/smash/get/diva2:744102/FULLTEXT01.pdf

/smash/get/diva2:744102/FULLTEXT01.pdf
Analysis and design of solar based
systems for heating and cooling of
buildings
Igor Shesho
Master's Thesis
Submission date: July 2014
Supervisor:
Vojislav Novakovic, EPT
Co-supervisor:
Laurent Georges, EPT
Norwegian University of Science and Technology
Department of Energy and Process Engineering
MASTER THESIS
Analysis and design of solar based systems for heating
and cooling of buildings
Igor Shesho
Supervisor: Vojislav Novakovic
Co-supervisor: Laurent Georges
Submission date: July 2014
Norwegian University of Science and Technology
Department of Energy and Process Engineering
Norwegian University
of Science and Technology
Department of Energy
and Process Engineering
EPT-M-2014-145
MASTER THESIS
for
Student Igor Shesho
Spring 2014
Analysis and design of solar based systems for heating and cooling of buildings
Analyse og design av systemer for bruk av solenergi for oppvarming og kjøling av bygninger
Background and objective
According the IEA (International Energy Agency) buildings represents 32% of the total
final energy consumption and converted in terms of primary energy this will be around 40%.
Inspected deeper, the heating energy consumption represents over 60% of the total energy demand
in the building. Space heating and hot water heating account for over 75% of the energy used in
single and multi-family homes. Solar energy can meet up to 100% of this demand.
Solar technologies can supply the energy for all of the building’s needs—heating, cooling,
hot water, light and electricity—without the harmful effects of greenhouse gas emissions created by
fossil fuels. Ussualy the maximum demand for cooling coincides with the maximum avalibility of
solar radiation, whereas conventional conventional electrical-compressor chillers have the problem
of providing their minimum capacity in the hottest hours.
The main objective of this thesis is optimizing solar driven air-conditioning systems
through developing inovative combination of solar heating, cooling and storage techniques in
regard of primary energy savings and economic viability. Objective will be achieved through
transitient dynamic simualtions and a holistic approach for varios configurations between solar
heating and cooling technologies, energy storage and auxiliary conventional heating and cooling
devices.
This assignment is realised as a part of the collaborative project “Sustainable Energy and
Environment in Western Balkans” that aims to develop and establish five new internationally
recognized MSc study programs for the field of “Sustainable Energy and Environment”, one at
each of the five collaborating universities in three different WB countries. The project is funded
through the Norwegian Programme in Higher Education, Research and Development in the
Western Balkans, Programme 3: Energy Sector (HERD Energy) for the period 2011-2014.
The following tasks are to be considered:
1. Literature review of solar based systems for heating and cooling of buildings.
2. Define the solar heating and cooling components, reference system and various configurations
of the SHC technologies subject for optimization
3. Numerical modelling of the solar heating and cooling system components (solar colectors,
storage tanks, absorption chiller, heat pump)
4. Dynamic simulation and optimization of the defined solar air-conditioning configurations
5. Simulation results, analysis and conclusions
6. Make a draft proposal (8-10 pages) for a scientific paper based on the performed work in the
master thesis.
-- ” -When the thesis is evaluated, emphasis is put on processing of the results, and that they are
presented in tabular and/or graphic form in a clear manner, and that they are analyzed carefully.
The thesis should be formulated as a research report with summary both in English and Norwegian,
conclusion, literature references, table of contents etc. During the preparation of the text, the
candidate should make an effort to produce a well-structured and easily readable report. In order to
ease the evaluation of the thesis, it is important that the cross-references are correct. In the making
of the report, strong emphasis should be placed on both a thorough discussion of the results and an
orderly presentation.
The candidate is requested to initiate and keep close contact with his/her academic supervisor(s)
throughout the working period. The candidate must follow the rules and regulations of NTNU as
well as passive directions given by the Department of Energy and Process Engineering.
Risk assessment of the candidate's work shall be carried out according to the department's
procedures. The risk assessment must be documented and included as part of the final report.
Events related to the candidate's work adversely affecting the health, safety or security, must be
documented and included as part of the final report. If the documentation on risk assessment
represents a large number of pages, the full version is to be submitted electronically to the
supervisor and an excerpt is included in the report.
Pursuant to “Regulations concerning the supplementary provisions to the technology study
program/Master of Science” at NTNU §20, the Department reserves the permission to utilize all the
results and data for teaching and research purposes as well as in future publications.
The final report is to be submitted digitally in DAIM. An executive summary of the thesis
including title, student’s name, supervisor's name, year, department name, and NTNU's logo and
name, shall be submitted to the department as a separate pdf file. Based on an agreement with the
supervisor, the final report and other material and documents may be given to the supervisor in
digital format.
Work to be done in lab (Water power lab, Fluids engineering lab, Thermal engineering lab)
Field work
Department of Energy and Process Engineering, 20. February 2014
Olav Bolland
Department Head
Research Advisor:
Vojislav Novakovic
Academic Supervisor
Lauren Georges, NTNU
Acknowledgement
This assignment is realised as a part of the collaborative project “Sustainable Energy and
Environment in Western Balkans” that aims to develop and establish five new internationally
recognized MSc study programs for the field of “Sustainable Energy and Environment”, one at
each of the five collaborating universities in three different WB countries. The project is funded
through the Norwegian Programme in Higher Education, Research and Development in the
Western Balkans, Programme 3: Energy Sector (HERD Energy) for the period 2011-2014.
First off all I would like specially to thank and express my sincere gratitude to my supervisor,
Professor Vojislav Novakovic for providing me the opportunity to work, research and broaden my
perspectives on such reputable University - NTNU.
Secondly I want to thank to my co-supervisor Laurent Georges for his advices and guidance,
helping me in setting the foundation of this thesis.
Also I want to thank to Associate professor Natasa Nord who helped me in the part of
building simulation.
I am thankful to my professors from Macedonia: Slave Armenski, Done Tashevski and Mile
Dimitrovski for their advices, who contributed in writing this thesis.
Last but not the least I want to give special thanks to my family for their support and
encouragement during my stay in Trondheim.
Igor Shesho
i
Abstract
It is well known that building sector in most of the countries consumes more than 40 % from
the primary energy. Also the increased requirements for thermal comfort especially for cooling
involve a considerable consumption of energy. Increased consumption of energy causes rising
environmental issues like increasing the greenhouse gasses emissions/pollution, global warming
etc. Thus the European Union established directive as a common framework of measures for the
promotion of energy efficiency. Due to the implementation of the recommended energy efficiency
measures new buildings have low heating and cooling energy consumption and also the Article 9
from the EU directive requires that “Member States shall ensure that by 31 December 2020 all new
buildings are nearly zero-energy buildings; and after 31 December 2018, new buildings occupied
and owned by public authorities are nearly zero-energy buildings”. This opens the possibility of
efficient harnessing the renewable energies since they are usually utilized at low temperature levels.
The Macedonian renewable energy market regarding the residential heating and cooling
systems mostly is covered by wood pellets, less with heat pumps and the solar energy is mainly
present in solar thermal systems for heating domestic hot water. Therefore the main idea of this
thesis is to analyze the thermal performance of solar assisted air-conditioning systems and their
feasibility for conditions in Macedonia.
Thermal performance of the solar thermal systems are estimated using numerical methods
and software since the solar processes are transitient in nature been driven by time dependent
forcing functions and loads. The system components are defined with mathematical relationships
that describe how components function. They are based on first principles (energy balances, mass
balances, rate equations and equilibrium relationships) at one extreme or empirical curve fits to
operating data from specific machines such as absorption chillers. The component models are
programed i.e. they represent written subroutines which are simultaneously solved with the
executive program. In this thesis for executive program is chosen TRNSYS containing library with
solar thermal system component models.
Validation of the TRNSYS components models is performed i.e. the simulation results are
compared with experimental measurements.
With the simulations are determined the long-term system performance i.e. data are
obtained for the energy consumption, solar fraction, collector efficiency also it is performed
parametric analysis to determine the influence of specific parameters like collector area, tilt and
orientation, mass flow rate etc. to the system performance. In this thesis are considered only the
residential buildings.
ii
Reference solar assisted air-conditioning system is defined for the analysis purposes, which
is numerically modeled in TRNSYS defined with the component models and their interconnections.
The solar system main components are: solar collectors, storage tank, auxiliary heater, absorption
machine reference building as load generator and the hydraulic components.
Analysis starts with assessment the solar thermal system performances when is used only
for providing heat energy for the heating system. In this case as parametric variables are
considered: heating system type (radiator and underfloor ), specific
building heat energy
consumption defined with three building types differencing only in thermal insulation , collector
type, collector area and storage tank volume. Solar fractions are in the range from 8% for radiator
heating system, building type I i.e. specific heat consumption of 70 kWh/m2 a, 16 m2 collector area,
storage tank 1000 l-radiator heating up to 52% solar fraction for underfloor heating system,
building type III i.e. specific heat consumption of 57 kWh/m2 a for 64 m2 collector area, storage
tank 2000 l.
Another analysis is performed for solar assisted cooling system in order to determine the
solar fractions and efficiencies for different collector areas, storage tanks. The obtained results
reveals the solar fraction regarding the ratio between specific collector area : 0,1 m2 / m2conditioned
can cover almost 30% , 0,2 m2 / m2conditioned covers 50% and 0,4 m2 / m2conditioned can cover 70% of
the total required heating energy for driving the absorption chiller. Also is the influence of
installing cold storage tank between the absorption chiller and building cooling system.
