roorkee
49 ix
31
Bird-eye view of AHEC, Roorkee
49
32
Plan of the Computer Lab, AHEC, IIT Roorkee
50
33
Panoramic view of Computer Lab, AHEC, IIT Roorkee
50
34
Monthly temperature at the Roorkee by Meteonorm 7
52
35
Motion of earth around sun
54
36
Various Solar angles for a surface
55
37
Incident radiation over Roorkee on June 21
64
38
Total radiation on vertical wall of lab on June 21
64
39
Hourly radiation on Horizontal surface on June 21 at Roorkee
65
40
Total incident radiation on Horizontal surface on June 21 at Roorkee
65
41
Hourly radiation over Roorkee on Jan 21
70
42
Hourly radiation on vertical wall of lab on Jan 21
70
43
Total hourly incident radiation on vertical wall of lab on Jan 21
71
44
Hourly incident radiation on Horizontal surface (roof) of lab on Jan
71
21 45
Total hourly incident radiation on horizontal surface (roof) of lab on
72
Jan 21 46
Overview of Radiant Time Series method
74
47
Hourly cooling load on June 21
86
48
LiBr-Water Absorption Cooling System
88
x
CHAPTER 1
INTRODUCTION
1.1 Energy Scenario The importance of energy in sustaining life is unquestionable. Energy is the most prime requirement of humankind in today’s developing world and with this world being continuously more and more populous the energy demand is on surge exponentially. Population and income growth are the key drivers behind growing demand for energy. As per the projections made, by 2030 world population would reach 8.3 billion, which means an additional 1.3 billion people will require energy; and world income in 2030 is expected to be roughly double the 2011 level in real terms [1]. Also any country’s economic and industrial development requires energy. The major sources of energy today are fossil fuels and the oil being the world’s leading fuel at 33.1% of global energy consumption [2]. The other fossil fuels are coal and gas which are burnt to extract the energy. The earth crust possesses a limited reserves of fossil fuels which once used can’t be replenished in a practicable duration of time instead they take millions of years to replenish and that’s why they are called as non-renewable sources of energy. With the current rate of consumption of fossil fuels all the reserves are going to be exhausted in coming years. This poses a risk of energy security in future.
300 237.79
250
Coal
244.41
200
2007
150
123.89
128.57
2008 109.47
100 50
2009 2010
50.95
2011 2012
0
Asia Pacific Europe and Middle East North Eurasia and Africa America
South and Central America
World
Figure 1. Reserve to production ratio of Coal in years(R/P ratio) [3]
1
Oil
140
121.91
120 100
2008
78.06
80
2009 52.93
60 40
13.64
20
2011
38.68
37.7
2010 2012
22.36
0
Africa
Asia Europe & Middle North Pacific Eurasia East America
S. & Cent. America
World
Figure 2. Reserve to production ratio of Oil in years (R/P ratio) [4]
Natural Gas 250 200 2008 146.79
150
2009 2010 2011
100 67.06 50
56.41
42.85
31.52
55.68
12.09 0
Africa
Asia Europe & Middle North Pacific Eurasia East America
S. & Cent. America
World
Figure 3 Reserve to production ratio of Natural Gas in years (R/P ratio) [5]
2
2012
In recent decades, apart from the energy security threats there have been several concerns over the usage of fossil fuels regarding the environment. The burning of fossil fuels causes serious damage to our environment. Their usage on large scale over the years is the reason for acid rain, ozone layer depletion and global climate change. The burning of coal, natural gas, and oil for electricity and heat is the largest single source of global greenhouse gas emissions (GHG) [6]. The analysis of data reveal that two sectors produced nearly two-thirds of global CO 2 emissions in 2010: electricity and heat generation accounted for 41% [7].
Figure 4. World CO 2 emission by sector in 2010 [7] *Other includes commercial/public services, agriculture/forestry, fishing, energy industries other than electricity and heat generation, and other emissions not specified elsewhere. The average temperature of earth has risen by about 0.6 K due to GHG, according to United Nations Intergovernmental Panel on Climate Change (IPCC) [1]. These concerns have compelled to search for the alternative sources of energy which are nonpolluting and neverending. These are called renewable sources of energy. The technologies to harness energy from these sources are of recent research topic across the globe.
1.2 Renewable Sources of Energy Renewable sources of energy are theoretically non exhaustible sources of energy which exist naturally. These are (i) Solar. (ii) Hydro, (iii) Biomass, (iv) Geothermal, (v) Wind and (vi) Ocean energy. Most renewable energy comes either directly or indirectly from the sun. Renewable energy currently constitutes 15% of the global energy mix [8].
3
Solar energy is harnessed in two ways: (i) Solar Photovoltaic and (ii) Solar thermal. Solar photovoltaic cells convert the solar radiation intercepted by the panel into electricity directly while in solar thermal applications the solar radiations are intercepted and converted into heat which is then carried away by some fluid for various purposes. Hydropower utilizes the potential and kinetic energy of water to generate electricity with the help of hydro turbines. Biomass energy is the energy which is extracted from organic materials by combustion process. Wood is the largest biomass energy resource. The other sources of biomass include plants, agriculture or forestry residues, and the organic constituents of municipal and industrial wastes. The fumes from landfills can also be used as a biomass source [9]. Geothermal energy is the thermal energy stored in Earth. Geothermal energy can be extracted from the shallow ground, hot water and hot rock found in the earth crust. Extremely high temperatures magma exists in the deeper portion of the earth crust, which can be a great source og heat. The upper layer of earth crust, approximately upper 10 feet of the earth's surface maintains nearly constant temperature which ranges from 10° to 16°C. Geothermal heat pumps or ground source heat pump use this thermal reservoir to heat and cool buildings [10]. Wind energy is harvested by tapping the kinetic energy of wind with the help of wind turbines to generate electricity. The ocean can produce two types of energy: (i) thermal energy and (ii) mechanical energy. The sun warms the upper layer of ocean more than the deep water and thus a temperature gradient is created. This difference in temperature is the cause for thermal energy. Ocean mechanical energy is basically caused due to tides. The gravitational pull of moon and sun drive the tides. This rise and fall of tides exhibits enormous amount of mechanical energy which can be converted into electrical energy by employing mechanical devices. Among all these green and naturally available sources, solar energy stands out. Solar energy is ubiquitous unlike other renewable energy sources and hence can be harnessed everywhere. It takes lesser space to employ solar energy technologies in comparison to other renewable technologies. Solar energy requires lesser maintenance. Most countries now accept that solar energy has huge potential because it is clean, freely available. Among all the renewable sources of energy, solar energy is the most abundant one and is available in both direct as well as indirect forms. 4
1.3 Solar Energy The Sun generates and emits energy at a rate of 3.8x1023 kW and approximately 1.8x1014 kW of it is intercepted by the earth. About 60% of intercepted amount reaches the earth surface. The remaining portion is either reflected back into space or gets absorbed by the atmosphere. Of 0.1% of this energy, if converted at an efficiency of 10% would generate four times the world’s total generating capacity of about 3000 GW [11]. Solar energy’s technical potential is hundreds of times greater than the world’s total electricity use. Solar energy is the result of electromagnetic radiation released from the sun by nuclear fission reaction. It reaches earth in form of visible, infrared and ultraviolet radiation. These radiations can be intercepted and converted into other useful forms of energy. There is a range of technologies [12] that have been developed to tap the solar energy.
1.4 Utilization of Solar energy Solar radiation can be utilized in several ways using different types of systems, either by converting it to other useful forms of energy or directly as heat, light etc. 1.4.1
Photovoltaic (Solar cells) Systems Solar cells convert sunlight directly into electricity. They are made of materials called
as semiconductors. When sunlight falls on such materials, the solar energy energizes electrons to loose from their atoms and to flow through the material thus producing electricity. This phenomenon of converting light to electricity is called the photoelectric effect. 1.4.2
Solar Hot Water The solar energy can be used to heat water for domestic use in buildings and swimming
pools, in general cases. Solar water heating systems for buildings in general, have a solar collector and a storage tank. The most commonly used collector in these systems are flat-plate collectors. Solar water heating systems can be either active or passive. Active systems require the input of mechanical energy to maintain the flow of the heat carrying fluid, thus they are equipped with the pumps, while passive systems work automatically without any external mechanical energy input by naturally created density difference with the help of gravity.
5
Figure 5. Solar photovoltaic panels [13]
Figure 6. Flat plate thermal system for water heating [14]
1.4.3
Solar Electricity These systems use sun’s heat to generate electricity. Solar electricity plants require high
temperature steam which can rotate the large turbine coupled with the generator to generate electrical power.
Figure 7. Parabolic-trough collector [15]
Figure 8. Dish/engine collector [16]
The higher temperature of steam is achieved by using concentrating type collectors which are of mainly three types: parabolic-trough, dish/engine, and power tower. In all kinds of concentrating collectors the solar energy is intercepted and concentrated to heat a heat carrying fluid. The concentrator directs the radiation onto the receiver where the solar radiation gets absorbed.
6
Figure 9. A Solar tower [17]
1.4.4
Passive Solar Heating and Day lighting In this type of application solar energy is used to heat and light the buildings. The
passive systems need no mechanical input as generally required by the solar thermal system for moving heat carrying fluids. Also, these types of systems can be integrated into building structure. The windows and rooms can act as the collectors and the storage system. The southfacing portion of the building always receive the maximum sunlight. That is why, the structures designed for passive solar heating usually have large windows facing south. The floors and walls get heated up during the day and release the heat at night. The sunspaces and trombe walls are the other passive solar heating design options. A sunspace is same as a greenhouse which is provided on the south part of a building while a trombe wall is a thick black painted wall facing south, which is made of a material that can absorb a good amount of heat. A glazing is provided a few inches in front of the wall to hold the heat. Day lighting is simply the use of natural sunlight to lighten up a building's interior.
1.4.5
Solar Process Space Heating and Cooling This application includes industrial and commercial utilization of the solar thermal
energy. Heating systems provide hot water or space heating for nonresidential buildings. A 7
typical system may include solar collectors, a heat exchanger, and a large storage tank. The heat from a solar collector can also be used to cool a building. Solar absorption coolers use a similar approach. Solar energy can also be used with evaporative coolers to extend their usefulness to more humid climates, using desiccant cooling.
1.5 Solar Photovoltaic versus Solar thermal The solar thermal systems are preferred over solar photovoltaic because solar thermal systems can utilize more of solar radiation. In photovoltaic panels, most of the intercepted radiation is lost in increasing the temperature of panel and cannot be converted into electricity. As the aim of using a thermal collector is to convert solar energy directly into heat, thermal collectors have no such limitation. The solar thermal collector efficiency can be as much as 95%. Figure 10 roughly shows the distribution of various radiations when the solar rays intercept a solar PV collector.
Figure 10. Solar spectrum used in a PV system [20] As per the Figure 10, 65% of the radiation falling over the PV collector is lost as heat and merely 35% of the radiation can be converted into electricity. Hence, it has also be seen that the collection efficiencies of thermal collectors are more than the double that of solar photovoltaic collectors [20]. The promise of the solar thermal system is obviously huge, therefore, the present work focuses on the refrigeration techniques, strictly based on the solar thermal systems.
8
1.6 Cooling demand trend In recent times there has been seen a huge demand in cooling and refrigeration sector because of changing climate. According to Japan Air Conditioning, Heating & Refrigeration News (JARN) estimates, the global market for air conditioners was about 89 million units in 2010, a 21% increase compared with 2009 [18]. Up to now, majority of cooling and refrigeration systems rely on electricity from fossil fuels and the electricity consumption in this sector is about 15% of worldwide electricity use and this percentage is even higher (more than 20%) in some developed countries[19]. Thus, this sector contributes to the CO 2 emission and global warming in the large magnitude. Availability of alternative technologies of cooling and refrigeration which do not depend on the fossil fuels will help in reducing the carbon emission and control global warming. These facts have attracted the attention of researchers worldwide towards the solar options. The solar options for cooling systems have been looked upon extensively in the past. It becomes however, very important to understand the various cooling and refrigeration technologies first, before exploring the solar based cooling technologies. The following section briefly discusses various cooling methods employed in today’s world.
1.7 Basics of cooling/refrigeration Refrigeration or cooling can be defined as the process of achieving and maintaining a temperature below that of the surroundings. One of the most important applications of cooling has been in the air conditioning systems for human thermal comfort and other important application being the preservation of perishable food products by storing them at low temperatures. Air conditioning means the simultaneously controlling air’s temperature, moisture content, cleanliness, odor and circulation, as required. The air can be conditioned for the occupants or a process or product in the space. The cooling can be achieved by several methods suchas vapor compression, vapour sorption, gas cycle systems etc.
9
1.7.1
Types of cooling processes 1.7.1.1 Vapour Compression System Vapour compression refrigeration or cooling systems are the most commonly used
among all refrigeration systems. A vapour compression system consists of (i) a compressor, (iii) a condenser, (iii) a thermal expansion valve, and (iv) an evaporator. In such systems, the working fluid that is refrigerant, undergoes phase change to produce the cooling effect. At low pressure, the refrigerant passes through the evaporator where it absorbs the heat and gets evaporated at low temperature and thus producing the cooling effect. The vaporized refrigerant is then made to pass through the mechanical compressor to increase its pressure and temperature. The refrigerant vapor then passed through the condenser to lose its heat to the environment or the cooling medium. The condensed refrigerant then gets expanded at the throttle valve to finally again come to evaporator. This cycle repeats in a cooling system. The input to such system is the mechanical energy to run the compressor. Hence these systems are also called as mechanical refrigeration systems.
Figure 11. Basic vapor-compression refrigeration system [21] 1.7.1.2 Vapour sorption systems Sorption refrigeration system utilizes physical or chemical attraction between a pair of substances to produce refrigeration effect. A sorption system converts the thermal energy to the cooling effect. Such systems use heat as the input to produce the cooling or refrigerating effect. The working substance is a solution of two substances, one sorbent and one sorbate. Among the two, the sorbate has lower boiling point and plays as the refrigerant. Absorption 10
refers to a sorption process where a liquid sorbent absorbs refrigerant molecules into its inside and changes physically and/or chemically in the process. Vapour absorption refrigeration system is a closed sorption system. The absorption cycle has three main processes (i) Evaporation: In evaporation, the liquid refrigerant gets evaporated at low pressure by extracting the heat from its surroundings. (ii) Absorption: Here, the evaporated refrigerant gets absorbed in another liquid called the sorbent which reduces its partial pressure in the evaporator and letting more refrigerant to evaporate. (iii) Regeneration: At the generator, the heat input makes the refrigerant laden solution to give off the refrigerant vapour. The vapour is then condensed in the condenser to again supply the liquid refrigerant in the evaporator.
Figure 12. Vapour absorption cooling system [21] Fig. 12 shows a schematic diagram of a sorption system or vapour absorption system. When the refrigerant leaves the evaporator, it is absorbed by sorbent in absorber and is then pumped to generator. After reaching generator, the refrigerant is separated by the application of heat. The refrigerant vapour then goes to condenser where the temperature of vapour refrigerant decreases as it loses its heat to the cooling medium around the condenser. Adsorption cooling systems is another type of sorption system where the sorbent is a solid which attracts refrigerant vapour onto its surface. Desiccation system is another sorption process where a sorbent is a desiccant which can absorb the moisture content from humid air. Desiccant cooling is basically used where humidity control is of major concern.
11
1.7.1.3 Gas Cycle Refrigeration System Air cycle refrigeration systems are the systems which employ an air standard cycle to produce the cooling effect. As the name implies, the air is used as the working fluid.in which a gas is used as the working fluid. There is no phase change of the working fluid takes place in such systems. Reverse Carnot cycle and reverse Brayton cycle may be used to achieve refrigeration employing a gas. An ideal reverse Carnot cycle will consist of four processes: (i) isentropic compression, (ii) reversible isothermal heat rejection, (iii) isentropic expansion and (iv) reversible isothermal heat absorption. Carnot cycle is a theoretical ideal cycle which has several practical constraint. The main problem with Carnot cycle with a gas is the difficulty of achieving isothermal heat transfer processes.
Figure 13. Schematic of a closed Reverse Brayton cycle [22] A Reverse Brayton refrigeration system is an important cycle frequently employed in gas cycle refrigeration systems. This may be thought of as a modification of reversed Carnot cycle, as the two isothermal processes of Carnot cycle are replaced by two isobaric heat transfer processes. In the Fig. 13, the working substance, air, enters the compressor at 1 and leaves at 2. The external work, Wnet is given at the compressor. The compressed air passes through condenser, HTHX to lose its heat to the cooling medium. The condensed gas is made to pass through the turbine T where expansion (3-4) takes place and the temperature and the pressure
12
is reduced. After turbine, the gas enters the evaporator LTHX, to absorb the heat, where eventually cooling effect (4-1) is produced. This cycle repeats to produce continuous cooling. 1.7.2
Coefficient of Performance
The performance of cooling devices is measured as Coefficient of performance (COP). COP is an alternative term to efficiency used in thermodynamics. COP of a refrigerator is given as the ratio of refrigeration effect to the external work done to the system. Therefore the COP of different solar cooling systems can be defined as follows: For solar electric compression refrigeration,
COPele =
Qe W
Where; COPele = COP of solar electric compression refrigeration; Q e = cooling effect in kW; W = electrical power input in kW. For solar mechanical compression refrigeration,
COPmech =
Qe Qboil
Where; COPmech = COP of solar mechanical compression cooling system; Q boil = input heat to boiler of Rankine cycle in kW. For solar absorption refrigeration,
COPab =
Qe Q gen
Where; COPab = COP of absorption cooling; Q gen = heat input to generator in kW. For solar adsorption refrigeration,
COPad =
Qe Qde
Where; COPad = COP of adsorption cooling; Q de = heat input at the generator in kW. For solar desiccant cooling,
13
COPdesi =
m a (ho − hs ) Qregen
Where; COPdesi = COP of desiccant cooling system; m a = mass flow rate of supply air in kg/s; h o = specific enthalpy of outdoor air in kW/kg; h s = specific enthalpy of supply air (kW/kg); Q regen = heat input for regeneration in kW.