At the end is made life cycle cost analysis for the solar assisted air-conditioning systems
with electrical heater or heat pump as auxiliary sources. Solar collector systems applied only for
heating and DHW with heat pump as auxiliary heat source and building with specific heat
consumption of have payback period starting from 7,5 years while the solar assisted cooling system
have negative NPV values which indicates that economically is not profitable measure. Solar
assisted cooling is not feasible since the electricity price in Macedonia is cheap and also the
absorption chiller price is relatively high. But it can provide more than 50 % savings in primary
energy if as conventional is considered the electrical energy.
iii
Nomenclature
and abbreviations
Symbols:
If no other units are mentioned in the text, the following units are used
Qu
Ac
S
UL
Ti
FR
τ
α
IT
IbT
IdsT
IdgT
collector
IAM
θsky
θground
η0
a1
a2
IT
performed
Gb , Gd
c1
c2
c5
Tm
Ta
∆Tdb
γ htr1
U
Cp
COP
Qsurf,I
Qinf,I
Qvent,I
Qg,c,I
Qcplg,I
Qr,wi
Qg,r,i,wi
Qsol,wi
Qlong,I
Qwall-gain
∆∆t
[kJ/h]
- Useful gain from the collector
[m2]
[kJ/h m2]
[kJ/h m2 K]
[K]
[-]
[-]
[-]
[kJ/h m2]
[kJ/h m2]
[kJ/h m2]
[kJ/h m2]
- Collector aperture area
- Solar radiation absorbed per collector unit area
- Overall collector heat loss coefficient
- Fluid inlet temperature
- Collector heat removal factor
- Collector absorber transmittance
- Collector absorber absorbance
- Total radiation incident on the solar collector
- Beam radiation incident on the solar collector
- Sky diffuse radiation on the solar collector
- Ground-reflected diffuse radiation on the solar
[-]
[-]
[-]
[-]
[W/m2 K]
[W/m2 K]
[W/m2]
- incidence angle modifier
- Effective incidence angles for sky diffuse
- Effective incidence angles for and ground reflected
- Optical efficiency (conversion factor)
- Heat transfer coefficient
- Temperature depending heat transfer coefficient
- Solar radiation at which the measurement is
[W/m2 ]
[W/m2K]
[W/m2K]
[W/m2K]
[K]
[°C]
[°C]
[-]
[kJ/kg K]
[kJ/kg K]
[-]
[kJ/h]
[kJ/h]
[kJ/h]
[kJ/h]
[kJ/h]
[kJ/h]
[kJ/h]
[kJ/h]
[kJ/h]
[kJ/h]
[K]
- Beam and diffuse solar irradiance
- Heat transfer coefficient
- Temperature depending heat transfer coefficient
- Temperature depending heat transfer coefficient
- Mean fluid temperature inlet/outlet solar collector
- Ambient temperature where the collector is installed
- Temperature difference dead band controller
- Control signal (0-1) for auxiliary heater
- Overall heat transfer coefficient
- Specific heat capacity
- Coefficient Of Performance
- Convective heat gain from inner surface of zone
- Infiltration gains
- Ventilation gains
- Internal convective
- Gains due to connective air flow
- Radiative gains for the wall surface temperature node
- Radiative airnode internal gains received by wall
- Solar gains through zone windows received by walls,
- Long-wave radiation exchange
- Predfiend heat flow to the wall or window surface.
- Characteristic temperature function
iv
QSH,
[kJ/h]
- Extracted heat energy from the storage tank
[kg/h]
- mass flow rate
⋅
m
Abbreviations
ETC
FPC
BHE
GCHP
PV
DEC
FPC
CPC
SAC
HVAC
DHW
CSHPSS
ROI
SC
PER
- Evacuated tube collector
- Flat plate collector
- Borehole heat exchanger
- Ground coupled heat pump
- Photovoltaic
- Desiccant cooling system
- Flat plate collector
- Compound parabolic collectors
- Solar air-conditioning
- Heating Ventilation Air-conditioning
- Domestic Hot Water
- Central Solar Heating Plants with Seasonal Storage
- Return of investment
- Solar cooling
- Primary Energy Ratio
Sub scripts
in
out
avg
amb
b
ds
dg
D
A
C
E
th
el
sol
- inlet
- outlet
- average
- ambient
- beam radiation
- sky diffuse radiation
- ground reflected radiation
- Desorber
- Absorber
- Condenser
- Evaporator
- Thermal
- Electrical
-Solar
v
Contents
Acknowledgement ........................................................................................................................... i
Abstract ........................................................................................................................................... ii
Nomenclature ..................................................................................................................................... iv
and abbreviations ............................................................................................................................... iv
List of figures ..................................................................................................................................... ix
List of tables ..................................................................................................................................... xii
Chapter 1 ............................................................................................................................................. 1
1.
Introduction ............................................................................................................................ 1
1.1 Background ........................................................................................................................... 2
1.2 Objectives ............................................................................................................................. 3
1.3 Literature review ................................................................................................................... 3
1.4 Structure of the report ........................................................................................................... 6
Chapter 2 ............................................................................................................................................. 8
2. Technologies for active solar thermal systems ........................................................................... 8
2.1 Solar assisted space heating and DHW system .................................................................... 9
2.2 Solar cooling systems ......................................................................................................... 17
2.3 Solar cooling technologies .................................................................................................. 22
Chapter 3 ........................................................................................................................................... 31
3. Numerical modeling of solar thermal system ........................................................................... 31
3.1 Energy simulation software ................................................................................................ 31
3.2 Numerical components models........................................................................................... 34
3.3 Modeling solar thermal assisted air-conditioning system .................................................. 52
Chapter 4 ........................................................................................................................................... 57
4. System and Components validation.......................................................................................... 57
4.1 Solar circuit component validation ..................................................................................... 57
4.2 Absorption chiller validation .............................................................................................. 62
Chapter 5 ........................................................................................................................................... 66
vii
5. Performance evaluation of solar air-conditioning system ........................................................ 66
5.1 Performance indicators ....................................................................................................... 68
Chapter 6 ........................................................................................................................................... 72
6. Simulation results and analysis ................................................................................................. 72
6.1 Reference building modeling and simulation ..................................................................... 72
6.3 Analysis of the solar thermal system in cooling mode for the building ............................. 82
6.4 Techno-economic analysis.................................................................................................. 89
Chapter 7 ......................................................................................................................................... 101
7. Summary and recommendations for further work .................................................................. 101
APPENDICIES ............................................................................................................................... 106
Appendix A................................................................................................................................. 107
Technical data solar thermal storage .......................................................................................... 107
Appendix B ................................................................................................................................. 109
Technical data flat plate solar collector ...................................................................................... 109
Appendix C ................................................................................................................................. 111
Technical data vacuum tube collector ........................................................................................ 111
Appendix D................................................................................................................................. 113
Parameters value for the absorption chillers ............................................................................... 113
Appendix E ................................................................................................................................. 114
Cooling tower performance data ................................................................................................ 114
Appendix F ................................................................................................................................. 115
Appendix G – Draft proposal scientific paper ........................................................................... 116
viii
List of figures
Figure 1. Solar thermal Market in EU27 and Switzerland [14]................................................................ 5
Figure 2. Total and newly installed capacity in EU27 and Switzerland [14] ........................................... 6
Figure 3. Direct collector loop: (a) pump control; (b) three-way valve control ..................................... 11
Figure 4. Indirect collector loop: a) Internal heat exchanger b) plate heat exchanger ........................... 12
Figure 5. Typical solar-thermal system for space heating and DHW (adapted from Beckman et
al.1977)………………………………………………………………………………………………
…13
Figure 6. Solar air heating system (adapted from Beckman et al.1977)................................................. 14
Figure 7. Scheme of drainback space heating and water heating system (gravity return) [15] .............. 14
Figure 8. Schematic diagram of a domestic water-to-air heat pump system .......................................... 15
Figure 9. Schematic diagram of a domestic water-to-water heat pump system ..................................... 16
Figure 10. Schematic diagram of combination solar energy and ground source heat pump .................. 17
Figure 11. Scheme of a solar cooling plant: 1 = solar collector field; 2 = hot storage; 3 = auxiliary
boiler; 4 = absorption chiller; 5 = cold storage; 6 =compression chiller; 7 = cooling towers ................ 18
Figure 12. Simplified scheme of solar cooling plant .............................................................................. 18
Figure 13. Overview on physical ways to convert solar radiation into cooling/air-conditioning [16] ... 19
Figure 14. COP-curves of sorption chillers and the upper thermodynamic limit (ideal) [16]................ 20
Figure 15. Solar electric vapor-compression cooling cycle.................................................................... 21
Figure 16. a) Vapor compression chiller
b) Absorption(thermally driven) chiller......................... 24
Figure 17. Schematic diagram of an single stage absorption chiller (LiBr-H2O) .................................. 25
Figure 18. Simple schematic drawing of an (solar driven) absorption chiller........................................ 29
Figure 19. Transversal and longitudinal directions ................................................................................ 40
Figure 20. Stratified Fluid Storage Tank ................................................................................................ 41
Figure 21. Graphical representation of energy flows into a node .......................................................... 43
Figure 22. Energy balance for he auxiliary heater model ....................................................................... 44
Figure 23. Radiative energy flows considering one wall ....................................................................... 50
ix
Figure 24. Two-node window model used in th TYPE56 energy balance equation .............................. 51
Figure 25. Subsystems definition of solar cooling system ..................................................................... 53
Figure 26. Functional scheme presenting inter conections between components of the system
moddeled in TRNSYS ........................................................................................................................... 53
Figure 27. Scheme of the experimental installation ............................................................................... 57
Figure 28. TRNSYS model for the experimental installation of solar thermal collector system ........... 59
Figure 29. Measured and simulated temperatures for the collector inlet as marked on ......................... 61
Figure 30. Measured and simulated temperatures at the collector outlet ............................................... 61
Figure 31. Measured and simulated temperatures inside storage tank ................................................... 61
Figure 32. Hourly measured and simulated solar radiation for the specific day .................................... 62
Figure 33. Measured ratios of external to internal heat transfer coefficients ......................................... 65
Figure 34. Measured cooling capacity for design and variable flow rates, by Kühn et al.