1.8 Solar Cooling and Heating 1.8.1 General Refrigeration, including air conditioning, is necessary for life and will continue to expand worldwide as the climate changes have been seen. In recent years there has been a sharp demand of refrigeration technologies due to increased needs of refrigeration & air conditioning for both commercial and residential purposes during hot season. And as already discussed fifteen percent of the electricity produced in the whole world is employed for refrigeration and air-conditioning processes of various kinds [23] and also there is a lack of electrical energy and storage to accommodate high energy consumptive systems such as refrigeration and cooling. This gap between demand and supply cannot be filled by merely efficiency improvement. Considering that cooling demand increases with the intensity of solar radiation, solar based refrigeration can be considered as a logical solution. The energy crisis has opened the doors for solar energy to not only meet the peak demands but also refrigeration and cooling needs. Air conditioning of buildings is responsible for a large percentage of the greenhouse and ozone depletion effect, as refrigerant harmful gases are released into the atmosphere from conventional cooling systems. The need to implement advanced new concepts in building air conditioning systems is more crucial than ever today. Solar cooling systems have the advantage of using absolutely harmless working fluids such as water, or solutions of certain salts. They are energy efficient and environmentally safe. Solar assisted refrigeration and cooling technologies can contribute by being a part of solutions for mitigating global warming. Therefore, solar cooling technology development has become a key research area worldwide. Solar refrigeration and cooling can be achieved fundamentally in two ways: a) Solar electric refrigeration b) Solar thermal refrigeration
14
1.8.2 Solar Electric Refrigeration or Cooling Solar electric systems use electricity as an input to the systems. 1.8.2.1 Vapour Compression Systems Vapour compression systems are the most widely used cooling systems in the world. As already discussed vapour compression cooling systems use a liquid refrigerant which is circulated through various components of the system. The refrigerant absorbs the heat at the evaporator and cools the space and rejects the heat at the condenser. In a solar based vapour compression refrigeration system, the electrical work required for compressor is supplied by solar photovoltaic (PV) panels.
1.8.2.2 Thermoelectric Refrigeration This kind of refrigeration systems are based on the Peltier–Seebeck effect. In Peltier– Seebeck effect, the the temperature difference develops the electrical voltage. In other words, the thermoelectric effect is the direct conversion of temperature differences to electric voltage and vice-versa. Thermoelectric elements are made of semiconductors. These semiconductors may be bismuth telluride (Bi 2 Te 3 ) and antimony telluride (Sb 2 Te 3 ) alloys. These kind of cooling systems don’t have any moving parts or the refrigerant liquid which makes them compact in size. They find their application in electronic cooling, portable refrigerators and in space applications like satellite and space ships where the size of a cooling system is extremely limited. COP of such systems is at present very low. COP can range from 0.3 to 0.6[1].
1.8.2.3 Stirling Refrigeration Stirling cycle is gas cycle which consists of four thermodynamic processes same as the vapor compression cycle. A Stirling engine is a heat engine which operates by cyclic compression and expansion the working fluid. The compression and expansion take place at different temperature levels in such a way that heat energy is converted to mechanical energy. The reverse Stirling cycle can be employed in a system for cooling effect cooling effect by connecting it to the PV panels.
15
1.8.3 Solar Thermal Refrigeration or Cooling Solar thermal systems use solar heat rather than solar electricity to produce cooling effect. 1.8.3.1 Absorption Cooling Absorption refrigeration has been most frequently and widely adopted for solar thermal refrigeration [24]. It is same as the vapour compression cooling system except for the compression process. The compressor in the vapour compression is replaced by an absorbergenerator unit. An absorber, pump, expansion valve, heat exchanger and generator are the main parts of the absorption cooling system. In a solar absorption cooling system, the heat input from the solar collector is used at the generator to separate the refrigerant vapour from the solution. In general, LiBr-water and NH3-water solution are used as the refrigerant solution in the absorption systems [25]. According to the solution regeneration and thermal operation cycle, the absorption systems can be divided into three categories: single-effect, half-effect and double-effect solar absorption cycles. 1.8.3.1.1 Single-effect Solar Absorption System The literature shows that the single effect solar absorption systems with the flat plate collector and evacuated tube collector are widely implemented among all the absorption cooling systems [26]. Figure 14 shows a single effect absorption cycle. The cycle starts at the absorber where it absorbs the refrigerant vapour from the evaporator to make a strong solution. The strong solution is then pumped through a heat exchanger to the generator part. The refrigerant vapour gets separated at the generator by getting heated from the heat input.
Figure 14. A solar assisted single-effect absorption cycle [27] 16
1.8.3.1.2 Half-effect Solar Absorption System The half effect systems are same as the single effect but different for the facts that the half effect systems can operate at lower temperatures and the COP of such systems is approximately half to that of the single effect systems [28]. 1.8.3.1.3 Double-effect Solar Absorption System Double-effect absorption cooling systems need heat source at higher temperature. The performance of double-effect absorption cooling systems is better than the single-effect and half-effect systems. Figure 15 illustrates a double effect absorption cooling system. A double effect system has two generators and two heat exchangers as shown in the Fig.15. In Fig. 15, the cycle starts at generator G-I which provides heat to generator G-II. The condenser passes the refrigerant and leads to the evaporator V by losing its heat. The refrigeration occurs at the evaporator and then the solution passes through the heat exchangers HX-I and HX-II from the absorber Ab to G-I by pumping. Through this process, HX-II can pass the fluids to G-II and then G-II passes to HX-I. The complete cycle follows three different pressure levels: high, medium and low.
Figure 15. Schematic of solar assisted double-effect WATER/LiBr absorption system [29]
17
1.8.3.2 Adsorption System In adsorption process, the refrigerant vapour is adsorbed over the solid surface by just physical attraction. There is no chemical changes involved in the adsorption process. The material getting adsorbed is known as the adsorbate and the surface over which adsorption takes place is known as the adsorbent.
Figure 16. Basic adsorption cycle Commonly used adsorbents are zeolite, silica gel, activated carbon and etc. The adsorbent can catch and hold the refrigerant because of the high surface volume ratio and high porosity of the adsorbent material. When saturated, the refrigerant vapour can be can be regenerated by simply applying the heat input. In the simplest case, an adsorption refrigerator can be assumed as a two vessels connected to each other as shown in Figure 16, one of which is filled with adsorbent and adsorbate. 1.8.3.3 Desiccant Systems Desiccant cooling systems process water vapor in the earth's atmosphere to produce cooling. Since mass transfer occurs between the system and its environment, they are commonly referred to as "open-cycle" systems. These systems use a liquid or solid material called a desiccant to remove water vapor from the air. The process by which water is removed 18
is most often adsorption on the solid desiccants and absorption in the liquid desiccants. Various desiccants are available in the market, these can be liquid as well as a solid, as said earlier. All the water absorbing sorbents can act as a desiccant to remove moisture. Commonly used desiccants are silica gel, activated alumina, zeolite, LiCl and LiBr. The main components of a solar assisted desiccant cooling system are shown in the figure 17. Warm and humid air enters the slowly rotating desiccant wheel and is dehumidified by adsorption of water (1-2). Since the air is heated up by the adsorption heat, a heat recovery wheel is passed (2-3), resulting in a significant pre-cooling of the supply air stream. Subsequently, the air is humidified and thus further cooled by a controlled humidifier (3-4) according to the set-values of supply air temperature and humidity. The exhaust air stream of the rooms is humidified (6-7) close to the saturation point to exploit the full cooling potential in order to allow an effective heat recovery (7-8). Finally, the sorption wheel has to be regenerated (9-10) by applying heat in a comparatively low temperature range from 50 °C-75 °C and to allow a continuous operation of the dehumidification process. Flat-plate solar collectors are normally applied as heating system in solar assisted desiccant cooling systems [30].
Figure 17. Schematic drawing of a desiccant cooling system [30]
19
1.8.3.4 Ejector Refrigeration (Thermo-mechanical) Systems The ejector system represents the thermo − mechanical cooling which is another alternative to the conventional vapour compression cycle. It is a thermally driven technology by low grade thermal energy. They have been developed by replacing the compressor with a boiler, an ejector and a pump in the vapour compression cycle. Their utmost advantage is capability of producing refrigeration from waste heat or solar energy by using heat source at temperatures above 80 °C.
Figure 18. Ejector refrigeration system [25]
1.9 CLOSURE Present chapter starts with the discussion on present energy scenario and related concerns. The repercussions of using conventional energy sources have been highlighted and renewable sources of energy have been briefly introduced. The aim of the chapter is to explain the contribution of cooling and heating systems in the CO 2 emissions and to introduce the solar energy based alternatives. Prior to discussing solar cooling options, various available cooling technologies have been presented. Out of the two solar cooling options, solar electric and solar thermal, the solar thermal options have been said preferable as it utilizes maximum of the radiation. The next chapter will follow with the literature review on solar thermal based cooling and heating systems with a special focus on absorption technology based systems.
20
CHAPTER 2 2.1
LITERATURE REVIEW
Theoretical studies S.O. Enibe [32] and R. Best & N. Ortega [33] discussed various solar cooling techniques
and stressed that the cost was the main concern for wide implementation. F. Assilzadeh et al. [37] simulated in TRNSYS program a LiBr-water based solar absorption cooling system. Evacuated tube collector was considered and it was found that a hot water storage tank is essential for the system. D.S. Kim & C.A. Infante Ferreira [1] and J.M. Abdulateef et al. [47] found that singleeffect solar thermal absorption cooling system was the best option. A single effect LiBr-water solar absorption system was designed by M. Mazloumi [39] with the capacity of 17.5 kW. The thermodynamic result showed that collector mass flow rate affected the thermal storage capacity. Two case studies were presented by Desideri et al. [41] in which industrial refrigeration and solar cooling system for hotel users were considered. For industrial refrigeration authors suggested LiBr-water absorption chiller with evacuated tube collector and storage tank and for hotels a hybrid, or thermo-solar, tri-generation plant which produces electricity, heating and cooling effect at the same time.
Fig 19. Solar hybrid tri-generation plant [41]
21
Ursula Eicker and Dirk Pietruschka [42] also developed a full simulation model for absorption cooling systems with a stratified storage tank. They found that doubling of mass flow decreased the required collector surface area and thus solar thermal system costs by 30%. Tiago Mateus and Armando C. Oliveira [44] studied integrated solar absorption cooling and heating systems using the TRNSYS simulation. The study found that the initial cost of the system was high although the exploitation cost was found to be considerable low. Boonrit Prasartkaew and S. Kumar [49] did the simulation by mathematical modeling of a solar–biomass based single effect LiBr–water absorption. They found that environmental impacts can be significantly reduced if investment is good on the solar collectors. T. Tsoutsos et al. [55] did the performance and economic evaluation of a LiBr-water solar heating and cooling system using TRNSYS. Investment cost was quite high as per their observation. Marwan Mokhtar et al. [56] evaluated the techno-economic performance of the solar collector/chiller system. Authors considered 25 feasible combinations of solar energy collection and cooling technologies and concluded that two very influential parameters determine the most economical solar cooling option; the cost and the performance of the refrigeration technologies. Lin et al. [64] performed a numerical investigation of a two-stage NH 3 -water air-cooled R
R
absorption solar refrigeration system with 5 kW cooling capacity. The simulation results showed that thermal COP is 0.34 and electrical COP is 26 under a typical summer condition with 85 °C hot water supplied from solar collector. X.Q. Zhai et al. [65] presented various cooling modes. As per their study half-effect absorption chillers and two-stage absorption chillers seem to be more suitable for air-cooled solar absorption cooling systems in hot and dry regions. They concluded that two parameters determine the most economical solar cooling option; the cost and the performance of the refrigeration technologies. They gave the specific hot water tank volume as 0.01–0.08 m3/m2 . P
P
P
P
Emilio J. Sarabia Escriva et al. [66] obtained an equation that shows the dependence of the generator/solar-collectors equilibrium temperature on basic design parameters of single effect absorption cooling machines. V. Boopathi Raja and V. Shanmugam [27] concluded that LiBr–water single effect absorption cooling method is more suitable for domestic purpose and flat plate and evacuated tube
22
solar collectors are more reliable and economical for the system. Authors also quoted the optimum generator and evaporator temperatures as 75 to 92 °C and 5–10 °C respectively. Ullah et al. [73] reviewed various solar thermal refrigeration systems and concluded that COP of absorption cooling systems is better than that of adsorption systems. F.J. Cabrera et al. [74] reviewed and stressed consideration of parabolic trough collectors (PTC) in process heat and double effect absorption chillers because PTCs solar fraction was found to be higher than that of other collectors and flat-plate collectors had the lowest. Renato M. Lazzarin [75] performed the analysis to compare the various popular solar cooling techniques in terms of the overall efficiency and the investment cost. The evaluation suggested that the investment costs are similar for PV solar cooling systems, flat plate collector, evacuated tube collector and parabolic trough collector driven absorption for a daily production of 10 kWh cooling. Ioan Sarbu & Calin Sebarchievici [29] stressed that solar hybrid cooling systems can provide higher capacity and better thermal COPs by eliminating some of the problems encountered with individual working pairs.
2.2 Experimental studies K Sumathy et al. [35] proposed a low driving temperature strategy. A two-stage lithium bromide absorption chiller was developed successfully and a practical integrated solar cooling and heating system with a 100 kW was constructed. This two-stage chiller could be driven by hot water of temperature from 75 to 60°C. Zambrano et al [38] developed a model of a single-effect absorption chiller of 35 kW capacity using hot water from flat plate collectors which fed. A dynamic model of a solar cooling plant was developed and validated with real data with satisfactory results. Ahmed Hamza H. Ali et al. [40] did the performance assessment of a 35.17 kW singleeffect LiBr-water absorption chiller using evacuated tube collectors in operation for the duration of five years. Hot water and cold water storage tanks were included in the system. The specific collector area of 4.23 (m2/kW cold ) was reported to harvest maximum possible solar heat fraction. P
P
R
R
Francis Agyenim et al. [47] developed and installed a fully functional solar absorption LiBr-water system with a capacity of 4.5 kW and proved the feasibility of the new concept of cold water storage.
23
C. Sanjuan et al. [50] proposed a different solar absorption cooling system with interior storage. The energy was storage in the form of crystallized salts. This doesn’t use liquid refrigerant. However, authors stated that these systems cannot substitute big absorption pumps. Pablo Bermejo et al. [51] tested a double effect LiBr-water absorption system which was powered by hot water from Fresnel collector. Average daily COP was 1.1-1.25. Ming Qu et al. [52] studied a small solar thermal cooling and heating system at Carnegie Mellon University. The system used 52 m2 of linear parabolic trough solar collectors; a 16 kW P
P
double effect, LiBr-water absorption chiller. TRNSYS results were validated and a sizable storage tank was suggested. K.F. Fong et al. [53] did a comparative study and found that solar electric compression refrigeration and solar absorption refrigeration had the highest energy saving potential in the subtropical Hong Kong. It was found that the parabolic concentrators had the primary energy consumption only slightly better than the flat plate collectors. J.D. Marcos et al. [59] designed an experimental solar energy facility which used a new type of flat plate vacuum solar collectors. The experiment proved that approximately 10 m2 of P
P
solar panels is needed to produce 1 kW of cooling capacity with an air-cooled LiBr-water absorption chiller. F. Palacín et al. [61] and C. Monné et al. [63] validated the TRNSYS results of real solar absorption cooling system. The performance was evaluated with new heat rejection systems. The COP was improved with the use of the geothermal sink. Syed A.M. Said et al. [38] investigated 5 kW NH 3 -water absorption chillers over a period R
R
of 24 h at a refrigeration temperature of −9 °C and concluded the refrigeration system with refrigerant storage is the best-suited design as collector size would be comparatively small compared to other alternatives. S. Du et al. [68] performed an experiment on an air-cooled two-stage NH 3 -WATER 2 kW R
R
absorption system which was driven by 85 °C hot water with an evaporating temperature of 8 °C and ambient air temperature of 29 °C. Its thermal COP was reported as 0.21. Naci Kalkan et al. [25] designed a LiBr-water absorption system with average chiller inlet temperature of 78 °C and thermal COP of 0.58 was reported. Hot water storage of 60-90 L/m2 was P
suggested for evacuated tube collector.
24
P
Pedro J et al. [71] did TRNSYS simulation of a 17.6 kW LiBr-water single-effect absorption chiller with a heat storage tank of 1000 L. The flat plate collector field of 38.4 m2 was P
P
used and a COP of 0.691 was found. Y.L. Yin et al. [77] designed an 8 kW LiBr-water absorption cooling system with a 3 m3 P
P
hot water storage tank. MATLAB program was used to predict the performance. It was found that the solar radiation intensity had a greater impact on the performance than the ambient temperature. Christine Weber et al. [79] found from their operational experience of NH 3 -water R
R
absorption chiller that concentrating solar collectors provide high efficiency at high driving temperatures favorable for thermally driven chillers. Hot water temperature of 100-205 °C and chiller temperature of -12 °C was reported with a COP of 0.8. Such systems with concentrating collectors were suggested for southern countries with high amount of direct irradiation. A minimum size of 175 kW is recommended for economic use as end losses increase significantly as well as the specific collector field cost for smaller units.
Fig 20. NH 3 -water absorption chillers with concentrating collector [79] R
R
2.3 Summary of review A summary has been presented in the Table 1.
25
Table 1. Summary of literature review S. No.
Author(s)
System/Parameter(s)
Result(s)
Theoretical Studies 1.
2.
3.
S.O. Enibe (1997) Discussed SPV based vapour compression
1. Cost is high.
[32]
refrigeration, absorption and adsorption system
2. Performance improvement required.
G.A. Florides et
Presented TRNSYS simulation of a 11 kW LiBr-
1. Optimum system: 15m2 compound parabolic
al. [34]
Water absorption solar cooling system.
Ibrahim Atmaca
A 10.5 kW LiBr-water solar absorption chiller
and Abdulvahap
was dynamically simulated.
P
P
collector and 600 l hot water storage tank. 1. The hot water inlet temperature decreases the 24T
24T
absorber and solution heat exchanger surface
Yigit [36]
area. 2. The storage tank mass should be kept at a minimum and for the study it was 3750 kg. 3. Evacuated tube collectors was the best choice for present study
4.
5.
1. Optimized system: 1 ton chiller with 35m2 of
Assilzadeh, F. et
TRNSYS simulation of a solar LiBr-water
al. [37]
absorption system was done for Malaysia
collector tilted at 20° to horizontal and storage
weather.
tank size of 0.8m3.
P
P
P
D.S. Kim & C.A.
A review of solar electric, solar thermal and new
Solar thermal with single effect absorption
Infante Ferreira
emerging technologies was presented.
system is the best option.
[1]
26
P
6.
7.
M. Mazloumi et
A single effect 17.5 kW LiBr-water absorption
1. Collector mass flow rate has has a significant
al. [39]
system was simulated Horizontal N-S parabolic
effect on optimum capacity of the storage
trough collector was used with an insulated
tank.
Desideri et al. [41]
thermal storage.
2. The minimum collector area was 57.6 m2.
Two case studies were presented for solar cooling
For industries, LiBr-water absorption chiller
in industries and in hotels.
with vacuum tube collectors and a storage tank
P
P
and for hotels, hybrid tri-generation plant was suggested which produces electricity, heating and cooling. 8.
Ursula Eicker and
A full simulation model for absorption cooling
Dirk Pietruschka
systems, combined with a stratified storage tank
[42]
was developed using TRNSYS.
1. Increasing the mass flow decreases the collector surface area and thus the costs. 2. For 3 m2 kW−1 collector area at 85 ° C and P
P
P
P
under low flow conditions, solar fraction of 80% was achieved. 3. Dynamic simulations are necessary for correct solar thermal system sizing. 10.
Tiago Mateus and
A TRNSYS simulation based study was done to
Armando C.
evaluate the potential of integrated solar
Oliveira [44]
absorption cooling and heating systems for building applications.
1. The cost and CO 2 can be saved by employing R
R
integrated systems. 2. Single-family house and the hotel showed the economic feasibility for the system. 3. Cost of chiller and the solar thermal systems need to go down.
27
11.
J.M. Abdulateef et al. [47]
Authors performed the steady-state analysis of the ejector refrigeration system.
1. Single-effect Solat absorption system appeared to be the best option and solar ejector refrigeration system were not economically competitive
12.
Boonrit
Authors studied the simulation by mathematical
Results indicated that solar–biomass hybrid air
Prasartkaew and
modelling of a solar–biomass based single effect
conditioning for tropical locations for residential
S. Kumar [49]
LiBr–water absorption chiller.
applications is feasible.
The overall system performance on a daily and monthly basis was evaluated. 13.
Berhane H.
Authors worked for reducing the life cycle impact
Gebreslassie et al. of cooling.
Environmental impacts can be largely reduced by investing on the solar collectors.
[54] 14.
15.
T. Tsoutsos et al.
TRNSYS simulation of a LiBr-water absorption
Investment cost was high and payback period
[55]
chiller with storage tank was done.
was 11.5 years without subsidy.
Marwan Mokhtar
A techno-economic analysis of the solar collector
1. Large-scale cooling plants were the most
et al. [56]
and chiller system was done by considering 25 feasible combinations of solar collection and
P. Lin et al. [64]
2. Selection of heat rejection system and storage tank size were found to be very important.
cooling systems. 18.
economical.
A numerical simulation of a 5 kW two-stage air-
A thermal COP of 0.34 and electrical COP of 26
cooled NH 3 -water solar absorption chiller was
in summer condition was fond with 85 °C hot
done.
water supplied from solar collector.
R
R
28
19.
X.Q. Zhai et al. [65]
Authors presented some new design options with
1. Half-effect and two-stage chillers seemed
regard to solar collectors, auxiliary energy
more suitable for air-cooled solar absorption
systems and cooling modes after reviewing solar
cooling systems in hot and dry regions.
single-effect absorption cooling systems.
2. Solar collection and chiller performance are very influential parameters for cost. 3. Ground source heat pumps may be integrated as auxiliary system. 4. Ground cooling can eliminate the need for cooling towers. 5. The specific tank volume was 0.01–0.08 m3/m2 . P
20.
P
P
P
Emilio J. Sarabia
Single effect solar absorption cooling system was
The increase in the evaporator temperature had a
Escriva et al. [66]
studied analytically and the effect of generator
negative effect
temperature was analyzed. 21.
V. Boopathi Raja
Study to minimize the cost of solar LiBr-water
and V.
absorption cooling system was done.
Shanmugam [27]
1. Single effect LiBr–water absorption cooling system was more suitable for domestic purpose. 2. Flat plate and evacuated tube solar collectors were more economical. 3. Optimum generator temperatures were around 75 to 92 °C and evaporation temperatures about 5–10 °C.
29
23.
Berhane H.
Mathematical modelling was done to optimize the The technology is not economically viable as per
Gebreslassie et al. economic and environmental performance of
25.
[72]
solar absorption cooling system.
F.J. Cabrera et al.
Parabolic trough collector (PTC) based air
[74]
conditioning systems were reviewed and the performance was compared with other types of solar collectors.
the current market.
1. Flat-plate collectors (FPC) had the lowest solar fraction and the PTC had the maximum. 2. By including hot water storage tank in the system the auxiliary heating can be also lowered.
26.
Renato M.
Author compared the various solar cooling
Lazzarin [75]
techniques in terms of the overall efficiency and the investment cost.
1. The specific cost of PV was found higher than solar thermal. 2. The investment costs were very similar for PV solar cooling systems and FPC, ETC and PTC driven absorption for a daily production of 10 kWh cooling.
27.
Ali M.
Authors compared the performance solar
Desiccant cooling system achieved a higher
absorption and solar desiccant cooling system.
solar fraction than absorption cooling system.
Yin Hang et al.
Solar absorption integrated systems were studied
The methodology can be applied to other
[78]
which could provide cooling, heating and hot
renewable energy systems
Baniyounes et al. [76] 29.
water. Optimization study was done on TRNSYS simulation program.
30
Experimental 1.
Ibrahim Dincer et
A water heating and cooling system based on
al. [31]
absorption cooling using a mixture of R22 and
found to be 0.6 and 0.45 respectively with a
Dimethyl Ether Tetra Ethylene Glycol as the
generator temperature of 90 degrees.
working fluid was constructed using flat plate
2. R22 has detrimental effect on environment. It should be replaced.
collectors. 2.
1. Theoretical and experimental COP were
R. Best & N.
A review on solar cooling technology was
Ortega [33]
presented and a solar cooling installation which included flat plate collector of 316 m2 , LiBrP
P
1. Performance was highly dependent on environmental conditions. 2. The cost of the system should be affordable.
water absorption chiller and a heat storage tank of 30m3 was studied. P
3.
4.
P
K Sumathy et al.
A solar based low temperature (75-60 °C) 2-stage
The proposed system could be integrated with
[35]
LiBr-water absorption cooling system was
existing solar hot water system hence is efficient
developed.
and cost effective
Zambrano et al.
A model of a solar cooling system was developed
Experiment results were in agreement with the
[38]
for demonstration purpose. The system included a
theoretical results.
35 kW single-effect absorption chiller with flat plate collector. 5.
Ahmed Hamza H.
A performance study was done for a 35.17 kW
Ali et al. [40]
solar single-effect LiBr-water absorption chiller in operation for the duration of five years.
31
The specific collector area was 4.23 (m2/kW cold ). P
P
R
R
Plant included vacuum tubes collectors, a hot water storage tank of 6.8 m3, a cold water storage P
P
capacity of 1.5 m3 and a 134 kW cooling tower. P
7.
P
Francis Agyenim
A solar cooling system consisting of a 12 m2
1. The concept of cold storage was well proved.
et al. [48]
vacuum tube solar collector, a 4.5 kW LiBr/H2O
2. Electrical COP and thermal COP were 3.64
P
P
absorption chiller, a 1000 l cold storage tank and a 6 kW fan coil was studied.
and 0.58 respectively. 3. 180 – 250 Liters of storage capacity was required per kW of cooling produced. 4. Chilled water temperature was 7 °C.
9.
10.
Pablo Bermejo et
A 174 kW double-effect LiBr-water absorption
al. [51]
chiller with 352 m2 solar field of a linear P
P
1. The daily average solar collector efficiency was found to be 0.35.
concentrating Fresnel collector was tested.
2. Daily average COP of 1.1–1.25 was reported.
Ming Qu et al.
A 16 kW double effect, LiBr-water absorption
1. The orientation of the collectors impacted the
[52]
solar thermal cooling and heating system with 52m2 of linear parabolic trough solar collectors P
P
was developed and studied. Models were developed in TRNSYS.
performance significantly. 2. The storage tank should be placed in solar collection loop. 3. A drain-back tank should be considered to achieve higher system performance.
11.
K.F. Fong et al.
A comparative study was done for a particular
When solar electric and solar absorption systems
[53]
building in Hong Kong. The systems included: (i)
were considered the yearly energy savings
solar electric compression, (ii) solar mechanical 32
compression, (iii) solar absorption (iv) solar
would be 15.6% to 48.3% compared to the
adsorption and (v) solar solid desiccant cooling.
conventional cooling.
Simulation models were developed and performances were evaluated. 12.
J.D. Marcos et al.
An experimental air cooled 4.5-kW LiBr/water
[59]
solar absorption cooling and heating facility with a net flat plate vacuum collector area of 42 m2 P
P
was developed and studied in Madrid. System
14.
conditioning demand was met. 2. Nearly 10 m2 of collectors was needed for 1 P
P
kW of cooling capacity.
also included a 1500-l hot storage tank.
3. The mean COP of absorption chiller was 0.55.
F. Palacín et al.
A model of 4.5 single effect, air cooled, LiBr-
1. When used dry cooling tower, the COP was
[62]
water absorption chiller was developed and simulated in TRNSYS program. Two heat
15.
1. 65.3% of heating demand and 46% of the air
highly influenced by the outdoor temperature. 2. Geothermal sink improved the performance
rejection sinks were analyzed with the system;
and performance was independent of outdoor
dry cooling tower and geothermal sink.
conditions.
C. Monné et al.
Developed a dynamic model of a 4.5 kW single
[63]
effect, LiBr–water rotary absorption chiller and a dry cooler tower using TRNSYS.
1. Cooling water temperature influenced the COP. 2. Geothermal sink improved the COP up to 42%.
16.
Syed A.M. Said
A detailed study on the alternate designs of solar
et al. [68]
absorption cooling technology was done. Authors investigated a 5 kW ammonia-water absorption
1. The refrigeration system with refrigerant storage was the best-suited design. 2. Refrigerant storage systems had the high COP.
33
chillers over a period of 24 h at a evaporator
3. The cost of the refrigerant-storage system was
temperature of −9 °C. 17.
S. Du et al. [69]
lesser as compared to other alternatives.
An experimental study on an air-cooled 2 kW
When the prototype was driven by 85 °C hot
two-stage NH 3 -water absorption refrigeration
water with an evaporating temperature of 8 °C
system was done.
and ambient air temperature of 29 °C, its thermal
R
R
COP and electric efficacy (ε) reached 0.21 and 5.1 respectively. 18.
Naci Kalkan et al.
A 4.5 kW solar LiBr-water single effect
1. Average chiller inlet temperature was 78 °C
[25]
absorption air-conditioning system was designed
and total electrical and thermal COP of 3.64
with improvements. The system included 15 m2
and 0.58 was achieved.
P
P
of evacuated tube collectors, a 1000 L cold water
2. 60 to 90 L of storage may be required for each m2 of evacuated tube collector.
tank and a 6 kW fan coil unit.
P
P
3. The system was uneconomic. 19.
20.
François
A 5 kW solar NH 3 -water absorption chiller was
The solution heat exchanger and the absorber
Boudéhenn et al.
developed and the numerical and experimental
were found as the optimization point.
[70]
analyses were presented.
Pedro J et al. [71]
A17.6 kW LiBr-water single-effect absorption
R
R
1. An average of 29% of the solar energy
chiller was studied and simulated in TRNSYS.
incident on the solar collectors’ surface was
The system consisted of 38.4 m2 flat-plate
transferred to the hot water storage.
P
P
collector.
2. Average COP of the absorption chiller was 0.691.
34
21.
Y.L. Yin et al.
A 8 kW solar LiBr-water absorption cooling
[77]
system with 96 m2 of collector area was P
1. Solar radiation intensity has a greater impact on the performance than the ambient
P
experimented. A water storage tank of 3 m3 was P
P
used to store the hot water. Theoretical modelling
temperature. 2. Indoor temperature decreases with the increase of the solar radiation intensity as well
was done on MATLAB.
as the decrease of the ambient temperature. 3. The systems were not competitive 22.
Christine Weber
Authors studied the operation of NH 3 -water
et al. [79]
absorption chiller with concentrating solar
chiller temperature of -12 °C was reported
collectors.
with a COP of 0.8.
R
R
1. Hot water temperature of 100-205 °C and
2. Such systems with concentrating collectors were suggested for southern countries with high amount of direct irradiation. 3. A minimum size of 175 kW is recommended for economic use
35
2.4 CONCLUSION The literature study shows that the absorption chillers have been the most popular and preferred technology in terms of solar thermal based cooling systems. Mostly, LiBr-water and the Ammonia-water are used as refrigerant solutions or working pairs. Among the two pairs, the LiBrwater can be easily employed in the cooling systems for human thermal comfort. The chillers with NH 3 -water as the working fluid can produce refrigerating effect of sub-zero temperatures but the R
R
temperature requirement at the generator is in the range of 95-120 C for water cooled chillers which will necessarily require concentrating collectors. The temperature requirements in LiBrwater chillers can be as low as 70 C. The temperature can be achieved by flat plate collectors very much suited for the LiBr-water chillers, thus saving the cost. The water cooled single effect LiBrwater solar absorption chiller seemed best suited for small cooling requirements and inclusion of hot water tank improved the performance. The chillers can serve as the heating system as well. In winter, the hot water from collector can be used directly or the air can be heated by using the fan coil.
2.5 OBJECTIVE Present work intends to design a LiBr-water based solar thermal powered absorption cooling and heating system. The theoretical design will be carried out for Students’ Computer Lab at AHEC, IIT Roorkee, India. The cooling and heating loads will be calculated for the design day and based on the cooling load, the absorption cycle will be simulated in spreadsheet and the various parameters will be calculated and analyzed. The aim of the effort is to find out the thermal energy requirement and the flat plate solar collector field area. The hot water storage tank is being included in the solar thermal loop and the volume of the tank will be calculated.
36
CHAPTER 3
SOLAR ABSORPTION COOLING AND HEATING SYSYTEM
3.1 System Description A solar absorption cooling and heating system being designed in present work consists of a LiBr-water absorption chiller which is hot water fired coming from solar thermal system. In heating mode, the chiller remains idle and the hot water is directly fed to the fan coil unit to warm the space. The complete system can be explained as follows:
Figure 21. Solar absorption cooling and heating system layout 37
3.1.1 Working of the system Cooling mode In the cooling mode, the chiller is hot water fired to produce cooling effect. The incident solar radiation is intercepted by the solar collector where the solar heat increases the temperature of incoming water. The hot water then flows into and is stored in the hot water storage tank. The hot water from the storage tank is fed to the LiBr-Water absorption chiller as shown in the figure where it loses its heat in the generator part of the chiller and comes out with low temperature which goes back to the hot water storage tank. The cooling effect is produced in the evaporator part of the chiller where the water is chilled. The chilled water coming out from chiller goes to fan coil unit where it cools the air stream going into the space to be conditioned. The cooling water continuously flows from the cooling tower to extract heat at absorber and the condenser part as explained in the section to follow. Heating mode In heating mode the chiller is left idle as there is no cooling required. The hot water from the hot water storage tank is made to flow directly to fan coil unit where it heats the air stream to be sent to the air-conditioned space. The operation during heating hours is very simplified as the absorption chiller and cooling tower do not function.
3.1.2 Components of the system The main components of the system are: 1. Solar collector 2. Hot water storage tank 3. LiBr-water absorption chiller 4. Fan coil unit 5. Cooling tower 3.1.2.1 Solar collector A solar collector absorbs the incoming solar radiation and converts the same into thermal energy. The thermal energy collected is then transferred by a transport medium which can be any 38
of heat carrying fluid like air, water and etc. The heated fluid from the collector can be directly used or can be stored in any hot water storage tanks. Any typical solar collector comprises of a glazing surface, absorber plate, headers, insulation and casing. There are two types of solar collectors. The two types are (i) stationary and (ii) concentrating. In a stationary and non-concentrating solar collector, the solar radiation is intercepted and absorbed with the same areas of interception and collection while in case of concentrating collectors, the radiation is intercepted on a larger area than the collection area. As the name suggests, the radiation is collected at a concentrated area. The ratio of the area of the interception to that of collection is known as the concentration ratio. 3.1.2.1.1 Stationary Collectors There are basically three types of stationary collectors: 1. Flat plate (FPC); 2. Stationary compound parabolic (CPC) and 3. Evacuated tube (ETC). Flat plate collectors (FPC) A flat-plate solar collector is shown in Fig. 22 and Fig. 23. The solar radiation passes through a transparent cover known as glazing and falls on the absorber surface. During the absorption, a large share of falling energy is absorbed by the plate and transferred to the heat carrying fluid in the tubes which can be then carried away for storage or direct use. The absorber plate and the sides of casing are properly insulated to reduce various thermal losses to the environment. The glazing at the top is employed to minimize the convection losses from the absorber plate. The transparent cover, glazing also reduces radiation losses from the collector as the glazing (glass) does not allow the long-wave thermal radiation to pass through which are emitted by the absorber plate. This is known as greenhouse effect. FPC are generally stationary and don’t require any kind of tracking of the sun. The collectors is oriented towards the equator. This is why it faces south in the northern hemisphere and faces north in the southern hemisphere.