2005&2007……………………………………………………………………………………………
..65
Figure 35. Solar cooling system variables [44] ...................................................................................... 67
Figure 36. Monthly energy consumption for the three building “types” ................................................ 74
Figure 37. Hourly dry bulb ambient temperatures for Skopje, R.Macedonia ........................................ 74
Figure 38. Cross section of the active layer for the underfloor heating ................................................. 76
Figure 39. Solar fraction for Building I, radiator and underfloor heating in regard of collector array
area and strorage volume ........................................................................................................................ 80
Figure 40. Solar fraction for heating with storage tank 1500l regard of collector array area building
type ......................................................................................................................................................... 81
Figure 41. Monthly average zone temperature, only heating analysed .................................................. 81
Figure 42. Scheme of the collector array conection 16/2 ....................................................................... 85
Figure 43. Monthly values of solar fraction for different tilt angles and tracking azimuth for solar
assited cooling system ............................................................................................................................ 86
Figure 44. Solar fraction and fan cooling tower energy consumption for system with and without
storage tank for solar assited cooling system ......................................................................................... 87
Figure 45. Monthly average storage tank temperture and solar collector yield in regard of system
with/without cold storage tank for solar assited cooling system ............................................................ 87
Figure 46. Solar fraction and thermal efficiency in regard of collector area for solar assited cooling
system ..................................................................................................................................................... 88
Figure 47. Hourly heating loads and useful heat yield for different collector array areas and constat
storage tank of 1000 l and one case for 64 m2 with 2000 l ..................................................................... 91
Figure 48. Hourly heating loads and collector energy yield for different areas ten day period ............. 91
x
Figure 49. Payback time for solar thermal systems for two building types regarding the specific
heat energy consumption ........................................................................................................................ 98
Figure 50. Primary energy consumption for solar thermal systems in regard of two buildings with
different specific heat energy consumption 57 kWh/m2 a and 70 kWh/m2 a ......................................... 99
Figure 51. Specific absorption chiller costs per kW cooling power [48] ............................................... 99
xi
List of tables
Table 1. Technical data for collector type Camel Solar Flat plate 2.0-4 ................................................ 55
Table 2. Technical data for collector type Camel Solar Vacumm tube SC 15 ....................................... 58
Table 3. Storage tank technical details ................................................................................................... 59
Table 4. Technical data for different market available small absorption chillers .................................. 63
Table 5. Averaged measured performance data by Kühn (2005) of a 10kW absorption chiller ............ 64
Table 6. Energy and thermal flows of solar cooling systems ................................................................. 67
Table 7. Legend of the performance indicators ...................................................................................... 69
Table 8. Calculation procedures for solar thermal system performance indicators................................ 71
Table 9. Reference building physical and thermal performance data .................................................... 73
Table 10. Active layer TYPE data .......................................................................................................... 75
Table 11. Parametric analysis of solar assited heating system with radiators ........................................ 78
Table 12. Parametric analysis of solar assited underfloor heating system ............................................. 79
Table 13. Cooling tower design parameters ........................................................................................... 83
Table 14. Monthly average solar fractions and efficiency in regard of collector areaa and storage
volume .................................................................................................................................................... 84
Table 15. Monthly avrage solar fractions and thermal efficiency in regard of colelctor array
interconnections ...................................................................................................................................... 84
Table 16. Solar fractions and thermal efficiency for solar assited cooling system in regard of
collector orientation i.e. azimuth ............................................................................................................ 85
Table 17. Average solar fraction and thermal efficiency of solar cooling system in regard of
collector area and strage tank volume for solar assited cooling system ................................................. 88
Table 18. Solar fraction and thermal efficiency for different specific vacuum collector areas storage
volumes for solar assited cooling system ............................................................................................... 88
Table 19. Annual energy balance, energy and system costs, CO2 emissions for conventional heat
source systems obtained with simulations, specific heat energy consumption 70 kWh/m2 a ................ 92
Table 20. Solar thermal system energy performance, costs, parameters and CO2 emissions, primary
energy consumption for building with specific heat energy consumption 70 kWh/m2 a, – part I ........ 93
xii
Table 21. Solar thermal system energy performance, costs, parameters and CO2 emissions, primary
energy consumption for building with specific heat energy consumption 70 kWh/m2 a, – part II ....... 93
Table 22. Solar thermal system energy performance, costs, parameters and CO2 emissions, primary
energy consumption for building with specific heat energy consumption 57 kWh/m2 a, – part I ........ 94
Table 23. Solar thermal system energy performance, costs, parameters and CO2 emissions, primary
energy consumption for building with specific heat energy consumption 57 kWh/m2 a, – part II ....... 94
Table 24. LCCA for solar thermal system assited buildign heating with specific .heat energy
consumption 70 kWh/m2 a...................................................................................................................... 97
Table 25. LCCA for solar thermal system assited buildign heating with specific .heat energy
consumption 57 kWh/m2 a...................................................................................................................... 98
Table 26. Solar thermal system energy performance, costs, parameters and CO2 emissions, primary
energy consumption for building with specific cooling energy consumption 12 kWh/m2 a................ 100
xiii
Chapter 1
1. Introduction
In the past twenty years the uncontrolled or mildly said “demanded” industry development
accompanied by the increased thermal comfort demand and the limited energy resources naturally
activated and imposed the need of the forgotten term of energy efficiency. In the past, the
cheap/affordable energy sources, high building energy consumption and the modest technology
development were the main limiting factors in the viability of utilization renewable energy sources.
Undertaken measures started with limiting the energy consumption through different directives
and regulations which translated into actions meant: improving efficiencies of existing systems,
decreasing the energy demand, up to developing new technologies, and all of that with one purpose
- doing more with less. Now with the implementation of the buildings directives such as the Energy
Performance Building Directive 2002/91/EC the EU 2020 strategy, technology development
enabled the renewable energy sources to be feasible for implementation in buildings airconditioning systems.
From the climatic point of view, the World Meteorological Organization - WMO (2012) has
determined that the “long term warming trend” continues, being the period of 2001 – 2011 the
world’s warmest decade since 1850. This trend has a significant impact in the energy sector, both in
the generation and the demand sides. Regard to the demand side, various studies has been
conducted, in particular with the change in thermal energy loads in buildings; a number of them
have been summarized by Yau and Pean (2011). Although the studies presented in the mentioned
review follow different methodologies, it is noticeable an increase tendency in energy demand due
to the effects of climate change. Considering the above, the integration of renewable energy sources
is a valuable strategy to reduce and/or offset the increase in cooling load; especially with the use of
active solar systems, because cooling loads are in phase with the amount of solar irradiation.
A significant contribution to the primary energy consumption of first and second world
countries is being made by the rapidly increasing use of electrical air-conditioning units worldwide.
1
Chapter 1. Introduction
Worldwide sales of room air-conditioners of all types amount to approximately 50 Mio. units p.a.,
with the U.S. China and Japan being the three main markets. In Europe, commercial air
conditioning has a share of 4% of the total annual electricity consumption while residential air
conditioning accounts only for 0.4%. Although the latter number is still comparatively low, Europe
has seen a seven-fold increase of residential air-conditioning sales between 1990 and 2004. The
reasons for the growing use of air-conditioning are twofold. First, the comfort demands from both
building users and owners have increased. The standard of living of the present generation is higher
than in the past, especially in private buildings. . Second, the trend towards commercial buildings
with large glazed facades has increased the internal heat load to be removed by air-conditioning.
Third, electricity prices are comparatively low. The additional cost caused by the use of airconditioning units is not in the order of magnitude to influence the consumer behavior significantly.
The obvious consequence of this growing air-conditioning use is increased power consumption.
Outside Europe, another consequence of excessive air-conditioning are locally higher temperatures
in metropolitan areas, commonly referred to as heat islands. As a part, these inner-city temperature
peaks are the result of heat conveyed from building inside to outside, released at a temperature
level above ambient temperature. Both these consequences are strong arguments for alternative airconditioning or cooling methods.
Priority is given to buildings and transport sector. Through directives to improve building
energy efficiency (energy performance of buildings directive or EPBD), the European Union
recognized high potential for energy savings from buildings and promote the installation of solar
thermal systems in the building sector. Solar technologies can supply the energy for all of the
building’s needs—heating, cooling, hot water, light and electricity—without the harmful effects of
greenhouse gas emissions created by fossil fuels thus solar applications can be used almost
anywhere in the world and are appropriate for all building types. The heat energy demand for
heating the building and/or DHW determines the solar collectors area which often can exceed the
available optimal area for installation of the collectors. Thus collectors are connected in arrays
which opens a variety of combinations regarding the number of collectors hydraulics and layout.
1.1 Background
It is well known that according the IEA(International Energy Agency) buildings represents 32%
of the total final energy consumption and converted in terms of primary energy this will be around
40%. Inspected deeper, the heating energy consumption represents over 60% of the total energy
demand in the building. Space heating and hot water heating account for over 75% of the energy
used in single and multi-family homes.
2
Chapter 1. Introduction
Solar technologies can supply the energy for all of the building’s needs—heating, cooling,
hot water, light and electricity—without the harmful effects of greenhouse gas emissions created by
fossil fuels. Usually the maximum demand for cooling coincides with the maximum availability of
solar radiation, whereas conventional electrical-compressor chillers have the problem of providing
their minimum capacity in the hottest hours.
As mentioned before in 2010 the European commission adopted communication “Energy 2020”
that defines new strategy toward 2020 for a competitive, sustainable and secure energy [1].
Without a doubt, the European goal of covering 20% of energy needs with renewable energy
can only be reached with a significant increase of renewable energy capacities in the heating sector.
The explosion of crude oil and natural gas prices along with increasing import dependency have
further increased public attention and interest.
1.2 Objectives
In this work/thesis the main accent is given on solar energy. The main objective is assessment
of thermal performance of the solar driven air-conditioning systems through analysis of different
innovative combination of solar heating, cooling and storage techniques in regard of primary
energy savings and economic feasibility compared to pre-defined reference system for the weather
conditions in R.Macedonia. Objective will be achieved through transient dynamic simulations and a
holistic approach for various configurations between solar heating and cooling technologies, energy
storage and auxiliary conventional heating and cooling devices. Representative building with
different specific heat and cooling energy will be defined in order to analyze the performance of the
solar assisted air-conditioning systems.
1.3 Literature review
Solar thermal systems for hot water production are already mandatory in new buildings
according to solar ordinances in Spain [2], Portugal, Italy, Greece and other European countries [3].
Systems combining production of domestic hot water (DHW) and space heating systems are
well suited to middle and high latitudes, due to significantly higher solar radiation in the
transitional period around winter (September-October and March-May) and the significant heating
demand in these latitudes at that time [4].
Installations with large solar collector areas and small size heat storage capacity can cover
around 50% of the total heat demand. This percentage can be higher in some cases of large storage
capacities and primary energy savings up to 80% [5]. Simulations of central solar heating plants
with seasonal storage (CSHPSS) have shown that the solar fraction of such systems varies between
3
Chapter 1. Introduction
50% and 100% [6, 7]. The heat produced by the collectors may be stored in thermal energy
storages in order to provide domestic hot water and space heating when required [8].
Solar air conditioning refers to any air conditioning system that uses solar power. This can be
done through solar thermal energy conversion, passive solar and photovoltaic conversion.
The addition of a solar cooling facility makes the system complete, covering all building
thermal and cooling demands. These systems which are combination of solar heating and cooling
seem to be proved advantageous since high cooling load coincident with high solar radiation and
consequently the readily available solar energy from the existing solar collector can be exploited by
a heat driven machine. The use of the solar collector is then extended to a whole year, making the
system financially attractive, when the number of annual full-load hours is high [9]. Furthermore
since the cooling machine is heat driven, the building electrical load share reduce the problems
associated with peak power demand during summer are minimized. Depending on the size of solar
collector field, hot water storage, local climatic conditions and building loads, these solar driven
systems can cover 10-60% of the combined space heating/cooling and DHW demand at southern,
northern and central European countries [10].
Urban areas are of particular interest where adverse outdoor conditions, as a result of higher
outdoor pollution and the urban heat island effect, encourage the use of mechanical air-conditioning
with a direct impact on peak electrical energy use [11]. Application of the solar energy for airconditioning in the commercial sector can be said that is relatively new. In the study of Lamp and
Ziegler [12] they give an overview of the European research on solar-assisted air conditioning up to
1996. It is investigated that the application of renewable technologies in the European tourism
industry and identified a large number of solar thermal systems but only a few solar cooling
systems [13].
Regarding the large scale application as main obstacles, beside the high first cost, also there is
the lack of practical experience and acquaintance among engineers with the design, control and
operation of these systems. For smaller scale systems, there is no market available technology.
Therefore, the development of low power cooling and air conditioning systems is of particular
interest. Heat-driven cooling technologies include mainly closed cycles (absorption, adsorption)
and open cycles (desiccant systems which will be detail described in the next chapter.
Despite their ecological advantages, solar cooling systems also have to yield an economic
advantage for the customer. At present, investment costs are higher for solar cooling systems than
for comparable compressor based cooling units. Thus, the full potential of solar cooling is far from
being realized, however building owners, occupants and architects are becoming more and more
sensitive towards energy issues. The economic advantage of solar cooling systems results from
4
Chapter 1. Introduction
much lower operation costs which include the costs for power, water and maintenance. Especially
the electrical power consumption of a solar cooling system influences the economics strongly. The
main idea of such a system is to use thermal energy for most of the process work, thus the
remaining power consumption should be kept as low as possible.
1.3.1 Solar thermal market in Europe
In this part are presented the results from the report of the European Solar Thermal Industry
federation for the current situation of the solar thermal market in Europe i.e. it is given overview
for ten year trend from 2002 – 2012.
The European Union market continues to suffer from the constraints imposed by the financial
and economic crises affecting most of the continent, resulting in a sluggish construction sector and
reduction of public support schemes for solar thermal. The annual market has been contracting
since the peak year of 2008. The 2.41 GWth sold in 2012 are well above the 2007 sales (2 GWth /
2.88 mio m²) but are a far cry from the 3.36 GWth (4.8 mio m²) reached in 2008. Over the past ten
years, there was a continuous steep uptrend in the growth rate up to 2008; followed by a decline,
steeper in the
first two years (2009, 2010) and then flattening out (2011, 2012). The variation in
the newly installed capacity is illustrated with the blue line in the graph on the right. In spite of the
decrease recorded over the last four years, the annual market size has doubled, over the past decade
at an average annual growth rate of 10%. Using the same comparison over the last five years (20072012), we can observe an absolute growth in the annual sales of 20% and an average annual growth
rate of 3.6%
Figure 1. Residential applications still represent the bulk of the solar thermal market.
Nevertheless, large installations are increasing apace. Large size systems above 35 kWth (50 m²)
for commercial heating and cooling applications have shown a positive development, but it is
mainly for very large systems (above 350 kW / 500 m²) that the market has been moving rapidly.
th
Figure 1. Solar thermal Market in EU27 and Switzerland [14]
5
Chapter 1. Introduction
In 2012 confirmed Denmark as the land of large solar district heating, with a total of 71.4
MWth (102 000 m²) installed, contributing to a total installed capacity of 196 MWth (280 000 m²),
solely in large solar thermal plants, that account for 65% of the European total installed capacity in
large systems, Figure 2 with regard to industrial process heat, several pilot projects have been
implemented, with other large ones in the pipeline. This is clearly a market segment to watch
closely in the coming years.
Figure 2. Total and newly installed capacity in EU27 and Switzerland [14]
Despite a below expectation growth of the total installed capacity (the evolution is shown by
the grey bars in the graph referred to above, the lighter shade of grey indicates the increment from
the newly installed capacity in 2012), solar thermal plays an increasingly important role in the
European energy strategy, namely through the National Renewable Energy Action Plans. The 28
GWth in operation generate an estimated 20 TWhth of solar thermal energy while contributing to a
saving of 2.5 Mt CO2 [14].
1.4 Structure of the report
The rest of the report is structured as follows. Chapter 2 gives overview for the existing
technologies for active solar systems. First is given general description of the solar assisted space
heating and DHW systems. Further are elaborated several physical configurations of active solar
heating systems and combinations with heat pumps. Second part from this chapter continues with
solar cooling technologies and components with special accent on the absorption, adsorption
chillers and desiccant cooling technologies, thermodynamic limitations and performances.
Chapter 3 concerns to the separate component models used in the simulation of the solar
assisted air-conditioning system like solar collector, storage tank, auxiliary heater, absorption
chiller, hydraulics etc. Simply said it is given overview of the mathematical relations which
6
Chapter 1. Introduction
describe how component models function. The elaborated component models are contained within
TRNSYS and Tess library.
Chapter 4 is dedicated to procedure of validation for the previously mentioned component
models in Chapter 3. Experimental set up of the system or component are modeled in TRNSYS and
simulation results are compared with the measured ones.
Chapter 5 deals with performance evaluation of solar air-conditioning system. There are given
performance indicators and procedures for calculation of system thermal performance.
In Chapter 6 is described the numerical development of the simulated system i.e. component
models characteristics with their interconnections, input data and simulation boundaries. Further
continues with developing different scenarios subject for simulations and analysis. Each of the
simulation results are subject of analysis in regard of which are drawn conclusions and
recommendations.
In the last Chapter 7 are condensed all of the results, analysis, conclusions covered within
previous chapters. At the end are given recommendation for further work. At the end are given the
Appendices.
7
Chapter 2
2. Technologies for active solar thermal systems
The term “active solar systems” is commonly applied to assemblies of equipment that provide
heating, cooling or hot water in dwellings and commercial buildings.
The distinction between active and passive systems is sometimes made according to other
criteria, as follows. Auxiliary energy is generally required to make the transfer fluid circulate in
active systems (forced circulation), whereas natural circulation (of air) takes place in passive
buildings. System operation is managed by a specific control device in active solutions, whereas
passive systems are more or less self-regulating.
Usually excluded is equipment that supplies electricity or products of biological processes. Also
excluded are materials and designs that by permitting the passive supply of solar energy to a
building, usually through windows or other transparent and translucent surfaces, can reduce the
demand for conventional energy.
Active solar space systems use collectors to heat a fluid, storage units to store solar energy until
needed and distribution equipment to provide the solar energy to the heated spaces in a controlled
manner. Additionally complete system includes pumps or fans for transferring the energy to storage
or to the load; these require a continuous availability of non-renewable energy, generally in the
form of electricity.
The load can be space cooling, heating or combination of these two with hot water supply.
When is combined with conventional heating equipment solar heating provides the same level of
comfort, temperature stability and reliability like conventional system.
It will be separately described the technology of solar space heating and cooling processes. In
further analysis will be used combination of solar space heating and cooling system with DHW
system.
8
Chapter 2. Technologies for active thermal solar systems
2.1 Solar assisted space heating and DHW system
A solar system for space and water heating can be designed to suit any particular application,
residential or commercial, new or retrofit. It is technically feasible to design a solar system that can
provide 100% of heating needs of a building but generally it is uneconomical to do so. Practical
solar heating systems are designed to displace up to 50% of conventional fuel needs and require
auxiliary heating systems that are fully capable of supplying the total heating load when no solar
energy is being collected and when stored solar energy has been depleted.
When space heating systems serve occupied buildings it is economical to include heating of
domestic hot water (DHW) in the system. For residential systems during the summer when there is
no space heating load, the entire solar system can be devoted to water heating so that solar can
supply a substantial portion if not all of the DHW heating needs. However for large commercial
systems summer operation to provide a relatively small DHW demand may not be worthwhile.
The solar loop also known as the collector loop, comprises the components which collect solar
energy (electromagnetic radiation from the sun), convert it to heat and transfer the heat to the
storage. The basic component of the solar loop is a set of solar collectors through which a heattransfer fluid circulates. This fluid is generally water or an anti-freeze solution, such as monoethylene glycol + water, but other fluids (e.g. air or an organic thermal fluid) can also be used.
Several types of collectors exist on the market. The most common for low-temperature applications
are flat-plate solar collectors , however other types can also be used, such as evacuated tubular
collectors (ETC).
The solar loop is completed by a piping circuit, a circulation pump, and several safety and
maintenance devices (e.g., an expansion tank, a filter, an air bleed, a safety valve, and a non-return
valve). The heat from the solar loop is usually transferred to a heat storage device, either directly or
through a heat-exchanger when different fluids are used in the collectors and in the storage. The
solar loop is controlled by a differential thermostat which switches the pump on and off to ensure
that fluid circulates only if a net gain of energy is possible.
Storage is required because energy demand and solar availability rarely coincide. Storage is
generally as sensible heat (sometimes latent heat) and the most common form is an insulated water
tank. A thermal stratification is observed in storage tanks: hot water, lighter than cold water, rises
to the top of the tank. In well-designed systems, this physical phenomenon is enhanced to improve
the performance. The connection with the load in the distribution loop varies according to the
application.
9
Chapter 2. Technologies for active thermal solar systems
The size of a solar system (primarily the collector area and storage volume) for a particular
building depends on the portion of the total load the system is expected to provide. Size also
depends on climate and location. The type of the system whether air or liquid based depends in part
on the application and to a large extend on the designers choice. Very often large systems may be
subdivided into several small systems to better suit the application. Smaller systems may be easier
to control, provide better overall efficiency and cost collectively less than a system with one large
collector array and storage.
Solar heating systems are as adaptable technically for commercial applications as for
residential applications, except that for very large buildings, available area for collector placement
may constrain system size. For most residential applications, selecting an air- or liquid based
system is largely a matter of designer or owner preferences. Air-based solar system can be as
effective as liquid-based systems, and costs for equally sized systems are essentially the same. Heat
transport in large air ducts is more expensive than a liquid through pipes.
Regarding new building or retrofit, there is considerable freedom in selecting a solar system
type for a new building and in designing that system to provide an arbitrary (but economical)
fraction of the total load. Orientation of the building, slope of the roof for mounting collectors,
location and space thermal energy storage and type of heat distribution system can be selected
according to system choice. Options are fewer for retrofit application. Building orientation is fixed,
roof slope may not be suitable for collector mounting, choice of storage location and size may be
limited and some interior changes may be necessary to accommodate the heat transport subsystem.
The existing heating and distribution system may also restrict the selection of solar system type and
design.
The size of the solar system is logically based on economics. Among the economic figures of
merit are total capital investment (first cost), least cost for energy, life cycle cost, life cycle cost
savings, payback, and return of investment (ROI). While positive life-cycle cost savings and shortterm payback are criteria most readily understood by homeowners a high ROI is usually the
criterion for commercial systems.