39
The tilt angle of the collector is kept equal to the local latitude with addition or subtraction of 10 – 15⁰.
Figure 22. A Flat plate collector [80]
Figure 23. Exploded view [80]
The various FPC components can be briefly discussed as below: Glazing: This is the layer(s) glass or other transparent material which can transmit radiations. Absorber plates: These are flat plates which are corrugated, or grooved and where tubes or the passages are attached where the fluid flows. Headers or manifolds: To admit the fluid in the collector and to discharge it. Insulation: It minimizes the thermal losses to the environment from the back and sides of the collector. Container or casing: It encloses all the components and keeps them free from dust, moisture, etc. The FPC are wide used because they can collect the solar energy at the lower cost.
Stationary compound parabolic collectors (CPC) CPC can reflect all the incident radiation within wide limits to the absorber. the two sections of the parabola are joined together facing each other to avoid the requirement of tracking the sun as shown in Fig. 24. 40
Compound parabolic concentrators have the capability of accepting incoming radiation over a wide range of incident angles. It uses multiple internal reflections by which, any radiation that enters the aperture within the collector acceptance angle, finally reaches to the absorber surface at the bottom of the collector. The absorber has many configurations. It can be cylindrical as shown in Fig. 24 or it can be flat. In the CPC shown in Fig. 24, the lower portion of the reflector (AB and AC) is kept circular, while the upper portions (BD and CE) are parabolic. T in the Fig.24 is the tube where the heat is collected in a heat carrying fluid.
Figure 24. Schematic of a compound parabolic collector [80] Evacuated tube collectors The FPC are generally well suited for areas with good amount of solar radiation. Their benefits are heavily decreased when weather conditions are hostile during cold, cloudy and windy days. The evacuated tube collectors are efficient options in such conditions. Evacuated solar collectors operate differently. The evacuated tube collectors consist of a heat pipe which is inside a vacuum tube, as shown in Fig. 25. The vacuum cover reduces the losses by conduction and convection and hence the collectors can achieve higher temperatures than FPC. Evacuated tube collectors can collect both direct and diffuse radiation as a FPC. The efficiency of EPC is higher at low incidence angles and this adds an advantage over FPC and there is an improvement in day-long performance. ETC use phase change materials to transfer heat. The material gets vaporized absorbing the heat. Evacuated collectors have a heat pipe which is a great conductor, placed inside a vacuum tube. 41
Figure 25. An evacuated tube collector [80] The pipe, which is a sealed copper pipe, is then attached to a black copper fin that fills the tube. Extending from the top of each tube there is a metal tip which attached to the sealed pipe which acts as condense). Each heat pipe contains a small amount of fluid e.g. methanol that evaporates and condenses. Solar heat evaporates the liquid and when the vapour reaches the heat sink region to condense and release its latent heat. The condensed fluid then returns back to the collector and this process repeats. Fig. 5 shows an evacuated tube collector. Water, or glycol, can be used as fluid to flow through the manifold to pick up the thermal energy from the tubes. The heated liquid can then be circulated through any heat exchanger to gives off its heat to a process or can be directly sent to the storage tank. Sun tracking concentrating collector The higher temperature of the fluid can be achieved if the solar radiation can be collected on a small area. In all the concentrating collectors solar radiation is concentrated and then is transferred into heat. Concentration is achieved by using reflection or refraction of radiation by using the mirrors or lens. After the reflection or refraction, the radiation is concentrated in a point or line which is the focal zone, thus increasing the energy intensity in the receiving surface increasing the temperature of the fluid.
42
Fresnel type PTC
Parabolic trough collector
Parabolic dish collector
Central tower system
Figure 26. Sun tracking concentrating collectors [80] Table 2. Solar energy collectors’ summary [80]
43
3.1.2.2 Hot water storage tank A hot water storage tank is simply a water tank which is employed for hot water storage. This can be done for many applications such as domestic use, space heating, or cooling. A properly insulated thermal tank can retain the thermal energy for several days. Hot water tanks may be supplied with an auxiliary heating source such as gas burner or oil burner, electric heaters etc. By integrating hot water storage tank with solar thermal systems the efficiency of the system can be improved and an intermittent systems can be developed which can be quite relevant with cooling and heating system. Purposes of a hot water tank can simply be summarized with two following points: 1. To avail the continuous run of hot water to the cooling system or FCU. 2. To avail the hot water in the times of non- sunshine hours. The obvious ideal condition for storage tanks is to be insulated properly to arrest the different heat losses to the surroundings. Very popular storage tank for the solar thermal systems are stratified hot water storage tank. Stratified hot water storage tank is preferred in solar thermal system because of their special capability of maintaining temperature difference between the different layers of water in the tank. The mixing of the tank water is prevented in the stratified storage tank thus there is a considerable difference in the temperature of water at the upper (hot) end of the tank and lower (cold) end of tank.
3.1.2.3
Absorption chiller
Absorption chiller is a very crucial component of the system. This is an alternative to the conventional cooling/refrigeration system which is vapor compression refrigeration. In vapor compression system huge amount of compression work is required which is given through electrical supply. In case of absorption chiller, the cooling effect is produced by inputting thermal energy even in the form of low temperature heat source. The complete absorption system can be explained as in figure 27.
44
Absorption machines are available in the market for two basic purposes. For applications in air conditioning which need cooling above 0⁰C and for applications requiring cooling below 0⁰C. In application where above 0 C cooling is needed, LiBr-water based machines are preferred and NH 3 -water for the other. In present work, the system is designed for air conditioning purpose hence the selected refrigerant absorbent pair is water-lithium bromide. Figure 27 shows a a typical LiBr-water machine. The complete process takes place in two vessels. The top vessel houses the generator and condenser and the lower vessel contains the absorber and evaporator. The heat input is given at the generator.
Figure 27. An absorption cooling machine This input heat evaporates the refrigerant, water in LiBr-water absorption machine, at the generator. The water vapor then passes into the condenser section where a cooling medium 45
condenses the water vapor to the liquid state. The liquid water then goes down to the evaporator where absorbs heat from the fluid to be cooled. Due to a very low pressure in the absorberevaporator section, boiling point of water decreases thus it boils off at a very low temperature. Evaporated water is then made to pass into the absorber where it gets mixed with a strong LiBrH2O solution. This strong solution absorbs the water vapor coming from the evaporator to make a weak solution. The solution is sent to the generator section by pumping. This cycle repeats to produce cooling effect at the evaporator In an absorption chiller, three fluid loops are there that have external connections: (i) generator heat input loop, (ii) cooling water loop, and (iii) chilled water loop.
3.1.2.4 Fan coil unit A fan coil unit is a device which consists of a cooling or heating coil and a fan. It is an integral part of a heating and cooling system employed in commercial, residential, and industrial buildings or applications. Fan coil unit is technically air to water heat exchanger. The fan coil receives hot or cold water, and removes heat from or adds heat to the air through heat transfer.
Figure 28. A fan coil unit [81]
46
Fan coil units can use a medium of chilled water to provide cooling to the conditioned space. Fan
coil units are capable of utilizing lower grade chilled water. Similar to the cooling, the fan coil unit can be used for heating purposes where it heats the incoming air stream by taking heat from the hot water supply. In general, they are supplied with low temperature hot water for heating purposes.
3.1.2.5 Cooling Tower A cooling tower is used to cool the hot water by losing its heat to the environment. The heat energy absorbed by the chiller needs to be rejected out of the system and into the atmosphere. Evaporative heat rejection devices called cooling towers are typically used to lower the water temperature in large chiller systems. The system in this study is considered to be water cooled chiller. Hence, a cooling tower is required to lose waste heat while the chiller is in operation to produce cooling effect.
3.2 CLOSURE The proposed system has been described in the present chapter. A typical layout of the system has been drawn and the working of the system in the cooling and heating mode has been explained. The system comprises of various components and each component has been illustrated in the chapter. The chapter also discusses the various major components’ characteristics and operation. The following chapter deals with the site details, meteorological information for the site and solar data calculation leading to the load calculations.
47
CHAPTER 4
METEOROLOGICAL DATA, SOLAR DATA AND LOAD CALCULATION
Solar thermal cooling and heating machine has various components as discussed in the Chapter 3. For designing the system, the various calculations are involved. The complete process for the design can be summarized as follows:
1
• Identify the site
2
• Get the complete details of site (Location, Dimensions, Construction details)
3
• Calculate cooling and heating load (Occupancy, solar gain, internal gain, )
4 5
• Design of absorption chiller • Solar collector area, storage tank volume and flow rate
Figure 29. Flow chart for the design process
4.1 Site The chosen site for the project undertaken is the Student Computer Laboratory of Alternate Hydro Energy Center (AHEC), Indian Institute of Technology (IIT), Roorkee, India. The computer lab is located on the second floor of the AHEC building which is open for the students all day in office hours, generally from morning 8 AM to evening 5 PM.
48
Figure 30. AHEC, IIT Roorkee
Figure 31. Bird-eye view of AHEC, Roorkee The computer lab has two adjacent rooms on its two sides. The windows are provided on the one side and one side faces the corridor. The exposed surface to solar radiation is the wall with the
49
fenestration which is around 15 ⁰ left to the north. The plan of the computer lab is shown in the figure 32.
Figure 32. Plan of the Computer Lab, AHEC, IIT Roorkee
Figure 33. Panoramic view of Computer Lab, AHEC, IIT Roorkee
4.2 Site details Latitude: 29.87 ⁰N Longitude: 77.9 ⁰E Floor area: 44 m2 50
Height: 3.1 m Total fenestration area: 12.75 m2 Window glass thickness: 0.3175 cm (1/8 in) (single glazed) Wall thickness (brickwork): ~23 cm (9 Inches) Total wall area exposed to surroundings: ~23 m2 (including fenestration) Total side walls area: ~36.6m2 Wall area towards the corridor: ~23 m2 Roof (concrete): ~15 cm (6 Inches)
4.3 Load calculation Cooling and heating load calculations require collection and processing of various meteorological data.
4.3.1 Meteorological data for Roorkee, India Roorkee is a city in the state of Uttarakhand, India which spreads over a plane landscape with the grand spectacle of Himalayas ranges adjoining it in the East and the North-east. Roorkee is located at 29.8749° N (latitude), 77.8899 (longitude) ° E. It is on the banks of the Ganges canal. The popular attraction of the city is the Upper Ganges Canal which flows south and bisects the Roorkee city. Table 3. Meteorological Data of Roorkee, India Latitude Longitude Elevation Coldest Month Hottest Month Average temperature in coldest month
29.87° N 77.89° E 274 m January June 12.6° C
Average temperature in hottest month
31.3° C
51
The average temperature at the Roorkee for the design months were calculated using the software Meteonorm 7. The monthly temperature plot is shown in the Fig. 34.
Figure 34. Monthly temperature at the Roorkee by Meteonorm 7 Using the meteorological data and the ASHRAE Handbook – Fundamentals [82] the climate design conditions were developed. Chapter – 14 of ASHRAE Handbook – Fundamental [82] provides the climate design information for the location Dehradun which was used to generate the climate design conditions for Roorkee since Roorkee can be approximated to have same weather conditions as that of Dehradun which is just at a distance of ~70 kms. This approximation is very much valid for generating climate design information for calculating cooling load and heating load calculations. Firstly the solar radiation available over Roorkee was calculated using the spreadsheet with the mathematical relations available in ASHRAE Handbook – Fundamental [82]. 52
4.3.2 Solar Radiation Calculation over Roorkee, India It is mandatory to know the clear-sky solar radiation on the design days for several calculations that require the heats gains to calculate the cooling and heating load. Accordingly, hourly clear sky beam and diffuse solar radiation were calculated using various expressions explained in following section. The obtained values of solar irradiance were transposed on the receiving surface according to their orientation. Solar Constant: According to ASHRAE, the solar constant Esc can be defined as the intensity of solar radiation on a surface n that is normal to the solar beam, just outside the earth’s atmosphere, at the average earth-sun distance. The widely accepted and used value that was proposed by the World Meteorological Organization in 1981, is Esc= 1367 W/m2. Since the earth revolves the sun in an elliptical orbit, the extraterrestrial radiant flux Eo changes throughout the year. It reaches a maximum of 1412 W/m2 in beginning of January, when earth is nearest to the sun, also known as the aphelion. The minimum value of extraterrestrial radiant flux is 1322 W/m2 in the starting of July, the day when the earth is farthest from the sun known as perihelion. Extraterrestrial solar irradiance incident on a surface normal to the sun’s ray can be approximated with the following equation:
(n − 3) = E Esc 1 + 0.033cos 360 365
(4)
Where n is the day of year (1 for January 1, 32 for February 1, etc.) and the argument inside the cosine is in degrees. Equation of Time (ET) and Solar Time: For the reason that the earth’s orbital velocity also changes throughout the year, the apparent solar time (AST) which is determined by a solar time sundial, changes a bit from the average time kept by a clock running at a uniform rate. This difference is called the equation of time and is calculated by the following formula:
0.0075 + 0.1868cos ( Γ ) − 3.2077 sin ( Γ ) − ET = 2.2918 1.4615cos ( 2Γ ) − 4.089sin ( 2Γ ) With ET expressed in minutes and 53
(5)
Γ =360
n −1 365
(6)
Once the equation of time is calculated, the standard solar time is calculated by adding the equation of time and the longitude correction as following:
AST =LST + ET / 60 + (LON − LSM) /15
(7)
Where LON is the longitude of the location and LSM is the local standard meridian for that time zone which is calculated as:
LSM = 15TZ
(8)
Where TZ is the time zone, expressed in hours ahead or behind coordinate universal time (UTC). Declination ( δ ): Solar declination is defined as the angle between the earth-sun line and the equatorial plane. Since the equatorial plane is tilted at an angle of 23.45⁰ to the orbital plane, the solar declination ( δ ) changes throughout the year as shown in the Fig. 82.
Figure 35. Motion of earth around sun [82] The solar declination angle ( δ ) is calculated for each day as follows:
δ = 23.45sin 360
n + 284 365
(9) 54
Where δ is in degrees and the argument in the sine is also in degrees. Sun Position: The sun position in the sky with respect to the location is easily expressed in terms of the solar altitude angle above the horizontal and the solar azimuth angle measured from the south as shown in the figure.
Figure 36. Various Solar angles for a surface [82] The altitude angle (β) is the angle between the horizontal plane and a line coming from the sun. The altitude angle ranges from 0° (sun is on the horizon) to 90° (sun directly overhead). Negative values of the altitude angle correspond to night times. The solar azimuth angle (φ) is explained as angle between the south of the projection on the horizontal plane and the earth to sun line. Conventionally, it is taken positive for afternoon hours and negative for morning hours. Solar altitude angle and the azimuth angle depend on various factors. These are local latitude L, the declination angle δ which varies each day and is function of date, and the hour angle H. The hour angle, H is the angular displacement of the sun east or west of the local meridian due to the rotation of the earth, and it is calculated in degrees by using the expression:
= H 15 ( AST − 12 )
(10)
55
Where AST is the apparent solar time. H is zero at solar noon, positive for the times in the afternoon and negative for the morning hours. Now the solar altitude angle can be calculates from the equation:
= sin β cos L cos δ cos H + sin L sin δ
(11)
The azimuth angle is determined by its sine and cosine as follows:
sin φ = sin H cos δ / cos β
= cos φ
(12)
( cos H cos δ sin L − sin δ cos L ) / cos β
(13)
Air Mass: The air mass m is defined as the ratio of the mass of the atmosphere through which the direct radiation passes to the mass it would pass if the sun was just overhead, the zenith. Air mass is a function of solar altitude angle β and is calculated from equation:
m= 1/ sin β + 0.50572 ( 6.07995 + β )
−1.6364
(14)
Clear sky solar radiation: The clear sky solar radiation is defined by its beam or direct and diffuse components. The direct or the beam component represents the solar radiation coming directly from the sun and the diffuse components comprises of all the radiations emanating from the rest of the sky. The both components can be obtained by using the formulae: = Eb E exp −τ b m ab
(15)
= Ed E exp −τ d m ad
(16)
Where
Eb = beam normal irradiance which is measured perpendicularly to rays of the sun Ed = diffuse horizontal irradiance which is measured on horizontal surface
E = extraterrestrial normal irradiance m = air mass 56
τ b and τ d = beam and diffuse optical depth ab and ad = beam and diffuse air mass exponents values of τ b and τ d are location specific and vary during the year. The air mass exponents ab and ad are calculated as follows:
ab = 1.219 − 0.043τ b − 0.151τ d − 0.204τ bτ d
(17)
ab = 0.202 − 0.852τ b − 0.007τ d − 0.357τ bτ d
(18)
Once the beam normal irradiance Eb and diffuse horizontal irradiance Ed have been calculated, the values can be transposed to receiving surfaces of various orientations. Transposition to receiving surfaces of various orientations: In present study, the cooling and heating load is to be calculated for a Computer Lab, which has vertical and horizontal exposed surfaces which receive solar gain. That’s why calculation of clear-sky solar radiation on exposed surfaces of arbitrary orientations is required. The solar radiation receiving surfaces can be vertical or tilted. The following section describes the method that can enable calculating solar irradiance incident on any surface by knowing the direct normal or beam and diffuse horizontal irradiance. To calculate the irradiance at any surface, some solar angles corresponding to the receiving surfaces must be known. The orientation of any surface is best described by its tilt angle and its azimuth angle. The tilt angle Σ is the slope of the surface, the anle between its plane and the horizontal. Tilt lies between 0 and 180° but generally, slope is between 0° for horizontal and 90° for vertical surface. The surface azimuth angle ψ is the angle between the south of the projection on the horizontal plane and the normal to the surface. The surfaces that faces west direction has a positive surface azimuth angle and those that facing east has a negative surface azimuth. Table 4 summarizes the solar azimuths for common orientations. Table 4. Surface Orientations and azimuths, measured from south Orientation Solar azimuth ψ
N 180⁰
NE -135⁰
E -90⁰ 57
SE -45⁰
S 0⁰
SW 45⁰
W 90⁰
NW 135⁰
The surface-solar azimuth angle γ is the angle between the solar azimuth angle φ and the surface azimuth angle ψ:
γ = φ −ψ
(19)
Values of γ more than 90⁰ or smaller than -90⁰ signifies that the surface is in shade. Finally, the angle of incidence θ is defined as the angle between the line normal to the irradiated surface and the earth-sun line. The incidence angle is very important in load calculations, and solar technology for the reason it decides the magnitude of the direct component of solar radiation falling on the surface and the surface’s ability to absorb, transmit, or reflect the sun’s rays. It is given by
= cos θ cos β cos γ sin ∑ + sin β cos ∑
(20)
For vertical surfaces ( ∑ =90 ) and
cos θ = cos β cos γ
(21)
Whereas for horizontal surfaces ( ∑ =0 ) and
θ= 90 − β
(22)
Now, Total clear-sky irradiance Et on the receiving surface is calculated by summing the three components: the direct orbeam component Et,b; the diffuse component Et,d,; and the groundreflected component Et,r emanating from the ground facing the receiving surface. Thus, Et = Et ,b + Et ,d + Et ,r
(23)
Beam Component, Et,b is obtained from
Et ,b = 0
(24)
This relation is valid when cos θ > 0 ; otherwise Et ,b = 0 .
58
Diffuse component, Et,d for vertical surfaces is calculated from = Et ,d Ed Y sin ∑
(25)
Where,
(
Y max 0.45, 0.55 + 0.437 cos θ + 0.313cos 2 θ =
)
(26)
For a non-vertical surface with slope ∑ , the following relationships are used.
= Et ,d Ed (Y sin ∑ + cos ∑ ) = Et ,d Ed Y sin ∑
if ∑ ≤ 90
(27)
if ∑ > 90
(28)
Ground-Reflected component, Et,r for surfaces of all orientations is given by = E t ,r
( Eb sin β + Ed ) ρ g
1 − cos ∑ 2
(29)
Where ρ g is the ground reflectance which is generally taken as 0.2. The solar radiation available at Roorkee has been calculated using the ASHRAE handbook of fundamentals (2013). Subsequently the incident radiation was calculated for each hour on the exposed wall of the Computer Lab, AHEC, IIT Roorkee i.e. the wall with the windows/fenestration and the roof top. The calculation was done using spreadsheet programming and tabulated in the following section. The important results of the radiation availability on design days have been plotted on the different charts. As mentioned in the previous section, hottest month at Roorkee is June and coldest month is January, the radiation data have been calculated for 21st date of the months. The June 21 is chosen to design day for calculation of cooling load and January 21 has been taken as design day for heating load calculations. All angles are in degrees and radiations in Watts in the tables.
59
Table 5. Solar radiation calculation data table over Student computer Lab, AHEC, IIT Roorkee (on June 21)
Declination Time LST (δ) (Hrs) (⁰) 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
23.45 23.45 23.45 23.45 23.45 23.45 23.45 23.45 23.45 23.45 23.45 23.45 23.45 23.45 23.45 23.45 23.45 23.45 23.45 23.45 23.45 23.45 23.45 23.45
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
Equation Of Time (ET) (Hours)
AST= LST+ET/60+ (LONLSM)/15
-1.3246 -1.3246 -1.3246 -1.3246 -1.3246 -1.3246 -1.3246 -1.3246 -1.3246 -1.3246 -1.3246 -1.3246 -1.3246 -1.3246 -1.3246 -1.3246 -1.3246 -1.3246 -1.3246 -1.3246 -1.3246 -1.3246 -1.3246 -1.3246
0.67099 1.67099 2.67099 3.67099 4.67099 5.67099 6.67099 7.67099 8.67099 9.67099 10.67099 11.67099 12.67099 13.67099 14.67099 15.67099 16.67099 17.67099 18.67099 19.67099 20.67099 21.67099 22.67099 23.67099
Hour Angle (H) (⁰) -169.935 -154.935 -139.935 -124.935 -109.935 -94.9352 -79.9352 -64.9352 -49.9352 -34.9352 -19.9352 -4.93515 10.06485 25.06485 40.06485 55.06485 70.06485 85.06485 100.0649 115.0649 130.0649 145.0649 160.0649 175.0649
Sin β
Solar Altitude Angle (β)
-0.58512 -0.52244 -0.41066 -0.25738 -0.07306 0.129747 0.337217 0.535214 0.710242 0.850374 0.946061 0.990782 0.981488 0.918814 0.80703 0.653754 0.469432 0.266625 0.059154 -0.13884 -0.31387 -0.454 -0.54969 -0.59441
-35.8112 -31.4962 -24.2462 -14.9148 -4.18978 7.454947 19.70744 32.3584 45.25459 58.2524 71.09572 82.21426 78.95826 66.75324 53.80672 40.82523 27.99743 15.46354 3.391267 -7.98084 -18.2926 -27.0008 -33.3457 -36.4706 60
sin(φ)
Solar Azimuth angle (φ)
Air mass (m)
-0.1977 -0.45581 -0.64762 -0.77831 -0.86475 -0.9218 -0.95949 -0.98378 -0.99737 -0.99843 -0.96546 -0.58259 0.83712 0.984702 0.99997 0.9939 0.976746 0.948336 0.904874 0.839142 0.739476 0.589621 0.37444 0.098143
-11.4028 -27.1168 -40.3624 -51.1064 -59.8538 -67.1904 -73.6357 -79.666 -85.8445 -86.7895 -74.8961 -35.6329 56.83727 79.96529 89.55611 83.66834 77.61961 71.50224 64.8063 57.04961 47.68679 36.13012 21.98973 5.632246
NA NA NA NA 9.491276 7.306459 2.943804 1.86392 1.40638 1.175185 1.056554 1.008967 1.018496 1.087823 1.238154 1.527452 2.123129 3.704359 13.90393 NA NA NA NA NA
Extraterrestrial Normal irradiance Eₒ 1322.989 1322.989 1322.989 1322.989 1322.989 1322.989 1322.989 1322.989 1322.989 1322.989 1322.989 1322.989 1322.989 1322.989 1322.989 1322.989 1322.989 1322.989 1322.989 1322.989 1322.989 1322.989 1322.989 1322.989
Beam Normal Irradiance Eb NA NA NA NA 14.08882 31.64677 200.3022 346.4244 447.1224 512.6218 551.293 567.9344 564.5468 540.7227 493.5809 417.1804 301.9325 140.4326 3.124773 NA NA NA NA NA
Table continued……
Time (Hrs)
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
Diffuse Horizontal Irrradiance Ed (W) 0 0 0 0 0.026527 0.272143 34.14769 120.6599 209.1295 277.4645 321.2843 340.8626 336.8434 309.0686 256.8032 180.6008 88.76334 14.34026 0.000269 0 0 0 0 0
Surface Azimuth Angle for Exposed wall ψ (⁰) 165 165 165 165 165 165 165 165 165 165 165 165 165 165 165 165 165 165 165 165 165 165 165 165
Surface solar azimuth angle for the exposed wall ƴ (⁰) -176.403 -192.117 -205.362 -216.106 -224.854 -232.19 -238.636 -244.666 -250.845 -251.79 -239.896 -200.633 -108.163 -85.0347 -75.4439 -81.3317 -87.3804 -93.4978 -100.194 -107.95 -117.313 -128.87 -143.01 -159.368
Cos Ɵ for vertical surface
-0.80935 -0.83368 -0.82391 -0.78071 -0.70701 -0.60786 -0.48999 -0.36145 -0.23099 -0.16444 -0.1625 -0.12678 -0.0597 0.034161 0.148412 0.114047 0.040356 -0.0588 -0.17667 -0.30521 -0.43567 -0.55915 -0.66725 -0.75258 61
incident angle for vertical exposed wall/fenestration Ɵ (⁰)
Beam component incident on the vertical wall Et,b (W)
144.0326 146.4785 145.4779 141.3252 134.9925 127.4347 119.34 111.1893 103.3555 99.46442 99.35218 97.28354 93.42267 88.04231 81.46509 83.45135 87.68715 93.37099 100.1757 107.7707 115.8277 123.9971 131.8548 138.8147
0 0 0 0 0 0 0 0 0 0 0 0 0 18.47188 73.25334 47.57811 12.18477 0 0 0 0 0 0 0
Y
0.45 0.45 0.45 0.45 0.45 0.45 0.45 0.45 0.465757 0.486605 0.487252 0.499628 0.525026 0.565294 0.62175 0.60391 0.568145 0.525386 0.482566 0.45 0.45 0.45 0.45 0.45
Diffuse component incident on the vertical exposed wall Et,d (W) 0 0 0 0 0.011937 0.122464 15.36646 54.29697 97.40363 135.0156 156.5464 170.3045 176.8516 174.7146 159.6675 109.0666 50.43047 7.534177 0.00013 0 0 0 0 0
Table continued….
Time (Hrs)
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
Ground reflected component on vertical exposed wall, Et,r
0 0 0 0 0 0.437820324 10.1693087 30.6070955 52.66945522 71.33849072 84.28412188 90.3561568 89.09392773 80.58919437 65.51376257 45.33341115 23.05000841 5.178311758 0.018511284 0 0 0 0 0
Ɵ for horizontal receiving surface
125.8112035 121.4961893 114.2461677 104.9147869 94.18977986 82.54505256 70.29256353 57.64160423 44.74540857 31.74760162 18.90428172 7.785742525 11.04173889 23.24676154 36.19328428 49.17477216 62.00256955 74.53646103 86.60873338 97.98083734 108.2926178 117.0007772 123.345713 126.4705778
cosƟ for Horizontal Surface
-0.58512 -0.52244 -0.41066 -0.25738 -0.07306 0.129747 0.337217 0.535214 0.710242 0.850374 0.946061 0.990782 0.981488 0.918814 0.80703 0.653754 0.469432 0.266625 0.059154 -0.13884 -0.31387 -0.454 -0.54969 -0.59441
Beam Component On Horizontal Surface Et,bh 0 0 0 0 0 4.10606 67.5454 185.411 317.565 435.9204 521.5569 562.699 554.0958 496.8233 398.3344 272.7333 141.7367 37.44285 0.184844 0 0 0 0 0
Diffuse Component On Horizontal Et,dh
0 0 0 0 0 0.272143 34.14769 120.6599 209.1295 277.4645 321.2843 340.8626 336.8434 309.0686 256.8032 180.6008 88.76334 14.34026 0.000269 0 0 0 0 0
Table concluded. From the Table 5, the solar beam, diffuse and reflected components of irradiance on vertical exposed surface/wall and horizontal roof of student computer lab can be tabulated as follow.
62
Table 6. Beam, diffuse and reflected components of radiation on Computer Lab’s wall & roof (June 21)
1 2 3 4
0.67099 1.67099 2.67099 3.67099
0 0 0 0
0 0 0 0
0 0 0 0
Diffuse Component Incident on the vertical exposed wall Et,d 0 0 0 0
5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
4.67099 5.67099 6.67099 7.67099 8.67099 9.67099 10.67099 11.67099 12.67099 13.67099 14.67099 15.67099 16.67099 17.67099 18.67099 19.67099 20.67099 21.67099 22.67099 23.67099
14.08881724 31.64677208 200.3022132 346.4243517 447.122434 512.6218186 551.292981 567.9344386 564.5467634 540.7226692 493.5809242 417.1803735 301.9324565 140.4325868 3.124773324 0 0 0 0 0
0.026527486 0.272143324 34.14768546 120.659944 209.1295143 277.464512 321.2843474 340.8625868 336.8434475 309.0686445 256.8032433 180.6008353 88.76333788 14.34026489 0 0 0 0 0 0
0 0 0 0 0 0 0 0 0 18.47188201 73.25333752 47.57811194 12.18477041 0 0 0 0 0 0 0
0 0.122464496 15.36645846 54.29697482 97.40363448 135.0156326 156.5463908 170.3045455 176.8516165 174.7145996 159.6674745 109.066573 50.43047338 7.534177071 0 0 0 0 0 0
AST= LST+ET/60+(LONLSM)/15
Time (Hrs)
Beam Normal Irradiance Bb
Diffuse Horizontal Irradiance Ed
Beam Component Incident On The vertical wall Et,b
63
Ground reflected component Et,r on vertical exposed wall 0 0 0 0 0 0.437820324 10.1693087 30.6070955 52.66945522 71.33849072 84.28412188 90.3561568 89.09392773 80.58919437 65.51376257 45.33341115 23.05000841 5.178311758 0 0 0 0 0 0
beam component on horizontal surface Et,bh
Total incident radiation on vertical wall
total incident radiation on horizontal surface
0 0 0 0
0 0 0 0
0 0 0 0
0 4.106059921 67.54540152 185.411011 317.5650378 435.9203952 521.5568714 562.6989812 554.0958298 496.8232992 398.3343824 272.7332763 141.7367463 37.44285269 0 0 0 0 0 0
0 0.56028482 25.53576716 84.90407032 150.0730897 206.3541233 240.8305127 260.6607023 265.9455443 273.775676 298.4345746 201.978096 85.6652522 12.71248883 0 0 0 0 0 0
0.026527486 4.378203244 101.693087 306.070955 526.6945522 713.3849072 842.8412188 903.561568 890.9392773 805.8919437 655.1376257 453.3341115 230.5000841 51.78311758 0 0 0 0 0 0
From above table it is evident that on June 21, the vertical wall of the Computer Lab remains most of the time in shade and receives direct or beam radiation for a brief period of time i.e. from 14 hours to 17 hours and that too in very minimal quantities. Maximum incident radiation on the exposed wall of the computer lab is from diffuse radiation.
Hourly Radiation(W/m2) on June 21 over Roorkee, IN 600 500 400 300 200 100 0 1
2
3
4
5
6
7
8
9
10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
beam normal irradiance Eb
diffuse horizontal irradiance Ed
Figure 37. Incident radiation over Roorkee on June 21
Hourly Total Radiation(W/m2) on vertically exposed wall of Student Computer Lab, AHEC, IITR, IN
350 300 250 200 150 100 50 0 1
2
3
4
5
6
7
8
9
10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
Ground reflected component Et,r ( for vertical exposed wall) Diffuse component incident on the vertical exposed wall Et,d beam component incident on the vertical wall Et,b
Figure 38. Total radiation on vertical wall of lab on June 21 64
Hourly Radiation(W/m2) on Horizontal surface on June 21 at Roorkee, IN 600 500 400 300 200 100 0 1
2
3
4
5
6
7
8
9
10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
diffuse horizontal irradiance Ed
beam component on horizontal surface Et,bh
Figure 39. Hourly radiation on Horizontal surface on June 21 at Roorkee
Total incident hourly radiation on the Horizontal surface on June 21 at Roorkee 1000 900
Radiation, W/m2
800 700 600 500 400 300 200 100 0 1
2
3
4
5
6
7
8
9
10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
Hours diffuse horizontal irradiance Ed
beam component on horizontal surface Et,bh
Figure 40. Total incident radiation on Horizontal surface on June 21 at Roorkee
65
Table 7. Solar radiation calculation data table over Student computer Lab, AHEC, IIT Roorkee (January 21)
Time Declination (Hrs) (Δ)
LST
Equation Of Time (ET)
1.00 2.00 3.00 4.00 5.00 6.00 7.00 8.00 9.00 10.00 11.00 12.00 13.00 14.00 15.00 16.00 17.00 18.00 19.00 20.00 21.00 22.00 23.00 24.00
1.00 2.00 3.00 4.00 5.00 6.00 7.00 8.00 9.00 10.00 11.00 12.00 13.00 14.00 15.00 16.00 17.00 18.00 19.00 20.00 21.00 22.00 23.00 24.00
-10.60 -10.60 -10.60 -10.60 -10.60 -10.60 -10.60 -10.60 -10.60 -10.60 -10.60 -10.60 -10.60 -10.60 -10.60 -10.60 -10.60 -10.60 -10.60 -10.60 -10.60 -10.60 -10.60 -10.60
-20.14 -20.14 -20.14 -20.14 -20.14 -20.14 -20.14 -20.14 -20.14 -20.14 -20.14 -20.14 -20.14 -20.14 -20.14 -20.14 -20.14 -20.14 -20.14 -20.14 -20.14 -20.14 -20.14 -20.14
AST= Hour LST+ET/60+(LONAngle(H) LSM)/15 0.52 1.52 2.52 3.52 4.52 5.52 6.52 7.52 8.52 9.52 10.52 11.52 12.52 13.52 14.52 15.52 16.52 17.52 18.52 19.52 20.52 21.52 22.52 23.52
-172.25 -157.25 -142.25 -127.25 -112.25 -97.25 -82.25 -67.25 -52.25 -37.25 -22.25 -7.25 7.75 22.75 37.75 52.75 67.75 82.75 97.75 112.75 127.75 142.75 157.75 172.75
Sin Β -0.98 -0.92 -0.82 -0.66 -0.48 -0.27 -0.06 0.14 0.33 0.48 0.58 0.64 0.64 0.58 0.47 0.32 0.14 -0.07 -0.28 -0.49 -0.67 -0.82 -0.92 -0.98 66
Solar Altitude Angle (Β)
Sin(Φ)
-78.01 -67.26 -54.61 -41.63 -28.67 -15.92 -3.54 8.24 19.08 28.46 35.59 39.51 39.44 35.41 28.19 18.75 7.87 -3.94 -16.33 -29.09 -42.06 -55.03 -67.66 -78.26
-0.61 -0.94 -0.99 -1.00 -0.99 -0.97 -0.93 -0.87 -0.79 -0.65 -0.44 -0.15 0.16 0.45 0.65 0.79 0.88 0.93 0.97 0.99 1.00 0.99 0.94 0.58
Beam Solar Extraterrestrial Air Normal Azimuth Normal Mass(M) Irradiance Angle (Φ) Irradiance Eₒ Eb -37.52 -69.92 -82.94 -88.81 -82.04 -75.58 -68.76 -61.03 -51.77 -40.28 -25.93 -8.84 9.43 26.45 40.70 52.11 61.30 69.00 75.79 82.25 89.05 82.62 69.32 35.65
0 0 0 0 0 0 20.73 6.67 3.03 2.09 1.71 1.57 1.57 1.72 2.11 3.09 6.96 13.06 0 0 0 0 0 0
1322.99 1322.99 1322.99 1322.99 1322.99 1322.99 1322.99 1322.99 1322.99 1322.99 1322.99 1322.99 1322.99 1322.99 1322.99 1322.99 1322.99 1322.99 1322.99 1322.99 1322.99 1322.99 1322.99 1322.99
0.00 0.00 0.00 0.00 0.00 0.00 50.64 287.35 537.63 655.85 715.79 741.56 741.15 714.44 653.11 532.19 275.11 120.77 0.00 0.00 0.00 0.00 0.00 0.00
Table continued……
Time (Hrs)
1.00 2.00 3.00 4.00 5.00 6.00 7.00 8.00 9.00 10.00 11.00 12.00 13.00 14.00 15.00 16.00 17.00 18.00 19.00 20.00 21.00 22.00 23.00 24.00
Diffuse Horizontal Irradiance Ed
0.00 0.00 0.00 0.00 0.00 0.00 19.45 47.76 79.50 98.34 109.43 114.60 114.52 109.17 97.86 78.71 46.37 28.80 0.00 0.00 0.00 0.00 0.00 0.00
Surface Azimuth Angle For Exposed Wall Ψ 165.00 165.00 165.00 165.00 165.00 165.00 165.00 165.00 165.00 165.00 165.00 165.00 165.00 165.00 165.00 165.00 165.00 165.00 165.00 165.00 165.00 165.00 165.00 165.00
Surface Solar Azimuth Angle For The Exposed Wall Ƴ -202.52 -234.92 -247.94 -253.81 -247.04 -240.58 -233.76 -226.03 -216.77 -205.28 -190.93 -173.84 -155.57 -138.55 -124.30 -112.89 -103.70 -96.00 -89.21 -82.75 -75.95 -82.38 -95.68 -129.35
Cos Ɵ For Vertical Surface
-0.19 -0.22 -0.22 -0.21 -0.34 -0.47 -0.59 -0.69 -0.76 -0.79 -0.80 -0.77 -0.70 -0.61 -0.50 -0.37 -0.23 -0.10 0.01 0.11 0.18 0.08 -0.04 -0.13
67
Incident Angle For Vertical Exposed Wall/Fenestration, Ɵ
101.06 102.83 102.56 102.03 110.01 118.19 126.16 133.40 139.20 142.65 142.98 140.09 134.68 127.66 119.79 111.61 103.56 95.99 89.24 83.66 79.62 85.64 92.15 97.41
Beam Component Incident On The Vertical Wall Et,b 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00
Y
0.48 0.47 0.47 0.47 0.45 0.45 0.45 0.45 0.45 0.45 0.45 0.45 0.45 0.45 0.45 0.45 0.46 0.51 0.56 0.60 0.64 0.59 0.53 0.50
Diffuse Component Incident On The Vertical Exposed Wall Et,d 0.00 0.00 0.00 0.00 0.00 0.00 8.75 21.49 35.78 44.25 49.24 51.57 51.53 49.12 44.04 35.42 21.55 14.63 0.00 0.00 0.00 0.00 0.00 0.00
Table continued….