While other, noneconomic factors can affect choice of system size for residential applications,
using solar systems that can provide from 30% to perhaps 50% of the total annual heating
requirements is a practical choice. Sizes of commercial systems, while also dictated by economic
constrains also depend on availability of suitable locations for collectors.
Solar radiation availability and heating demand are inversely related; in sunny winter climates,
the need for space heating is low while in areas with prevailing clouds in winter, temperatures are
generally low and heating demand is high. It is commonly assumed that solar heating systems are
10
Chapter 2. Technologies for active thermal solar systems
therefore suitable for climates between these two extremes. While solar radiation is least during the
coldest periods of winter the fall and spring periods may provide substantial opportunity for
displacement of conventional heating fuels.
2.1.1 Physical configurations of active solar heating systems
Although it is possible to imagine a large variety of active solar systems (even with a simple
application such as space or water heating), there are far fewer possibilities for the solar loop itself.
A small number of basic designs exist, with only a few variants, and these are independent of the
application. A distinction can be made between a solar loop without a heat exchanger (direct loop),
and one with a heat-exchanger (indirect loop).
In this simple configuration of direct loop Figure 3(a), solar heat from the collector is
transferred directly to the storage (or to the distribution loop when there is no storage in the
system). The same fluid circulates throughout, consequently the direct loop is unsuitable when antifreeze fluid is required in the collector. The direct loop is usually controlled by a differential
thermostat (between the collector outlet and the bottom of the storage tank). This thermostat
controls the circulation pump and lets it operate only if the temperature difference between the two
sensors is greater than some set value.
Figure 3. Direct collector loop: (a) pump control; (b) three-way valve control
In a variant of this system Figure 3 (b), the pump operates continuously during daylight hours
and the differential thermostat controls a three-way valve. These two types of control are roughly
equivalent in terms of thermal performance. Therefore the former type (the simplest) is often
preferred. Nevertheless, it is preferable to avoid starting and stopping the pump too frequently.
Frost protection of a collector loop is usually achieved by the use of an anti-freeze solution, but
this is rarely used as a storage fluid owing to its cost. A heat-exchanger is therefore required
between the collector loop and the storage. The simplest case is that of a heat-exchanger submerged
within the storage tank, most often in the lower part. This design is often used for single-family
11
Chapter 2. Technologies for active thermal solar systems
solar water heaters. The control strategies mentioned above still apply here with similar sensor
locations. As a general rule, this configuration does not allow the storage to remain stratified when
the collector operates, thus low flow-rates are not suitable.
Another combination is with indirect loop with external heat exchanger i.e. plate type. In this
loop, the primary side of a tubular or plate heat-exchanger is inserted in the collector loop, and the
secondary side is connected to the storage Figure 4.
A secondary pump is needed and the control is more sophisticated. One simple configuration
has the primary pump PI working continuously (during daylight hours), while a differential
thermostat between the collector outlet and the bottom of the storage controls the secondary pump
P2. Another possible solution for large collector arrays has been proposed. This solution
incorporates three temperature sensors, one at the collector outlet (SO), one at the heat exchanger
inlet (collector loop, SO1), and one at the bottom of the storage (SI). The control operates the
primary pump PI as soon as TSO > TS1; after a delay, the pump is then controlled by the sensor
SO1 which replaces SO. The primary pump goes on working only if TSO' > TS1; in this case, the
secondary pump P2 is also switched on. For large plants, many sophisticated control strategies have
been proposed in the past, including attempts at optimum control of several storage tanks at
different temperatures. In most cases however, the practical operating problems outweighed the
improvement in performance.
Figure 4. Indirect collector loop: a) Internal heat exchanger b) plate heat exchanger
On the Figure 5 is presented schematic diagram of a typical space heating system. The system
consists of three loops-collector, storage and load. In addition as previously mentioned, most spaceheating systems are integrated with a domestic water-heating system to improve the yearlong solar
load factor.
Since space heating is relatively low-temperature use of solar energy, a thermodynamic match
of collector to task indicates that an efficient flat plate collector or low-concentration solar collector
is the device of choice.
12
Chapter 2. Technologies for active thermal solar systems
Figure 5. Typical solar-thermal system for space heating and DHW (adapted from
Beckman et al.1977)
The collector fluid loop contains fluid manifolds, the collectors the collector pump and heat
exchanger, an expansion tank and other subsidiary components. A collector heat exchanger and
antifreeze in the collector loop are normally used in all solar space heating systems since the
existence of a significant demand implies the existence of some subfreezing weather.
The storage loop contains the storage tank and pump as well as the tube side of the collector
heat exchanger. To capitalize on whatever stratification may exists in the storage tank, fluid
entering the collector heat exchanger is generally removed from the bottom of the storage. This
strategy ensures that the lowest temperature fluid available in the collector loop is introduced at the
collector inlet for high efficiency.
On the Figure 6 shows assumed configuration for a solar air heater with pebble-bed storage
unit. Energy for domestic hot water is provided by heat exchange from air leaving the collector to a
domestic water preheat tank as in the liquid system. The hot water is further heated if necessary by
a conventional water heater. During summer operation, a seasonal, manually operated storage
bypass damper is used to avoid heat loss from the hot bed into the building.
The standard solar domestic water heater collector heats either air or liquid. Collected energy is
transferred by a heat exchanger to a domestic water preheat tank that supplies solar-heated water to
a conventional water heater. The water is further heated to the desired temperature by conventional
fuel if necessary.
13
Chapter 2. Technologies for active thermal solar systems
Figure 6. Solar air heating system (adapted from Beckman et al.1977)
Another possible system for combined space heating and DHW is the drainback solar system.
In this case, a large atmospheric pressure storage tank is used from which water is pumped to the
collector by pump P1 in response to the differential thermostat T1. Drainback is used to prevent
freezing because the amount of antifreeze required would be prohibitively expensive. Service hot
water is obtained by placing a heat exchanger coil in the tank near the top, where even if
stratification occurs the hottest water will be found.
Standby heat becomes increasingly important as heating requirements increase. The heating
load winter availability of solar radiation and cost and availability of the auxiliary energy must be
determined. It is rarely cost-effective to do the entire heating job for either space or service hot
water by using the solar heat collection and storage system alone. On the Figure 7 is presented solar
drainback system for space heating and DHW. The space heating is provided by heat exchanger
water-air placed in the ventilation duct of the building.
The storage tank has two heat-exchangers for indirect fluid heating for the space heating which
is serial connected with the auxiliary heater and another for the DHW.
Figure 7. Scheme of drainback space heating and water heating system (gravity return) [15]
14
Chapter 2. Technologies for active thermal solar systems
A disadvantage of system depicted in Figure 7 is that water must be circulated against full static
head losses as well as friction head losses in the supply piping and through the collector.
2.1.2 Solar energy and heat pumps
The solar energy can be used as supplementary thermal energy for the heat source of the heat
pumps. It can be used to preheat the air entering the heat pump evaporator or regenerate the ground
temperature for the ground source heat pumps.
Heat pumps are usually vapor compression refrigeration machines, where the evaporator can
take heat into the system at low temperatures and the condenser can reject heat from the system at
high temperatures.
Heat pumps have been used in combination with solar systems in residential and commercial
applications. The additional complexity imposed by such a system and extra costs are offset by the
high coefficient of performance and the lower operating temperature of the collector subsystem. A
schematic of common residential heat pump system is shown in Figure 8.
Figure 8. Schematic diagram of a domestic water-to-air heat pump system
During favorable weather conditions it is possible with this arrangement to have solar energy
delivered directly to the forced air system while the heat pump is kept off. The arrangement shown
on Figure 8 is a series configuration where the heat pump evaporator is supplied with energy from
the solar system. Energy from collector system is supplied directly to the building when the
temperature of the water in the storage temperature is high. When the storage temperature cannot
satisfy the load the heat pump is operated thus it benefits from relatively high temperature of the
solar energy system which is higher than the ambient and thus increases the heat pump’s COP.
A parallel arrangement is also possible where the heat pump serves as an independent auxiliary
energy source for the solar energy system as shown in Figure 9. In this case a water-water heat
pump is used.
15
Chapter 2. Technologies for active thermal solar systems
Figure 9. Schematic diagram of a domestic water-to-water heat pump system
The series connection is usually preferred because it allows all the solar collector power to be
used, leaving the tank at low temperature which allows the solar energy system to work more
efficiently the next day.
At the ground source heat pumps most embedded heat exchangers for private residences are
installed either horizontally or vertically in the ground. Interest in vertical heat exchangers, also
called borehole heat exchangers (BHE), has increased in the housing sector over the last decade
because they offer better performances. Indeed, horizontal heat exchangers are directly affected by
local climatic conditions as they are buried at depths between 0.80 and 1.50 m, while boreholes can
exploit the ground temperature regularity below 6 m in depth, which ensures good performance
throughout the year whatever the local climatic conditions. However, it should be pointed out that
the high cost of boreholes is the major drawback of BHE systems, as installation requires drilling
technologies. Nevertheless, the soil surface area occupied by a BHE is very small compared to that
occupied by a horizontal ground heat exchanger, an advantage in areas of high land prices.
Nevertheless, the use of a geothermal heat pump (GHP) with BHE to heat and/or to cool
buildings can create annual imbalances in the ground loads. In the case of heating dominated
buildings, a thermal heat depletion of the soil can occur, which progressively decreases the heat
pump’s entering fluid temperature. On the contrary, cooling-dominated buildings heat the soil,
which progressively increases the heat pump’s entering fluid temperature. As a consequence, the
heat pump’s performance coefficient decreases and the installation gradually becomes less
efficient. One solution for this problem could be combining solar collectors and the Ground
Coupled Heat Pump (GCHP). This type of system has been increasingly recognized since the oil
crisis in the 1970s , but the technology has not been widely adopted. In addition, experimental and
theoretical results on the combination of thermal solar collectors with a GCHP used in heating
dominated detached houses are relatively scarce.
The schematic diagram of the combination solar energy with ground source heat pump is
presented on Figure 10. Solar heat is used in priority to heat DHW and is injected into the ground
16
Chapter 2. Technologies for active thermal solar systems
via boreholes only when the DHW temperature setting is reached. The advantage of this operation
is that it contributes to the balance of the ground loads, optimizes the use of solar heat provided by
solar collectors and prevents overheating problems. The heat pump can be used in heating mode or
in cooling mode. In cooling mode, heat is injected into the ground which also contributes to the
balance of the ground loads.
Figure 10. Schematic diagram of combination solar energy and ground source heat pump
It should be noted that regarding the control strategy compared to conventional heating or
cooling, system controlling this installation is relatively complex. The power provided by
geothermal energy is nearly the same throughout the year, as opposed to solar energy for which the
power provided depends on solar radiations. To ensure proper operation of the installation, it is
simpler to use two existing control systems: one adapted to ground-coupled heat pump systems and
another adapted for solar heating systems. However, these two control systems must have good
operational flexibility in order to ensure that the GCHP system combines well with the solar
collector.