CosƟ For Horizontal Surface
Beam Component On Horizontal Surface Et,Bh
Diffuse Component On Horizontal Et,Dh
168.01
-0.98
0.00
0.00
0.00
157.26
-0.92
0.00
0.00
3.00
0.00
144.61
-0.82
0.00
0.00
4.00
0.00
131.63
-0.66
0.00
0.00
5.00
0.00
118.67
-0.48
0.00
0.00
6.00
0.00
105.92
-0.27
0.00
0.00
7.00
1.63
93.54
-0.06
0.00
19.45
8.00
8.89
81.76
0.14
41.18
47.76
9.00
25.53
70.92
0.33
175.77
79.50
10.00
41.09
61.54
0.48
312.56
98.34
11.00
52.61
54.41
0.58
416.62
109.43
12.00
58.64
50.49
0.64
471.76
114.60
13.00
58.53
50.56
0.64
470.83
114.52
14.00
52.31
54.59
0.58
413.93
109.17
15.00
40.63
61.81
0.47
308.48
97.86
16.00
24.98
71.25
0.32
171.04
78.71
17.00
8.40
82.13
0.14
37.66
46.37
18.00
2.05
93.94
-0.07
0.00
28.80
19.00
0.00
106.33
-0.28
0.00
0.00
20.00
0.00
119.09
-0.49
0.00
0.00
21.00
0.00
132.06
-0.67
0.00
0.00
22.00
0.00
145.03
-0.82
0.00
0.00
23.00
0.00
157.66
-0.92
0.00
0.00
24.00
0.00
168.26
-0.98
0.00
0.00
Ground Reflected Component Et,r ( For Vertical Exposed Wall)
For Horizontal Receiving Surface Ɵ
1.00
0.00
2.00
Time (Hrs)
Table concluded
68
Table 8. Beam, diffuse and reflected components of radiation on Computer Lab’s wall & roof (January 21)
Time (Hrs)
1.00 2.00 3.00 4.00 5.00 6.00 7.00 8.00 9.00 10.00 11.00 12.00 13.00 14.00 15.00 16.00 17.00 18.00 19.00 20.00 21.00 22.00 23.00 24.00
Beam AST= Normal LST+ET/60+(LONIrradiance LSM)/15 Eb 0.52 1.52 2.52 3.52 4.52 5.52 6.52 7.52 8.52 9.52 10.52 11.52 12.52 13.52 14.52 15.52 16.52 17.52 18.52 19.52 20.52 21.52 22.52 23.52
0.00 0.00 0.00 0.00 0.00 0.00 50.64 287.35 537.63 655.85 715.79 741.56 741.15 714.44 653.11 532.19 275.11 120.77 0.00 0.00 0.00 0.00 0.00 0.00
Diffuse Horizontal Irradiance Ed 0.00 0.00 0.00 0.00 0.00 0.00 19.45 47.76 79.50 98.34 109.43 114.60 114.52 109.17 97.86 78.71 46.37 28.80 0.00 0.00 0.00 0.00 0.00 0.00
Beam Component Incident On The Vertical Wall Et,b 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 69
Diffuse Component Incident On The Vertical Exposed Wall Et,d 0.00 0.00 0.00 0.00 0.00 0.00 8.75 21.49 35.78 44.25 49.24 51.57 51.53 49.12 44.04 35.42 21.55 14.63 0.00 0.00 0.00 0.00 0.00 0.00
Ground Beam Reflected Total Component Component Incident On Et,r ( For Radiation Horizontal Vertical On Surface Exposed Vertical Et,bh Wall) Radiation 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 1.63 0.00 10.38 8.89 41.18 30.39 25.53 175.77 61.30 41.09 312.56 85.34 52.61 416.62 101.85 58.64 471.76 110.21 58.53 470.83 110.07 52.31 413.93 101.43 40.63 308.48 84.67 24.98 171.04 60.40 8.40 37.66 29.95 2.05 0.00 16.68 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00 0.00
Total Incident Radiation On Horizontal Surface 0.00 0.00 0.00 0.00 0.00 0.00 19.45 88.95 255.27 410.90 526.05 586.36 585.34 523.10 406.34 249.76 84.03 28.80 0.00 0.00 0.00 0.00 0.00 0.00
Hourly Radiation over Roorkee on January 21 800.00
Radiation W/m2
700.00 600.00 500.00 400.00 300.00 200.00 100.00 0.00 1
2
3
4
5
6
7
8
9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
Hours beam normal irradiance Eb
diffuse horizontal irradiance Ed
Figure 41. Hourly radiation over Roorkee on Jan 21
Hourly radiation on vertical wall on January 21 70.00
Radiation W/m2
60.00 50.00 40.00 30.00 20.00 10.00 0.00 1
2
3
4
5
6
7
8
9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
Hours beam component incident on the vertical wall Et,b Diffuse component incident on the vertical exposed wall Et,d Ground reflected component Et,r ( for vertical exposed wall)
Figure 42. Hourly radiation on vertical wall of lab on Jan 21
70
Total hourly incident radiation on the vertical wall on Jan 21 Radiation W/m2
120.00 100.00 80.00 60.00 40.00 20.00 0.00 1
2
3
4
5
6
7
8
9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
Hours Ground reflected component Et,r ( for vertical exposed wall) Diffuse component incident on the vertical exposed wall Et,d beam component incident on the vertical wall Et,b
Figure 43. Total hourly incident radiation on vertical wall of lab on Jan 21
Hourly incident radiation on Horizontal surface on Jan 21 500.00 450.00
Radiation W/m2
400.00 350.00 300.00 250.00 200.00 150.00 100.00 50.00 0.00 1
2
3
4
5
6
7
8
9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
Hours diffuse horizontal irradiance Ed
beam component on horizontal surface Et,bh
Figure 44. Hourly incident radiation on Horizontal surface (roof) of lab on Jan 21
71
Total hourly incident radiation on Horizontal surface 700.00
Radiation W/m2
600.00 500.00 400.00 300.00 200.00 100.00 0.00 1
2
3
4
5
6
7
8
9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
Hours diffuse horizontal irradiance Ed
beam component on horizontal surface Et,bh
Fig 45. Total hourly incident radiation on horizontal surface (roof) of lab on Jan 21
4.4 Cooling Load Calculation of cooling load has been done according to the ASHRAE Handbook of fundamentals (2013). Cooling load is the result of heat transfer through the building walls and roofs by conduction, convection and radiation plus the heat from internal sources. The various factors affecting the building cooling load are: 1. External source: Walls, roofs, windows/fenestrations, and etc. 2. Internal source: Lightings, occupants, appliances, and etc. 3. Infiltration: Air leakage and moisture migration 4. System: Outdoor air, duct leakage and heat gain, reheat, fan and pump energy, and energy recovery. Space Heat Gain: This rate of heat gain is defined as the rate at which heat enters into the building and generated inside the space. Heat gain can enter through many modes. The different entry modes are (i) solar radiation through windows (transparent surfaces); (ii) heat conducted through exposed walls and roofs; (3) heat conducted through ceilings, floors, and interior partitions; (4) heat generated by people, lightings, and appliances; (5) energy transfer through ventilation and infiltration of outdoor air; and (6) miscellaneous heat gains. Sensible heat is added by conduction, convection, and/or radiation and latent heat gain takes place when moisture is added somehow to the space
72
Space Cooling Load: This is the rate of heat removal from the space to maintain a constant temperature and humidity. By adding all the instantaneous heat gains at any given time does not equal to the cooling load at that given time. The cooling load at any instant depends on the heat gains in past hours in addition to the concurrent heat gains. In present work, the cooling load has been calculated using Radiant Time Series Method (RTS). This method is a simple way to calculate cooling load which is rigorous but quantifies each component’s contribution to the total cooling load. This method is employed to calculate peak design load. RTS method considers the delay of conductive heat gain through wall, roofs and floors and delay of radiative heat gain conversion to cooling loads. The heat conduction through walls and roof takes place due to temperature difference between the outdoor and indoor space. Also, the walls and roof absorb the solar radiation and then transfer the heat to the indoor space. Due to thermal capacity of the building materials, there remains a considerable time delay in heat input. The heat addition to the conditioned space is mostly by combination of convection and radiation. The convective part adds to cooling load immediately while there is a time delay in the radiative part because it is first absorbed, and then heat is transferred by convection from the surface. RTS method takes into consideration both the time delay, conduction and convection by multiplying hourly heat gains by the 24 hour time series. The series coefficients called as radiant time factors and conduction time factors can be referred from the ASHRAE handbook of fundamentals. The whole procedure of RTS method can be overviewed in Fig. 46. 4.4.1 Heat Gain through exterior surface Sol-Air Temperature: Sol-Air temperature is defined as the temperature of outdoor air that gives the same rate of heat transfer into the surface exposed as the combination of incident solar radiation, radiant energy exchange with the sky and other outdoor surroundings, and the convective heat exchange with the outdoor air. The Sol-air temperature is calculated using the equation,
te =+ to
α Et ho
−
ε∆R ho
Where
α = absorptance of surface for solar radiation
Et = total solar radiation incident on surface, W/m2 ho = coefficient of heat transfer by long wave radiation and convection at outer surface, W/m2-⁰C 73
to = outdoor air temperature, ⁰C
te = sol-air temperature temperature, ⁰C ε = hemispherical emittance of the surface ∆R = difference between long wave radiation on surface from sky and surroundings and radiation
emitted by blackbody at outdoor temperature, W/m2 -⁰C
Fig. 46. Overview of Radiant Time Series method [82] The sol-air temperature has been calculated for vertical (wall) and horizontal (roof) exposed surface in Table 6.7. For vertical surface
ε∆R ho
is considered as 0 while for horizontal surface it is
to be taken as 4 K as described by the handbook.