2.2 Solar cooling systems
In 2011, worldwide, about 750 solar cooling systems were installed, including installations with
small capacity (<20kW) (Mugnier and Jakob, 2012). Recently a number of very large installations
have been completed or are under construction. Examples are the system at the headquarters of the
CGD bank in Lisbon, Portugal with a cooling capacity of 400 kW and a collector field of 1 560m²;
and the system installed at the United World College in Singapore, completed in 2011, with a
cooling capacity of 1 470 kW and a collector field of 3 900m².
17
Chapter 2. Technologies for active thermal solar systems
Solar cooling of buildings is an attractive idea. Cooling is important in space conditioning of
most buildings in warm climates and in large buildings in cooler climates. The biggest advantage is
that cooling loads and solar availability are approximately in phase. The combination of solar
cooling and heating should greatly improve use factors on collectors compared to heating alone.
Solar air conditioning can be accomplished by three classes of systems: absorption cycles,
desiccant cycles and solar mechanical processes. In
Figure 11 is presented general scheme of solar assited absorption cooling system. The cooling
demand can be satisfied by the absorption chiller, the compression chiller or the cold storage. The
thermal power needed by the absorption chiller is supplied by the solar collector field, the auxiliary
heater or the hot storage. Cooling towers for the chillers heat rejection complete the scheme.
Figure 11. Scheme of a solar cooling plant: 1 = solar collector field; 2 = hot storage; 3 =
auxiliary boiler; 4 = absorption chiller; 5 = cold storage; 6 =compression chiller; 7 = cooling towers
In order to reduce the plant complexity (and the cost of the system) some simplified
configurations can be drawn. In the simplest solar cooling plant presented onFigure 12, (1 + 4 + 7)
the entire thermal load is satisfied by a totally solar driven absorption chiller. In this case, the
absorption chiller has to be sized with respect to the pick demand (QE ≥ QL,max). The solar collector
field must be oversized, because it has to drive the chiller also during the periods of low irradiation
(QColl ≥ QG).
Figure 12. Simplified scheme of solar cooling plant
In order to avoid wasting the heat collected, a hot storage can be inserted between the solar field
and the absorption chiller (1 + 2 + 4 + 7). The storage capacity must be evaluated by matching the
18
Chapter 2. Technologies for active thermal solar systems
thermal power produced by the collectors and the chiller heat demand. When QColl > QG , the heat
surplus is stored in the tank; when QColl < QG, the hot storage is needed to drive the absorption
chiller: this situation typically occurs at the beginning and at the end of the day, when the solar
radiation level is not high enough to drive the chiller at its lowest capacity. The collector area is
partially reduced with respect to the previous case: in fact, when the irradiation is low, the hot
storage can contribute to supply the chiller generator. To significantly reduce the collector area, an
auxiliary boiler could be introduced (1 + 3 + 4 + 7). In this case, a fraction of the energy input to
the generator can be supplied by the burner. The heat collected by the solar devices is wasted for
few time, when the building load is low and the solar radiation is high (QColl > QG).
2.3.1 Thermodynamic schemes and limits for solar cooling systems
From a thermodynamic point of view there are many processes conceivable for the
transformation of solar radiation in cooling. Solar cooling technologies can be broadly classified
according to Figure 13.
. Although the conversion of electricity by photovoltaic and the subsequent use of this
electricity in a classical motor driven vapor compression chiller is a technically feasible concept, it
is not further considered here. Reason is that in industrialized countries, which have a welldeveloped electricity grid, the maximum use of photovoltaic is achieved by feeding the produced
electricity into the public grid. From an economic point of view this is even more valid if the price
for electricity generated by solar energy is higher than that of electricity from conventional sources
(e.g., feed-in laws in Germany or Spain). From the thermally driven technologies, which may use a
solar thermal collector to provide heat to drive a cooling process, the technologies based on heat
transformation are best developed. Therefore only these technologies are considered further.
Figure 13. Overview on physical ways to convert solar radiation into cooling/air-conditioning
[16]
19
Chapter 2. Technologies for active thermal solar systems
Figure 14. COP-curves of sorption chillers and the upper thermodynamic limit (ideal) [16]
From the above presented technologies this report will concentrate on those for small to scale
A/C applications, which are currently available in the market.
In contrast, solar cooling systems based on the Rankine cycle have not been subject of further
study, due to their complexity and higher costs. Therefore this technology will not be further
discussed in this work.
Double effect absorption cycles have considerably higher COPs at around 1.1 - 1.4, but require
significantly higher driving temperatures between 120°C and 170°C, so that the energetic and
economic performance of the solar thermal cooling system is not necessarily better.
Vapor-compression cycle chillers are the most common for air-conditioning applications in
commercial and residential buildings, since the technology is mature and significant improvements
have been done in order to increase the COP. Therefore, coupling this cycle with PV, as seen in
Figure 15 is an option worth. Furthermore, this cycle has been subject of study as part of innovative
applications for hybrid photovoltaic/thermal collectors to produce electricity and heat to drive a
hybrid air conditioner system.
20
Chapter 2. Technologies for active thermal solar systems
Figure 15. Solar electric vapor-compression cooling cycle
The use of a conventional vapor-compression chiller powered by photovoltaic panels for A/C
has been proved technically viable in Turkey without the use of an auxiliary device to provide
cooling [17]. This study concludes that the required panel area and COP vary depending on the
evaporator temperature and the month of the year; which corresponds to a specific cooling capacity
in the range of 0.10 – 0.35 kW/m2. This value is lower in comparison with other technologies – see
Table 1, because, although the chiller COP is high, the global efficiency of the whole system is
reduced because of the current low efficiencies of PV modules, between 12% to 15%.
Solar refrigeration can also be applied by using thermal energy supplied from solar collectors,
rather than electricity from PV cells; approach that has received particular attention for the past
decades. Mainly because the efficiency of solar collectors is higher in comparison with PV
modules, hence the former are able to use more of the received solar energy. Furthermore, with the
use of solar thermal cooling it is possible to benefit from the existing infrastructure in residential
and commercial buildings that currently provides hot domestic water – DHW– and in some regions
space heating in winter and, reduce the environmental impact related with the use of refrigerants, as
R-22 and R-410A, in conventional A/C systems.
Absorption chillers are available on the market in a wide range of capacities and designed for
different applications. However, only very few systems are available in a range below 100 kW of
cooling capacity. Today, also a few commercial systems for small power, e.g., below 30 kW, are
available. Today absorption chillers are mainly applied if a ‘cheap’ heat source is available, such as
waste heat, district heat or heat from co-generation plants. For air conditioning applications mainly
absorption chillers using the sorption pair H2O–LiBr are applied.
21
Chapter 2. Technologies for active thermal solar systems
2.3 Solar cooling technologies
The key components of solar air conditioning systems are the solar collector subsystem and the
thermally driven cooling subsystem. Additional key components are a heat rejection unit to reject
the waste heat of the thermally driven chiller and storages (hot, cold).
Solar energy can be converted into cooling using two main principles:
•
Electricity generated with photovoltaic modules can be converted into cooling using
well-known refrigeration technologies that are mainly based on vapour compression
cycles
•
Heat generated with solar thermal collectors can be converted into cooling using
thermally driven refrigeration or air-conditioning technologies. Most of these systems
employ the physical phenomena of sorption in either an open or closed thermodynamic
cycle. Other technologies, such as steam jet cycles or other cycles using a conversion of
heat to mechanical energy and of mechanical energy to cooling are less significant
Techniques which allow the use of solar thermal collectors for air-conditioning of buildings
regarding the thermally driven chillers can be distinguished in two main types:
– Closed cycle where thermally driven chillers are used to produce chilled water which can be
used for any type of air-conditioning equipment
– Open cycles, also referred to as desiccant cooling systems, are used for direct treatment of air
in a ventilation system.
Many details about components and systems for using solar thermal energy for air-conditioning
application may be found in [18] .
Solar or waste heat driven closed absorption or adsorption chillers and open desiccant
evaporative cooling systems (DEC) offer the potential to provide summer comfort conditions in
buildings at low primary energy consumption. The future of many of the methods will depend on
development beyond the cooling process itself. Temperature constrains in the operation of collector
limit what can be expected of solar cooling processes.
In solar cooling system mostly used solar collectors are flat plate solar collectors (FPC),
evacuated tube collectors
(ETC), are stationary type and due to their temperature range as
presented in Table 2 are the most commonly used for solar single-effect cooling systems. ETC, like
FPC, are able to collect both beam and diffuse radiation; however has better overall performance
than FPC, since their efficiency is greater at low incidence angles [15] and have less thermal losses
[16].
22
Chapter 2. Technologies for active thermal solar systems
Table 2. Solar collectors for single-effect absorption cooling systems
Collector type
Absorber type
Concentration
ratio1
Temperature
range2 °C
30-803
50-200
Flat plate collector - FPC
Flat
1
Evacuated tube collector - ETC
Flat
1
1-5
60-240
Compound parabolic collector-CPC
Tubular
*1 Concentration ratio = the aperture area divided by the receiver/absorber area of the collector
*2 At nominal conditions
*3 There are collectors, using vacuum insulation or transparent insulation TI, which can achieve higher values.
Also with the use of highly selective coatings, temperatures of 100 °C can be obtained [15].
Compound parabolic collectors (CPC) are able to reflect the incoming direct radiation to the
absorber over wide-ranging angles. In contrast, its performance is significantly reduced when the
amount of beam radiation is diminished in cloudy days.
The performance of a solar cooling system depends strongly on its components, how are
interconnected and the control strategy adopted this includes the coupling between the chiller and
the solar field. Situation that is particularly important for solar absorption systems, where there is a
negative relation between the COP of the chiller and the efficiency of the solar collector. Since the
former generally increases with the generator operating temperature while the performance of the
solar collector is reduced. Therefore, there is an optimum temperature to drive the generator that
maximizes the global efficiency of the solar system. This temperature is not a constant value, rather
is a function of several parameters, as the ambient temperature and humidity, solar irradiation,
which vary significantly during the operation of the system [19].
In the last decade, the main progress was made in the field of small capacity thermally driven
chillers and SAC has significantly contributed to stimulate this development. Today, numerous
systems from various manufacturers are offered on the market and have reached a considerable
technical maturity. Most of the manufacturers are small start-up companies. Some of those
companies achieved a status where they set up a manufacturing capacity on an industrial scale and
others are on the way to follow.