α ho
is considered as 0.026 for the light coloured surface and
0.052 for the dark coloured surface. For the present calculations,
α ho
has been taken as 0.052. The
hourly temperature of outdoor air is required to calculate the sol-air temperature. The 24 h 74
temperature profile was generated for a typical day in June at Roorkee using the ASHRAE handbook of fundamentals data for Dehradun. Table 9. Sol-air temperature calculation for the wall and the roof on June 21
Hour
Tdb 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
24.37 23.81 23.40 22.98 22.70 22.98 23.95 26.31 28.96 31.32 33.40 34.79 35.91 36.60 36.60 35.77 34.65 33.26 31.18 29.65 28.40 27.15 26.18 25.20
Twb
Total incident radiation on Horizontal surface (Roof), Et(Roof) 0.00 0.00 0.00 0.00 0.03 4.38 101.69 306.07 526.69 713.38 842.84 903.56 890.94 805.89 655.14 453.33 230.50 51.78 0.00 0.00 0.00 0.00 0.00 0.00
Total incident radiation on vertical exposed wall, Et(wall)
18.18 17.98 17.84 17.70 17.60 17.70 18.03 18.85 19.76 20.58 21.30 21.78 22.16 22.40 22.40 22.11 21.73 21.25 20.53 20.00 19.57 19.14 18.80 18.46
0.00 0.00 0.00 0.00 0.00 0.56 25.54 84.90 150.07 206.35 240.83 260.66 265.95 273.78 298.43 201.98 85.67 12.71 0.00 0.00 0.00 0.00 0.00 0.00
Sol-air temperature for vertical wall 24.37 23.81 23.40 22.98 22.70 23.06 27.78 39.05 51.47 62.27 69.53 73.89 75.80 77.67 81.37 66.06 47.50 35.17 31.18 29.65 28.40 27.15 26.18 25.20
Sol-air temperature for roof 20.37 19.81 19.40 18.98 18.70 19.63 35.20 68.22 103.96 134.33 155.83 166.33 165.55 153.48 130.87 99.77 65.23 37.03 27.18 25.65 24.40 23.15 22.18 21.20
Heat input at the exterior of the wall and the roof is done by using the conduction time series (CTS). Conductive heat input from wall and roof at the exterior is defined by the conduction equation as = qi ,θ − n UA(t e,θ − n − t rc )
where qi ,θ − n = conductive heat input for the surface n hours ago, W
U = overall heat transfer coefficient, W/m2-K= A = surface area, m2= (23 m2)
75
t e,θ − n = sol-air temperature n hours ago, ⁰C
t rc = room temperature, ⁰C (= 25 ⁰C) Conduction heat gain for the current hour can be calculated using the heat input for the current hour and past 23 hours in the conduction time series:
q= c0 qi ,θ + c1qi ,θ −1 + c2 qi ,θ − 2 + c3 qi ,θ −3 + ... + c23 qi ,θ − 23 θ where
qθ = hourly heat gain through conduction for the surface, W qi ,θ = heat input for the current hour qi ,θ − n = heat input n hours ago
c0 , c1 , c2 , etc.= conduction time factors
4.4.1.1 Wall cooling load using sol-air temperature conduction time series and radiant time series The wall cooling load can be calculated once the conductive heat gain using the sol-air temperature has been obtained. The conductive heat gain is used then used to calculate the wall heat gain using conduction time series (CTS). The wall heat gain can be split into convective and radiative part of the load by using the convective fraction and radiant time series (RTS) according to the wall construction details. In present case, the wall is made of brick with R5 insulation, sheathing and gyp board. The hourly detailed calculation has been presented in the Table 10. For example. At 3PM the, heat gain can be calculated using the conductive time series as,
q= c0 qi ,15 + c1qi ,14 + c2 qi ,13 + c3 qi ,12 + ... + c23 qi ,1 15 Or
q15 = 490.6466 W
Now the total cooling load for the vertical wall can be calculated by summing the convective and radiative portions. The convective and radiative portion has been considered according to ASHRAE handbook of fundamentals. For vertical wall, convective fraction is 0.54. Convective fraction, Qc = 490.6466*0.54 = 264.9 W The radiant cooling load suing the RTS, = qr ,15 r0 qi ,15 (0.46) + r1qi ,14 (0.46) + r2 qi ,13 (0.46) + .... + r23 qi ,16 (0.46) 76
= 234.50 W
Table 10. Hourly heat input profile for the vertical wall using the sol-air temperature with indoor temperature 25⁰C
conduction time factors
Hour
conductive heat gain qi,n
wall heat gain using CTS qi
Radiant heat gain qr,15
Net wall cooling load
61.54482323
29 50.70717
112.252
87.10957
47.03916564
15 42.60173
89.6409
10 35.95078 70.25488
-8.30006 113.9719
Convective heat gain Qconvec
RTS for wall
1
5
2
14
3
17
-21.0785 63.52612
34.30410321
4
15
-26.5549 44.00126
23.76068274
7
29.85676 53.61744
5
12
-30.2059 27.83232
15.02945429
6
23.4356 38.46506
6
9
-25.4512 14.47547
7.816754831
5
18.55364 26.37039
7
7
36.52767 6.238567
3.368826208
4
22.50494 25.87376
8
5
184.5135
12.0233
6.492580047
3
42.66375 49.15633
9
4
347.5775 44.11268
23.82084704
3
73.69091 97.51175
10
3
489.4816 104.2709
56.30630785
3
108.9692 165.2755
11
2
584.7807
184.063
99.39404483
2
141.777 241.1711
12
2
642.1 270.8074
146.2360219
2
169.4948 315.7308
13
1
667.1148 353.7724
191.0371099
2
191.9786 383.0157
14
1
691.6672 426.9627
230.5598344
2
212.3297 442.8895
15
1
740.244 490.6466
264.9491856
1
234.5037 499.4529
16
1
539.2766 536.2479
289.5738576
1
223.9784 513.5523
17
1
295.5422 541.7615
292.5512223
1
190.7878
18
0
133.5741 501.2332
270.6659068
1
154.1289 424.7948
19
0
81.14881 431.8237
233.1848072
1
127.9136 361.0984
20
0
61.06845 355.9177
192.1955461
1
109.3418 301.5373
21
0
44.63907 287.5758
155.2909342
1
94.35689 249.6478
22
0
28.20968 230.1474
124.2795973
0
81.0969 205.3765
23
0
15.43128
183.854
99.28117202
0
70.04558 169.3268
24
0
2.652866
145.982
78.83027476
0
60.1764 139.0067
-15.602
Now the total wall cooling load at 3 PM is calculated as:
77
483.339
Qwall = Qc + Q r,15 = 264.9+234.5 = 499.45 W 4.4.1.2 Roof cooling load using sol-air temperature conduction time series and radiant time series Similar to wall cooling load the roof cooling load has been calculated. Conduction time series (CTS) has been adopted for 6 inch lightweight concrete from the handbook. The radiant time series (RTS) for roof has been considered from the handbook for medium construction without carpet. The hourly cooling load from roof has been presented in Table 11. To understand the calculation, the load at 3 PM has been discussed. At 3PM the, heat gain can be calculated using the conductive time series as, q15 = 987.546 W Now the total cooling load for the roof can be calculated by summing the convective and radiative portions. The convective and radiative portion has been considered according to ASHRAE handbook of fundamentals. For roof, convective fraction is 0.46. Convective fraction Qc = 987.546*0.54 = 533.275 W The radiant cooling load, Qr,15= 561.568 W Now the total roof cooling load at 3 PM is calculated as: Qroof = Qc + Qr,15 = 533.275+561.568 = 1094.84 W 4.4.1.3 Window cooling load using radiant time series: The incident radiation over the wall with windows has already been calculated and tabulated in the previous sections. Referring the Table 8 we can have: Incident beam radiation on window, Et,b = 73.25 W/m2 Incident diffuse radiation on window, Et,d = 159.67 W/m2 Reflected component of radiation, Er = 51.51 W/m2 At 3PM, Incidence angle, Ɵ = 81.47 ⁰
78
Table 11. Hourly heat input profile for the roof using the sol-air temperature with indoor temperature 25⁰C wall Radiant Total Conductive heat Convective RTS conduction heat roof Hour heat gain gain heat gain, for time factors gain cooling qi,n using Qconvec roof qr,15 load CTS qi 1 -61.958 2 386.419 208.666 29 95.3872 304.053 2 -69.395 8 325.109 175.559 15 78.8781 254.437 3 -74.972 11 269.941 145.768 10 64.9637 210.732 4 -80.55 11 222.254 120.017 7 52.7823 172.799 5 -84.216 10 181.575 98.0504 6 40.916 138.966 6 -71.766 9 147.164 79.4687 5 31.1663 110.635 7 136.502 7 122.692 66.2537 4 49.0636 115.317 8 578.173 6 123.139 66.4951 3 112.857 179.353 9 1056.16 5 167.28 90.3313 3 207.131 297.463 10 1462.34 5 262.651 141.831 3 311.988 453.82 11 1749.97 4 401.766 216.954 2 411.866 628.82 12 1890.39 3 566.745 306.042 2 494.486 800.528 13 1879.94 3 735.688 397.272 2 551.202 948.474 14 1718.6 3 887.13 479.05 2 575.023 1054.07 15 1416.13 2 987.546 533.275 1 561.568 1094.84 16 1000.07 2 1070.09 577.85 1 510.492 1088.34 17 538.103 2 1080.68 583.569 1 429.14 1012.71 18 160.933 1 1032.94 557.789 1 337.859 895.648 19 29.1463 1 938.877 506.993 1 269.086 776.079 20 8.6944 1 825.585 445.816 1 223.35 669.167 21 -8.039 1 714.529 385.846 1 188.887 574.733 22 -24.772 1 614.377 331.764 0 160.098 491.862 23 -37.787 1 527.9 285.066 0 135.581 420.646 24 -50.802 1 453.429 244.852 0 114.241 359.092 For calculation of heat gain, solar heat gain coefficient (SHGC) values for beam and diffuse components of radiation are required. From ASHRAE handbook, the value of SHGC and indoor attenuation coefficient (IAC) have been considered for uncoated, single glazing (ID 1a) with sheer louver. SHGC(Ɵ) = 0.42 SHGCd = 0.78 IAC = 0.7 & U = 5.9 W/m2 -⁰C Now, solar beam heat input at 15 hours, qb15 = A * Et ,b * SHGC (θ ) *(IAC) 79
= 13*73.25*0.42*0.7 = 280 W Solar diffuse heat input at 15 hours, qd15 = A*(Et,d+Er)*SHGCd*(IAC)D = 13*(159.67+65.5)*0.78*0.7 = 1598.3 W Conductive heat gain at 15 hours, qc15 = UA*(tout-tin) = 5.9*13*(36.6-25) = 889.72 W Table 12. Window hourly cooling load calculation
hour
1.0 2.0 3.0 4.0 5.0 6.0 7.0 8.0 9.0 10.0 11.0 12.0 13.0 14.0 15.0 16.0 17.0 18.0 19.0 20.0 21.0 22.0 23.0 24.0
incident beam diffuse reflected angle component component beam component on on on SHGC on window windows window window Er Ɵ Et,b Et,d 144.0 146.5 145.5 141.3 135.0 127.4 119.3 111.2 103.4 99.5 99.4 97.3 93.4 88.0 81.5 83.5 87.7 93.4 100.2 107.8 115.8 124.0 131.9 138.8
0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 18.5 73.3 47.6 12.2 0.0 0.0 0.0 0.0 0.0 0.0 0.0
0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.0 0.4 0.4 0.4 0.4 0.0 0.0 0.0 0.0 0.0 0.0 0.0
0.0 0.0 0.0 0.0 0.0 0.1 15.4 54.3 97.4 135.0 156.5 170.3 176.9 174.7 159.7 109.1 50.4 7.5 0.0 0.0 0.0 0.0 0.0 0.0
0.0 0.0 0.0 0.0 0.0 0.4 10.2 30.6 52.7 71.3 84.3 90.4 89.1 80.6 65.5 45.3 23.1 5.2 0.0 0.0 0.0 0.0 0.0 0.0 80
Solar Solar beam diffuse Conductive Total heat heat heat input, window input, input, qci cooling qbi qdi load 0.0 0.0 -48.5 -48.5 0.0 0.0 -91.1 -91.1 0.0 0.0 -123.1 -123.1 0.0 0.0 -155.1 -155.1 0.0 0.0 -176.4 -176.4 0.0 4.0 -155.1 -151.1 0.0 181.3 -80.5 100.8 0.0 602.6 100.8 703.4 0.0 1065.2 303.3 1368.6 0.0 1464.7 484.6 1949.3 0.0 1709.4 644.5 2353.9 0.0 1850.2 751.1 2601.3 0.0 1887.7 836.4 2724.1 70.6 1812.1 889.7 2772.5 280.0 1598.3 889.7 2768.0 181.8 1095.9 825.8 2103.5 46.6 521.6 740.5 1308.6 0.0 90.2 633.8 724.1 0.0 0.0 473.9 473.9 0.0 0.0 356.7 356.7 0.0 0.0 260.7 260.7 0.0 0.0 164.8 164.8 0.0 0.0 90.1 90.1 0.0 0.0 15.5 15.5
According to ASHRAE Handbook Fundamentals, if there is any kind of inside shading in the window/fenestration, the window cooling load can be calculated by adding the direct beam, diffuse and conductive heat gains together. In present case, the windows are provided with the sheer louver. So, the total window cooling load at 3 PM, Qwindow = 2768 W 4.4.2 Lighting/appliances cooling load In the computer lab the internal heat gain is by lighting, computers, fan motors, all-in-one printer, and scanners and occupants. The heat gains from the appliances in the lab can be calculated by knowing their wattage or rating or the average power consumed. The ASHRAE Handbook of Fundamentals 2013 gives the guidelines to calculate the heat gain from appliances.
1 − Em Heat gain from the fan motor is calculated as, qem = P Em
FUM FLM where;
qem is the heat gain, P is the rating of the motor, Em is the efficiency of the motor, FUM is the utilization factor and FLM is the motor load factor. The present calculations have been done considering an efficiency of 90% and the load and utilization factor as 1.
Table 13. Heat gain from lighting and appliances S.
Appliance
No.
Quantity Wattage/rating/Average
Heat
(Nos.)
power (W)
Gain (W)
1.
Fan (Motor)
4
70
28
2.
Fluorescent tubes
12
36
518.4
3.
Computer (2.3 GHz, 3
14
97+19=116
1624
GB Ram) + LCD Screen (380mm) 4
Printer/Copier
1
1350
775
5
Scanner
1
60
15
Total
2960.4
Heat Gain from lighting has been calculated as qel = WFul Fsa ; where W is the wattage of the lighting and Fsa is the special allowance factor which considers the ballast in the fluorescent lights. Here, the ballast factor considered is 1.2 and utilization factor as 1. The average power for 81
computer has been referred from the ASHRAE handbook of fundamentals. The heat gains from printer and scanner have been calculated directly by multiplying the wattage to the utilization factor. Utilization factor for printer and scanner are 0.5 and 0.25 respectively. The hourly heat gain profile has been presented in the Table 14. Using non solar RTS, the hourly load has been calculated for 24 hours and tabulated in Table 14.
Table 14. Lighting/Appliances heat gain profile based on occupancy schedule and hourly load Lighting heat gain
Hour
Convective portion
Radiative portion
Radiant load
RTS
Total lighting load by lighting
1
0
0
0
16
203.972
203.97156
2
0
0
0
10
188.281
188.28144
3
0
0
0
7
172.591
172.59132
4
0
0
0
5
156.901
156.9012
5
0
0
0
4
141.211
141.21108
6
0
0
0
3
125.521
125.52096
7
0
0
0
3
627.605
627.6048
8
2960.4
1391.388
1569.012
2
862.957
2254.3446
9
2960.4
1391.388
1569.012
2
1004.17
2395.55568
10
2960.4
1391.388
1569.012
2
1098.31
2489.6964
11
2960.4
1391.388
1569.012
2
1161.07
2552.45688
12
2960.4
1391.388
1569.012
1
1208.14
2599.52724
13
2960.4
1391.388
1569.012
1
1239.52
2630.90748
14
2960.4
1391.388
1569.012
1
1270.9
2662.28772
15
2960.4
1391.388
1569.012
1
1302.28
2693.66796
16
2960.4
1391.388
1569.012
1
1333.66
2725.0482
17
2960.4
1391.388
1569.012
1
847.266
2238.65448
18
0
0
0
1
627.605
627.6048
19
0
0
0
1
486.394
486.39372
20
0
0
0
1
392.253
392.253
21
0
0
0
1
329.493
329.49252
22
0
0
0
0
282.422
282.42216
23
0
0
0
0
251.042
251.04192
24 0 0 0 33 219.662 219.66168 The convective portion of load can be calculated by simply multiplying the heat gain with the convective fraction. 82
From the handbook, for non-in-ceiling fluorescent luminaire the radiative fraction can be taken as 0.53 which implies the convective fraction to be 0.47. At 3 PM, the convective portion, Qc = 2960.4*0.47= 1391.4 W. The radiant portion of the cooling load is calculated using lighting heat gains for the current hour and past 23 h. The radiant time series (RTS) for medium-weight construction, 50% glass and non-carpeted floors can be referred from the handbook. Thus, the radiant cooling load for lighting is:
= Qr r1 (0.53)q15 + r2 (0.53)q14 + r3 (0.53)q13 + ... + r24 (0.53)q16 where r1, r2, r3….etc. are the RTS or radiant factor coefficients. The radiant portion, Qr =1302 W The total lighting and appliances cooling load, Qlight+app = Q c+Qr = 1391.4+1302 = 2693 W 4.4.3 Internal/occupants cooling load Table 15. Hourly lighting/appliances heat gain and cooling load
Hour 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24
Lighting heat gain 0 0 0 0 0 0 0 2960.4 2960.4 2960.4 2960.4 2960.4 2960.4 2960.4 2960.4 2960.4 2960.4 0 0 0 0 0 0 0
Convective portion
Radiant portion
0 0 0 0 0 0 0 1190 1190 1190 1190 1190 1190 1190 1190 1190 1190 0 0 0 0 0 0 0
0 0 0 0 0 0 0 630 630 630 630 630 630 630 630 630 630 0 0 0 0 0 0 0 83
RTS 16 10 7 5 4 3 3 2 2 2 2 1 1 1 1 1 1 1 1 1 1 0 0 33
Radiant load 81.9 75.6 69.3 63 56.7 50.4 252 346.5 403.2 441 466.2 485.1 497.7 510.3 522.9 535.5 340.2 252 195.3 157.5 132.3 113.4 100.8 88.2
Total occupant cooling load 81.9 75.6 69.3 63 56.7 50.4 252 1536.5 1593.2 1631 1656.2 1675.1 1687.7 1700.3 1712.9 1725.5 1530.2 252 195.3 157.5 132.3 113.4 100.8 88.2
Human body continuously emits sensible and latent heat in different states of activity. The occupants inside the space add significantly to the cooling load. For moderately active work a human emits 75 W of sensible and 55 W of latent heat. No of occupants, N: 14 Nos. Sensible heat gain, qs = qs,per*N= 75*14 = 1050 W Latent heat gain, ql = ql,per *N = 55*14 = 770 W As per the ASHRAE Handbook fundamentals, 60% of the sensible load adds to the radiant load. So, Convective portion of heat gain can be calculated as, Qconv = 770+1050*0.4 = 1190 W and the Radiant portion can be obtained as, Qrad = 1050*0.6 = 630 W. The radiant cooling load can be found out using the RTS from handbook. The hourly heat gain schedule and the cooling load due to lighting and appliances has been calculated and tabulated in Table 15.
4.4.4 Total cooling load
The total load can thus be found by adding the hourly loads from all the sources i.e. wall cooling load, roof cooling load, window cooling load, lighting cooling load and the occupants cooling load. The 24 hour cooling load profile can be obtained by adding the hourly loads from all sources. Table 16 summarizes the cooling loads for 24 hours in hourly basis.
From Table 16, the peak cooling demand on the hottest day is 8767 Watts or 8.77 kW. In terms of refrigeration ton (RT), the peak cooling load is 2.5 Ton. The cooling demand has an increasing trend in the morning hours from 6AM till 3PM in the afternoon. The peak cooling demand occurs at 3 PM and minimum cooling demand is 161 W at 6 AM. There is a sharp increase in demand at 8 AM and the demand declines significantly at 6 PM. The hourly cooling demand has been plotted in the fig. 47.
84
Table 16. Total cooling load hourly profile for 24hrs on June 21 Total hourly cooling load (W)
Window cooling load (W)
Wall cooling load (W)
Roof cooling load (W)
Lighting load (W)
Occupant load (W)
1
-48.47
112.25
304.05
203.97
81.90
653.70
2
-91.12
89.64
254.44
188.28
75.60
516.84
3
-123.10
70.25
210.73
172.59
69.30
399.77
4
-155.09
53.62
172.80
156.90
63.00
291.23
5
-176.41
38.47
138.97
141.21
56.70
198.93
6
-151.11
26.37
110.64
125.52
50.40
161.82
7
100.79
25.87
115.32
627.60
252.00
1121.59
8
703.43
49.16
179.35
2254.34
1536.50
4722.79
9
1368.57
97.51
297.46
2395.56
1593.20
5752.30
10
1949.29
165.28
453.82
2489.70
1631.00
6689.08
11
2353.93
241.17
628.82
2552.46
1656.20
7432.57
12
2601.29
315.73
800.53
2599.53
1675.10
7992.18
13
2724.09
383.02
948.47
2630.91
1687.70
8374.19
14
2772.47
442.89
1054.07
2662.29
1700.30
8632.02
15
2768.03
499.45
1094.84
2693.67
1712.90
8768.89
16
2103.53
513.55
1088.34
2725.05
1725.50
8155.97
17
1308.60
483.34
1012.71
2238.65
1530.20
6573.50
18
724.08
424.79
895.65
627.60
252.00
2924.13
19
473.93
361.10
776.08
486.39
195.30
2292.80
20
356.66
301.54
669.17
392.25
157.50
1877.11
21
260.70
249.65
574.73
329.49
132.30
1546.88
22
164.75
205.38
491.86
282.42
113.40
1257.81
23
90.12
169.33
420.65
251.04
100.80
1031.94
24
15.49
139.01
359.09
219.66
88.20
821.45
Hour
85
Hourly Total Cooling Load 10000.00 9000.00
Cooling load in Watts
8000.00 7000.00 6000.00 5000.00 4000.00 3000.00 2000.00 1000.00 0.00 1
2
3
4
5
6
7
8
9
10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 Hour
Fig. 47. Hourly cooling load on June 21
4.5 Heating load Space heating load is determined by computing the heat transfer rate through building envelop elements plus heat required because of outside air infiltration. For the coldest month, January the 99.6 % design temperature for heating is 5.3 ⁰C from the data available from ASHRAE handbook of fundamentals. The 99.6% design condition can be considered for peak design as only for 0.4% occurrences the temperature falls below 5.3 ⁰C. Heat transfer from the space to the surroundings can simply be calculated from the heat transfer equation, Q = UA∆T . The overall heat transfer coefficients (U) can be obtained from the handbook according to the build, as already done in the previous sections. Window heating load, Qheating, window = 5.9*12.75*(25-5.3) = 1482 W Wall heating load, Qheating, wall = 0.514*10.25*(25-5.3) = 104 W Roof heating load, Qheating, roof = 0.834*44*(25-5.3) = 723 W Heat lost due to infiltration air also needs to be taken into account. Considering the air infiltration rate to be 55l/s/100sq-m plus 3.5 l/s/occupant, for 14 occupants, the net infiltration rate can be calculated= as ma 15*
44 + 3.5*14 = 55.6 l/s. 100 86
Now, net heat loss due to infiltration, Q = ma C ∆T ; where, heat capacity of air, C =1.23 J/l-⁰C a and temperature difference, ∆T = 25-5.3 ⁰C. So, heat loss due to infiltration Qa = 55.6*1.23*(25-5.3) = 1347.24 W Also, contribution of lighting and appliances to the heat addition in the space by factor of 0.5, Qlight, heat = 0.5*2960 = 1480 W Now, net heating load, Q heat =Qheating, window + Qheating, wall + Qheating, roof + Qa - Qlight, heat = 2176.24 W
4.6 CLOSURE Selected site for the design of the system, Students’ Computer Lab, Alternate Hydro Energy Centre at IIT Roorkee, India has been introduced. Various required details for the design purpose have been collected and presented in the chapter. The chapter starts with a brief information of the location of the site. The construction details are very important for the building cooling and heating load evaluation, hence, have been detailed. The meteorological data for the location have been listed. The heating and cooling load are governed by solar radiation available at the by a large extent, hence hourly solar data have been extensively calculated and tabulated for each exposed surface (wall and roof). The cooling load was calculated using the RTS method as explained in the chapter after calculating the lighting and electronic appliances contribution to it. An hourly cooling load profile was generated using the radiant time series (RTS) method. The peak cooling load was 2.5 Ton at 3PM on June 21. A peak heating load of 2176.24 W was calculated in a typical coldest day in January. The following chapter will deal with the analysis of the absorption cycle and design of the system based on the calculated cooling load calculated in the present chapter.
87
CHAPTER 5
DESIGN OF SOLAR BASED LiBr-WATER ABSORPTION COOLING SYSTEM
5.1 Design of Absorption Chiller The cooling load has been calculated in Chapter 4 as 8.77 kW. Once the load is known the different component of the absorption chiller must be sized and different parameters need to be found out. An absorption cooling system can be plotted in a Pressure – Temperature (P-T) diagram as follows:
Fig. 48. LiBr-Water Absorption Cooling System At the evaporator, Qg amount of heat flows in from the surroundings to evaporate the refrigerant water. The refrigerant vapor gets absorbed by the LiBr solution at the absorber releasing Qa amount of heat. The weak solution (Point 1) gets then pumped generator at higher pressure through the heat exchanger. The solution gives off the refrigerant vapor by absorbing the Qg amount of 88
heat from some external source, here by the solar hot water. The refrigerant vapor gets condensed at the condenser by releasing the Qc amount of heat. The pressure of the condensed refrigerant water is reduced by passing it through the throttle valve before it reaches to the evaporator to again extract heat from the conditioned space. Once the refrigerant vapor is given off by the solution at the generator, LiBr solution becomes stronger i.e. the concentration of LiBr is more (Point 4). This strong solution is passed through the solution heat exchanger to lose some heat. The solution is passed through the throttle valve to decrease its pressure before it reaches to the absorber to again absorb the refrigerant vapor coming from the evaporator. To size the chiller components and to evaluate the various parameters some assumptions and input values need to be considered. Assumptions: 1. The steady state conditions exist. 2. The refrigerant is pure water. 3. There are no pressure changes except through the throttle valve and the pump i.e.
P= P10= P= P6 and P= P= P= P= P= P5 9 1 8 7 4 3 2 4. Referring to fig.1, at points 1, 4, 8 and 11, there is only saturated liquid. 5. At point 10, there is only saturated vapor. 6. Throttling is isenthalpic. 7. Pumping is an isentropic process. For designing the components, mathematical modeling has been performed for each component. Before the mathematical modeling, design parameters are defined as in Table 15. Table 17. Design parameters for absorption chiller Parameter
Value
Evaporator temperature, T10
10 ⁰C
Generator solution exit temperature, T4
75 ⁰C
Weak solution concentration (refrigerant laden), X1
55% LiBr
Strong solution concentration, X4
60% LiBr 65 ⁰C
Solution temperature at the inlet of generator, T3
70 ⁰C
Generator vapor exit temperature, T7
89
5.1.1 Mathematical Modeling of different chiller components Evaporator m9 = m10
(i)
Q= m10 (h10 − h9= ) m9 (h10 − h9 ) e
(ii)
Absorber = m m10 + m6 1
(iii)
x1m1 = x1m6
(iv)
Qa = m10 h10 + m6 h6 − m1h1
(v)
Pump = W p m1v1 ( P1 − P2 )
(vi)
Heat Exchanger m2 h2 + m4 h4 = m3 h3 + m5 h5
(vii)
Generator m= m4 + m7 3
(viii)
m3 x3 = m4 x4
(ix)
Qg = m4 h4 + m7 h7 − m3 h3
(x)
Condenser = Qc m7 h7 − m8 h8
(xi)
90
Throttle Valve h5 = h6
(xii)
h8 = h9
(xiii)
Here, m is the mass flow rate, Q is the heat, h is the enthalpy, x is the mass fraction of LiBr, Wp is the pump work, P is the pressure and v is the specific volume.
5.2 Parameters calculations Using the assumptions and the input values the parameters can be calculated from the mathematical equations. At the evaporator the temperature T10 taken as input is 10 °C. The refrigerant vapor in the evaporator is in saturated state. The pressure and the enthalpy can be known using the steam tables.
P10 = 1.228 kPa and h10 =2619 kJ/kg The cooling load, Qe = 8.77 kW Once the value h9 is known, the required refrigerant mass flow m9 rate can be calculated from equation (i). At point (8), the condensed water (refrigerant) is saturated liquid. The enthalpy at this point can be known from the steam tables, if the pressure is known. The pressures at point (4) and point (8) can be assumed to be equal for present study. At generator, the solution (55% LiBr) leaves at 75 ⁰C ( T4 =75). Referring the ASHRAE Handbook of Fundamentals [82] we have;
h4 = 183.23 kJ/kg and T ' = 31.46 ⁰C = 304.46 K
Using the empirical method to find the pressure;
log P = C + D / T '+ E / T '2 ; Where C = 7.05, D = -1596.49 and E = -104095.5
P4 = 4.824 kPa Now corresponding to pressure P8 = 4.824 kPa, the enthalpy at point 8 can be evaluated from steam tables as
h8 = 135.1136 kJ/kg and as h8 = h9 ; h9 = 135.1136 kJ/kg. Using the equations (i) and (ii) we get; Mass flow rate of refrigerant, m= m= m= m10 = 3.679x10-3 kg/s 7 8 9 91
m1 = 44.14x10-3 kg/s and m6 = 40.46x10-3 kg/s Following the ASHRAE Handbook Fundamentals [82], the enthalpy of the solution at point (1) can be calculated empirically knowing the pressure level and concentration. Using the steam tables (2), corresponding to 10 ⁰C saturated water, we have
P1 = 1.228 kPa.
From pressure P1 = 1.228 kPa, the refrigerant saturation temperature is calculated as,
T1 ' = 9.325 ⁰C and solution temperature T1 = 39.40 ⁰C Solution temperature and concentration are input to calculate the enthalpy, h1 and same has been calculated as;
h1 = 92.41kJ / kg The pump takes the solution from the absorber, point (1) and delivers at the heat exchanger inlet, point (2) increasing the pressure from 1.228 kPa to 4.824 kPa. The electrical input power to the pump can be calculated using the equation (vi). The specific volume v1 does not possess considerable change with change in pressure. The specific volume can be calculated as presented by Stankus, S .V. et. al.[83].
ρ1 = 1568.28kg / m3 or v1 = 6.376 x10−4 m3 / kg Now using the equation (vi): W p = 0.0981W h1 + W p / m1 = 92.418kJ / kg Now, h2 =
It’s evident that there is no appreciable change in the enthalpy of the solution due to pump since the required electrical input is very minimal. Hence, there is no noticeable change in the solution temperature before and after the pump.
92
So,
= T2 39.4°C
Corresponding to pressure P3 the temperature T3 is less than the saturation temperature, so the solution is sub-cooled. In sub-cooled conditions pressure doesn’t affect the enthalpy of the solution considerably. Hence, at point 3 the temperature ( T3 ) and the mass fraction of the solution ( x3 ) can be used as inputs to get the enthalpy ( h3 ). Calculated enthalpy at point 3 is:
h3 = 145.38kJ / kg Now from the equation (vii), enthalpy at point 5 can be calculated as:
h5 = 125.46kJ / kg Throttling model gives:
h= h= 125.46kJ / kg 5 6 The solution temperature at point 5 is calculated from the empirical relation given in ASHRAE Handbook of fundamentals [82] using the enthalpy value.
= T5 45.05°C Now the temperature at point 6 can be calculated as:
= T6 49.15°C Now, using the equation (v)
Qa = 8.61kJ / kg At point 7, the refrigerant vapor is at temperature (T7) 70 ⁰C and pressure (P7) 4.824 kPa.
Referring the steam tables (2) the enthalpy for this superheated vapor can be found out as:
h7 = 2631.247kJ / kg Using (x), heat added at the generator, Qg can be calculated as: 93
Qg = 10.67 kW
The amount of heat rejected at the condenser, Q c is calculated from equation (xi) as:
Qc = 9.18kW All the parameters of LiBr-water absorption chiller are tabulated in Table 16 Table 18. All calculated parameters of the absorption chiller
Point
h(kJ/kg)
m(kg/s)
P(kPa)
T(⁰C)
92.41863
0.037045
1.228184
2
92.41871
0.037045
4.824532
39.40458144
3
145.3803
0.037045
4.824532
65
4
183.2322
0.033958
4.824532
75
5
125.4559
0.033958
4.824532
45.05422587
6
125.4559
0.033958
1.228184
49.15158446
7
2631.247
0.003087
4.824532
70
8
135.1136
0.003087
4.824532
32.24113675
9
135.1136
0.003087
1.228184
10
10
2519.23
0.003087
1.228184
10
1
All the energy flows are listed in Table 17. A COP of 0.82 was achieved for the present considered parameters.
94
39.40458144
Table 19. Energy Flows in the absorption chiller
Particulars
Quantity
Cooling load, Qe
8.77 kW
Heat rejected at absorber, Qa
10.26 kW
Pump work, Wp
0.0981 W
Heat supplied at generator, Qg
10.67 kW
Heat rejected at condenser, Qc
9.18 kW
5.3 Solar Collector area The heat required at the generator needs to be supplied from the hot water coming from the solar collector. The quantity of heat required to be absorbed by the solar collector and the average radiation incident on the collector need to be known to calculate the total area of the collector field. For solar collectors, optimum tilt angle is latitude plus 10 degrees in the northern hemisphere of earth. So, tilt angle at Roorkee; Ɵt = 29.9+10 = 39.9 = 40 ⁰ Now, solar radiation at any surface with tilt angle of 40 ⁰ at Roorkee can be calculated hourly as shown in Table 18. Considering average radiation incident on collector between the sunshine hour from 0800 hours to 1700 hours, Etotal,avg = 246.2 W/m2. As in the present study the hot water is supplied from the flat plate solar collector. Considering the efficiency of the solar collector to be 80%, to supply the 8.959 kW of heat to the generator, the amount of solar heat required to be absorbed by the collector can be calculated as: 10.67/0.8 = 13.33 kW. The area of collector required can be calculated as: 13330 W/246.2 W/m2 = 54 m2 95
Table 20. Radiation components over the collector at an angle of 40 ⁰
Beam Component Incident On The Collector
Time(Hrs)
Diffuse Component Incident On Solar Collector
Ground Reflected Component Et,R ( For Vertical Exposed Wall)
Total Radiation incident on the collector
8
51.51329
72.79092
29.68802806
153.9922
9
22.04417
117.8294
51.80702348
191.6806
10
16.1639
155.1596
70.64914538
241.9726
11
48.1764
188.8559
83.81640341
320.8487
12
64.84681
205.4392
90.12770094
360.4137
13
62.08251
202.7244
89.10203644
353.9089
14
40.537
181.0676
80.81403542
302.4186
15
5.658417
143.3295
65.91639009
214.9043
16
31.89286
106.8487
45.84748868
184.589
17
55.41687
58.16302
23.55496223
137.1348
5.4 Flow rate of hot water The solution enters and leaves the generator at the temperatures 65 C and 75 C respectively. Let the hot water coming from the solar collector is at 80 C. Consider the water enters the solar collector at a temperature of 25 C. The mass flow rate of the water through the collector can be calculated using simple energy balance. Heat carried by the hot water, Qw = mwC p (80 − 25) =Qg ; Where; mw is the mass flow rate of water and Cp is the heat capacity of the water. Now, mw = 10670/(4200*55) = 0.0462 kg/s = 2.77 kg/m
96
5.5 Volume of hot water storage tank Literature review [65] concluded that optimum specific hot water storage tank volume should be 0.01 to 0.08 m3/m2. For heating and cooling system the higher value should be chosen for design. Taking specific volume of the storage tank as 0.08 m3/m2 , the volume of hot water storage tank can be calculated as, V = 0.08*45 = 4.32 m3.
5.6 RESULTS AND DISCUSSION The cooling load was calculated for each hour on June 21 using the radiant time series method (RTS) which includes the effect of heat gains in last 24 hours by convection and radiation. The obtained cooling load profile for 24 hours shows that the maximum load is 8.77 kW at 3PM while minimum cooling demand occurs at 6 PM. The chiller was simulated and other parameters of the chiller were obtained using the spreadsheet program using the Macros feature. The simulation of the program also gave the related heat transfers at various points in the absorption chiller. For a peak cooling load of 8.77 kW, the heat input required at the generator was found as 10.67 kW. This demand of thermal energy can be fulfilled by a solar collector field of 54 m2 which supplies the water at 80 C with a flow rate of 2.77 kg/min. For Roorkee, optimum collector tilt angle was taken as 40⁰. The optimum volume of the storage tank was calculated using the results available in the literature, which is 4.32 m3 . Total heat rejected from the chiller i.e. at condenser and absorber, has been calculated as 19.44 kW. It can be noticed that to meet the cooling requirement of 8.77 kW, the electricity required is 0.0981 W. This proves that solar absorption systems can save the electricity significantly and thus save the environment. Also, the heating requirement can be fulfilled by direct use of hot water from the solar collector or the hot water storage tank. The hot water can also be fed to the fan coil unit bypassing the absorption chiller to heat the cold air in winter.
97
CHAPTER 6
CONCLUSION AND RECOMMENDATION
6.1 Conclusion The study shows that the solar absorption cooling and heating system prove to be best among all the other alternative options. For the selected site, Students’ Computer Lab, Alternate Hydro Energy Centre, IIT Roorkee, the LiBr-water based absorption cooling and heating system is proposed which can be powered by the hot water from flat plate collector field. A theoretical design has been presented in this dissertation. The design procedure was performed for the peak cooling load condition. The cooling load was calculated following the ASHRAE guidelines and Radiant Time Series Method and accordingly the various design parameters were calculated. The cooling and heating loads were 8.77 kW and 2.18 kW respectively. For the design equilibrium generator temperature of 80 C, the thermal energy required at the generator was 10.67 kW. For an 80% efficient flat plate collector the required collector field comes out to be 54 m2 according to the calculated average radiation available at the site. The rate of flow of the hot water was calculated as 2.77 kg/m. For continuous and efficient operation, the hot water storage tank volume has been included and its volume came out to be 4.32 m3. The use of flat collector in the present work which gives hot water at 80 C proves that the chiller can be integrated with the domestic solar hot water systems. The integration can fulfill the cooling demand as well. This technology can thus save electricity and thus the CO 2 emission by large quantity as cooling and heating systems constitute a large portion in electricity consumption. The use of solar thermal energy thus can reduce the burden of fossil fuel import and can help improve the economy of the country. The use of this technique is the novel step towards the sustainable energy development and in saving the environment.
98
6.2 Recommendations According to the literature and the present dissertation following recommendations can be made. 1. The study can be continued with the design for the complete AHEC building at IIT Roorkee. 2. An experimental setup can be developed and tested. 3. A detailed cost analysis can be performed for such systems. 4. The change in the performance of the cycle can be studied with the change in cooling medium (water to air). 5. The system can be tested for innovative heat rejection systems like geothermal sink as mentioned in the literatures. 6. The design can be done by considering the annual energy consumption and nominal loads. 7. The system can be evaluated by integrating it with the biomass plants. 8. A tri-generation plant producing heating, cooling and electricity simultaneously can also be designed. 9. The performance and the cost analysis can be done for the proposed site for different solar collection systems. 10. The design can be continued with the transient analysis at the solar thermal storage tank and the chiller.
99
List of Publications 1. Arvind Gupta, R P Saini, ‘Solar Absorption Cooling and Heating Systems: A Review’, National Conference on Recent Advances in Mechanical Engineering - 2014, 28-29 June 2014, GB Pant Engineering College, Pauri Garhwal, (UK), India (Accepted). 2. Arvind Gupta, R P Saini, ‘Design of LiBr-water based solar absorption cooling and heating system’, International Conference on Energy Technology, Power Engineering and Environmental Sustainability’ 28 June 2014, JNU, New Delhi, India (Communicated).
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