However, solar cooling is still in the early stages of market development; costs need to be
reduced through further development and increased deployment. A standardized, effective and
simplified range of technology arrangements require development – particularly for single family
and multi-family dwellings – to enable solar cooling to compete with conventional and supported
renewable technologies and achieve widespread deployment. Quality assurance and system
certification procedures are also needed to help stimulate the market by building buyer confidence.
23
Chapter 2. Technologies for active thermal solar systems
2.3.1 Absorption cooling
The first evolution of an absorption system began in the 1700s. It was observed that in the
presence of H2SO4 (sulfuric acid), ice can be made by evaporating pure H2O (water) within an
evacuated container. In 1810, it was found that ice could be produced from water in a couple of
vessels connected together in the presence of sulfuric acid. As the H2SO4 absorbed water vapor (to
reduce heat), ice formed on the surface of water. However, difficulties emerged with leakage and
the corrosion of air into the void vessel. In 1859, a French engineer named Ferdinand Carrede
signed a machine that used a working fluid pair of water and ammonia. This machine was used for
making ice as well as storing food. In 1950, a new system was introduced with a water/lithium–
bromide pair in gas working fluids for commercial purposes [20].
Currently, absorption chillers are the most common thermally-driven cooling process in solar
cooling installations. Common absorption cooling pairs include ammonia-water and water-lithium
bromide, with many sorption chillers available commercially over a range of capacities, but few at
capacities of 100 kWth or less. The so-called “single effect” absorption chillers typically need heat
with temperatures in the range of 70 to 100°C, and achieve a coefficient of performance (COP) of
about 0.7.
The primary advantage of an absorption system is that it has a larger COP (coefficient of
performance) than other thermally operated technologies.
In closed absorption cycles the refrigerant is conserved and reused repeatedly in successive
cycles. Heat exchange but not fluid exchange takes place between the refrigerant and the
atmosphere. Closed sorption cooling cycles work under the same principle as vapor-compression
machines, except that the mechanical compressor is replaced by a thermally driven compressor:
Figure 16. a) Vapor compression chiller
b) Absorption(thermally driven) chiller
Absorption air-conditioning is compatible with solar energy since a large fraction of the energy
required is thermal energy at temperatures that currently available flat-plate collectors can provide.
Solar absorption air conditioning has been subject of investigation by a number of researchers [2130].
24
Chapter 2. Technologies for active thermal solar systems
A single stage absorption chiller is essentially a three temperatures device. It removes a
quantity of heat QE from a source at low temperature TE, it rejects a quantity of heat QM to a
source with a temperature close to the environment TM, it is driven by a quantity of heat QG taken
from a source at high temperature TG. The quantity of heat QE is the refrigeration capacity of the
chiller.
On Figure 16 is shown a schematic of an absorption refrigeration system. As mentioned before
absorption refrigeration differs from vapor-compression air-conditioning only in the method of
compressing the refrigerant. In absorption pressurization is accomplished by first dissolving the
refrigerant in a liquid (the absorbent) in the absorber section, then pumping the solution to a high
pressure with an ordinary liquid pump. The low boiling refrigerant is then driven from solution by
the addition of heat in the generator. By this means the refrigerant vapor is compressed without the
large input of high grade shaft work that the vapor-compression air-conditioning demands. On
Figure 17 , two vessels can be individuated, a condenser (C) and an evaporator (E). These are also
found in many other refrigerating machines (like vapor compression chillers). A refrigerant, that in
the present case is pure water, is condensed in the high pressure vessel C, until it reaches the state
of saturated liquid (point 8). The condensation process releases the latent heat QC to an external
stream. The refrigerant is then brought to the low pressure vessel E through a lamination valve or a
U tube (point 9). Due to the lamination it is now a mixture of vapor and liquid. In the evaporator
the liquid part of the refrigerant is brought to the state of saturated vapor. The evaporation takes
place thanks to a quantity of heat QE extracted from an external stream.
If water vapor (point 10) is taken again to the high pressure level the cycle can be closed. How
this is accomplished is the key of the absorption technology. The saturated vapor enters the vessel
A, named absorber, which is at a pressure almost identical to that of the evaporator.
Figure 17. Schematic diagram of an single stage absorption chiller (LiBr-H2O)
25
Chapter 2. Technologies for active thermal solar systems
For the thermodynamic analysis of the absorption system the principles of mass conservation
and the first and second laws of thermodynamics are applied to each component of the system.
Each component can be treated as a control volume with inlet and outlet streams, heat transfer and
work interactions. In the system mass conservation includes the mass balance of each material of
the solution. The governing equations of mass and type of material conservation for steady-state,
steady-flow systems are [31]:
•
•
(1)
∑ mi −∑ mo = 0
•
•
∑ (m, x) −∑ (m, x)
i
o
(2)
=0
•
where m is the mass flow rate and x is mass concentration of LiBr in the solution. The first law
of thermodynamics yields the energy balance of each component of the absorption system as
follows:
•
•
∑ (m, h)i −∑ (m, h)o + Qi − Qo + W = 0
(3)
An overall-steady state energy balance on the absorption cooler indicates that the energy
supplied to the generator and to the evaporator must equal the energy removed from the machine
via the coolant flowing through the absorber and the condenser plus whatever net losses may occur
to the surroundings
QG + QE = QA + QC + Qlosess
(4)
The useful output energy of the system for heating applications is the sum of the heat rejected
from the absorber and the condenser while the input energy is supplied to the generator.
The thermal coefficient of performance COP is defined as the ratio of energy into the
evaporator QE to the energy into the generator QG.
COP =
(5)
QE
QG
The coefficient of performance is useful index of performance in solar cooling, where collector
costs (and thus cost of QE) are important. Many LiBr-H2O machines have nearly constant COP as
the generator temperatures vary over the operating range as long as the temperatures are above the
minimum.
Others types of COP can be defined. A COPe is the ratio of cooling to electrical energy used to
provide air and liquid flows, operate controls etc.
26
Chapter 2. Technologies for active thermal solar systems
COPe =
(6)
Qe
Electric Input
Solar fractions on the heating energy demand of the absorption chillers therefore need to be
higher than about 50% to start saving primary energy [32].
The most common used refrigerant-absorber pairs are: Lithium Bromide (absorber) and water
(refrigerant) LiBr-H2O , and Ammonia (refrigerant) – Water(absorber) NH3-H2O. Because the
LiBr-H2O has high volatility ratio, it can operate at lower pressures and therefore at lower generator
temperatures achievable by flat-plate collectors. A disadvantage of this system is that the pair tends
to form solids. LiBr has tendency to crystalize when air cooled and system cannot be operated at or
below the freezing point of the water. Therefore the LiBr-H2O system is operated at evaporator
temperatures of 5 °C or higher. Using the mixture of LiBr with some other salt as the absorber can
overcome the crystallization problem. It over comes the problem of using a filter as the pair is not
volatile, and it has a very high latent heat of vaporization. However, because water is used as the
refrigerant in this pair, there are problems with low temperature operation at temperatures below
0°C. Moreover, the system requires vacuum conditions and at high concentrations, the pair tends to
be crystalline. There are also issues of corrosion with some metals. A brief thermodynamic analysis
of this pair is available in previous investigations [33, 34].
The ammonia-water system has advantage that it can be operated down to very low
temperatures. However for temperatures much below 0°C water vapor must be removed from
ammonia as much as possible to prevent ice crystals from forming. This requires a rectifying
column after the boiler. Also ammonia is a safety Code Group B2 fluid (ASHRAE Standard 341992) which restricts its use indoors [35].
Single effect LiBr-H2O absorption chillers operate with thermal COP limited to 0,7 and actual
operating coefficients may be very much less due to cycling and other problems. Double effect
chillers in which two generators in series are used can have COP values in the range 1.0 to 1.5.
These improvements are obtained at the expense of considerable complication of the machines and
will probably require the use of collector operating at temperatures beyond the range of flat-plate
collectors.
2.3.2 Adsorption cooling
Adsorption cycles for refrigeration were first used in the early 1900s. Plank and Kuprianoff
(1960) reported on manufactured machines using ammonia/CaCl and methanol/activated carbon.
Hulse (1929) reported on a sulfur dioxide/silica gel machine for the air-conditioning of rail freight
cars in the United States. Critoph (2012), in his historical review, reports on a domestic activated
27
Chapter 2. Technologies for active thermal solar systems
carbon/methanol refrigerator called ‘Eskimo’ and sold in the 1930s by the Norwegian Amundsen
Refrigerator Company. In general , this technology was recognized in 19th century so its practical
application in the field of refrigeration is relatively recent.
The concentration of adsorbate vapors in a solid adsorbent is a function of the temperature of
the pair i.e. the mixture of adsorbent and adsorbate and the vapor pressure of the latter. The
dependence of adsorbate concentration on temperature under constant pressure conditions makes it
possible to adsorb or desorb the adsorbate by varying the temperature of the mixture.
An adsorbent – refrigerant working pair for a solar refrigerator requires following
characteristics:
Refrigerant should have large latent heat of evaporation. A low heat of desorption under the
envisaged operating pressure and temperature conditions and low thermal capacity.
The conventional adsorption cycle has been presented extensively in the literature [31-33] and
it mainly includes two phases:
1. Adsorbent cooling with adsorption process which results in refrigerant evaporation inside
the evaporator and thus in the desired refrigeration effect. At this phase the sensible heat and the
adsorption heat are consumed by a cooling medium which is usually water or air
2. Adsorbent heating with desorption process also called generation which results in
refrigerant condensation at the condenser and the heat release into the environment. The heat
necessary for the generation process can be supplied by a low-grade heat source such as solar
energy, waste heat etc.
When the temperature of the heating medium ranges from 60 to 95 ◦C, the coefficient of
cooling efficiency for adsorption chillers is 0.6–0.7 [9–11]. With a solid sorbent, the coefficient of
cooling efficiency (COP) in open systems is 0.6–1.0 (when the temperature of the heating medium
ranges from 45 to 95 ◦C), whereas with a liquid sorbent the COP is about 1.0 (when the temperature
of the heating medium ranges from 60 to 80 ◦C) [34].
Closed cycle adsorption cooling
A closed cycle desiccant system has been developed in which the refrigerant (water) is cycled.
During the cooling process water previously condensed is injected into a flash evaporator and
evaporates to provide cooling. The vapor is adsorbed in a zeolite desiccant which allows low ( 10
°C – 15 °C) temperature in the evaporator. The water evaporated from the zeolite (or silica gel)
during regeneration is condensed in sensible exchanger by heat rejection to the surroundings. The
condensed water is then recycled in the evaporator.
28
Chapter 2. Technologies for active thermal solar systems
Figure 18. Simple schematic drawing of an (solar driven) absorption chiller
In contrast to absorption chillers, adsorption chillers use solid sorption materials. Market
available systems use water as refrigerant and silica gel as sorbent. Because of the fact that the solid
sorbent cannot be circulated adsorption chillers consist of two separate chambers Figure 18, which
both contain the adsorbent. Besides these two adsorbent chambers, there is one evaporator and one
condenser (coupled to the heat sink).
The adsorbent containing the water is heated (by the solar collectors) and the adsorbed water is
expulsed as water vapor and condenses in the condenser and the condensed water is transferred to
the evaporator. The adsorbent is cooled again, leading to a lower pressure in the sealed system
where the water in the evaporator evaporates, taking up the heat from the chilled water circuit, after
which water vapor is adsorbed in the adsorbent (adsorption heat is evacuated).
One of the first companies that offers commercial adsorption chillers is the SorTech AG,
founded in 2002 as a spin-off of the German “Fraunhofer-Institut für Solare Energiesysteme
ISEscale” solar cooling application. Starting from 2007 the SorTech chillers are silica gel based
with cooling capacities starting from 5 kW to 15 kW per unit. In the meantime more than 200
projects mainly in Europe, but as well in Africa, North America, Asia, and Australia could be
realized with cooling capacities up to 150 kW.
2.3.4 Open-cycle solar desiccant cooling
Characteristic for open systems, working as an aerial set is that they become particularly
ineffective when the heat gains are significant. In that case the excess heat from the building is
removed by increasing the air circulation in a room. A critical defect of open systems is the limited
possibility of their use. They may be operated only in such conditions when the air is cooled at a
central air-conditioning unit and when the air treatment depends on its cooling. An additional
constraint is the inability to achieve a low temperature for the air supply. All these have negative
29
Chapter 2. Technologies for active thermal solar systems
impact on obtaining the desired levels of the supply air. Therefore, the adsorption refrigerating
systems supplied by solar energy appears to be the most promising solution to the problem.
Solar desiccant air conditioning systems use a desiccant to remove moisture from an air stream.
The dried air is processed by evaporative coolers to produce a relatively dry cool air stream which
then cools the building space. Solar energy can be used to regenerate the desiccant. These systems
offer the promise of a high thermal coefficient of performance using the moderately low
temperatures compatible with solar collectors.
Mainly are used two types of desiccant systems: solid desiccant system in which the desiccant
is a stationary bed or rotating matrix and liquid desiccant system in which desiccant is pumped
round between air streams.
Air flows through the components in open-cycle systems and moisture and heat are transferred
between the air stream the atmosphere and heat sources. The ambient serves as a heat sink for the
discharged moisture regarding the closed cycle which only exchange heat with the surrounding
atmosphere. The desiccant is regenerated with solar energy.
Warm and humid air (fresh air) enters the slowly rotating desiccant wheel and is dehumidified
by adsorption of water. Because the air is heated up by the adsorption heat there is also installed a
heat recovery wheel. The result is a significant pre-cooling of the supply air stream. After the precooling the air is humidified and further cooled by a controlled humidifier according to the desired
temperature and humidity of the supply air stream. The exhaust air stream (warm, humid) of the
rooms is humidified close to the saturation point to exploit the full cooling potential in order to
allow an effective heat recovery. To allow a continuous operation of the dehumidification process
the sorption wheel has to be regenerated by applying heat in a comparatively low temperature range
from50-75°C
30
Chapter 3
3. Numerical modeling of solar thermal system
The method for the modeling is separated into three distinct stages: Building modeling stage
and system plant modeling stage, and stage 3 manual calculations. This can be summarized in the
following key stages:
Stage 1: Building model
•
Development of building models (geometrics, construction details) for the building
•
Apply future weather data to building model
•
Simulate cooling demand from building models for various scenarios
Stage 2: TRNSYS solar model
•
Development of solar powered absorption plant
•
Apply loads generated from stage 1 model for specific future weather years
•
Simulate cooling delivered by solar absorption cooling system
Stage 3: Manual calculations
•
Calculate COP of solar powered cooling system
•
Calculate additional cooling to be delivered by cooling/heating system
3.1 Energy simulation software
The development of solar simulation capabilities greatly assist in the promotion of practical
solar systems. Simulations, like any other calculations are only as good as the models that are the
basis of the program and the skill which they are used. Some of the programs that have been
applied to solar processes have been written specifically for study of solar energy systems. Other
were intended for non-solar applications but have had models of solar components added to them to
make them useful for solar problems.
Simulations are numerical experiments and can give the same kinds of thermal performance
information as physical experiments. They are however relatively quick and inexpensive and can
31
Chapter 3. Numerical modeling of solar thermal system
produce information on effect of design variable changes on system performance by series of
experiments all using exactly the same loads and weather. These design variables could include
selectivity of the absorbing surface, number of covers on the collector, collector area etc. With cost
data and appropriate economy analysis, simulation results can be used to find least cost systems.
There are two basic kinds of data that can be obtained from simulations. First, integrated
performance over extended periods can be determined. This is normally wanted for a year that
represents the long-term average climate in which is proposed process would operate. A year is the
time base of most economic studies, but information may be needed for other periods from days to
spans of many years.
The extent to which simulations represent the operation of real physical systems depends on the
level of detail included in the numerical experiment. Component models can vary in complexity, as
can system. In principle simulations can be as detailed as the user wishes. In practice there may be
factors in system operation which are difficult to simulate such as leaks in air systems and
operation of real systems may be less ideal than the simulation indicates.
3.1.1 Review of simulation software’s used in solar air-conditioning
Here are presented the results from the review i.e. the technical report of Subtask C, Task 38
where is given brief description of some simulation tools applicable in solar air-conditioning
systems. These tools are the most commonly used by IEA task 38 participants.
SPARK
SPARK is a general simulation environment that supports the definition of simulation models
and solution of these models via a robust and efficient differential/algebraic equation solver. In
SPARK, the modeler describes the set of equations defining a model as equation-based objects. At
the lowest level, an atomic object characterizes one equation and its variables. Then, macroscopic
objects can be created as an assembly of various atomic or macroscopic objects.
SPARK has its own HVAC library based on some simple models.
ENERGY PLUS
EnergyPlus is an energy analysis and thermal load simulation program. Based on a user’s
description of a building from the perspective of the building’s physical make-up, associated
mechanical systems, etc., EnergyPlus will calculate the heating and cooling loads necessary to
maintain thermal control set points, conditions throughout an secondary HVAC system and coil
loads, and the energy consumption of primary plant equipment as well as many other simulation
details that are necessary to verify that the simulation is performing as the actual building would.
Many of the simulation characteristics have been inherited from the legacy programs of BLAST
32
Chapter 3. Numerical modeling of solar thermal system
and DOE–2. it is the intent of EnergyPlus to handle as many building and HVAC design options
either directly or indirectly through links to other programs in order to calculate thermal loads
and/or energy consumption on for a design day or an extended period of time (up to, including, and
beyond a year).
EES
EES is an acronym for Engineering Equation Solver. The basic function provided by EES is the
numerical solution of a set of algebraic equations. EES can also be used to solve differential and
integral equations, do optimization, provide uncertainty analyses and linear and non-linear
regression, convert units and check unit consistency and generate publication-quality plots.
EASYCOOL
EasyCool provides 11 pre-defined system configurations for solar thermally driven cooling
applications, of which 4 configurations are foreseen for reference calculations (non-solar,
conventional system solutions). The program reads annual time series with hourly building load
data and meteorological data of the respective site (these data set has to be prepared in advance)
and calculates annual energetics and economic performance data as well as environmental figures
such as CO2 savings.
INSEL
The acronym INSEL stands for INtegrated Simulation Environment Language. This graphical
programming language has been developed at the Faculty of Physics of Oldenburg University
(Germany) in the early 1990’s and was originally designed for the modelling of renewable
electrical energy systems. The graphical programming language INSEL is based on the principle of
“structured programming” on blocks diagrams. It consists of connecting blocks in order to obtain
block diagrams that express a solution for a certain simulation task.
TRNSYS
Trnsys is a widely used modular thermal process simulation program. Originally developed for
solar energy applications, it now is used for simulation of wider variety of thermal processes.
Subroutines are available that represents the components in typical solar energy systems.
TRNSYS is a complete and extensible simulation environment for the transient simulation of
systems, including multi-zone buildings. It is used by engineers and researchers around the world to
validate new energy concepts, from simple domestic hot water systems to the design and simulation
of buildings and their equipment, including control strategies, occupant behavior, alternative energy
systems (wind, solar, photovoltaic, hydrogen systems), etc. The simulation engine is programmed
in Fortran and the source is distributed. The engine is compiled into a Windows Dynamic Link
33
Chapter 3. Numerical modeling of solar thermal system
Library (DLL), TRNDll. The TRNSYS kernel reads all the information on the simulation (which
components are used and how they are connected) in the TRNSYS input file, known as the deck
file (*.dck). It also opens additional input files (e.g. weather data) and creates output files.
Users can readily write their own component subroutines. By a simple language, the
components are connected together in manner analogues to piping, ducting and wiring in a physical
system. The DLL-based architecture allows users and third-party developers to easily add the
custom component models, using all common programming languages (C, C++, PASCAL,
FORTRAN, etc.). In addition, TRNSYS can be easily connected to many other applications, for
pre- or post-processing or through interactive calls during the simulation (e.g. Microsoft Excel,
Matlab, COMIS, etc.).
Current versions of TRNSYS have in executive program convergence promoters and other
means of speeding computations. There are three integration algorithms in TRNSYS, the user can
choose the one best suited to the problem at a hand. The one that is extensively used is the modified
Euler method. It is essentially a first order predictor corrector algorithm using Euler’s method for
the predicting step and the trapezoid rule for the correction step.
TRNSYS applications include:
• Solar systems (solar thermal and PV)
• Low energy buildings and HVAC systems with advanced design features (natural ventilation,
slab heating/cooling, double façade, etc.)
• Renewable energy systems
• Cogeneration, fuel cells
• Anything that requires dynamic simulation!
As an additional component library (also used in this work) is the developed by the company
Thermal Energy System Specialist (TESS). The TESS Applications Library is an assortment of
scheduling and setpoint applications that use the TRNSYS Simulation Studio plugin feature. These
components are extremely useful for creating daily, weekly, monthly schedules, normalized
occupancy, lighting, or equipment schedules, and setpoints for thermostats.
3.2 Numerical components models
Mathematical description of the used component models are given as described in the TRNSYS
manual Mathematical references since all of the simulations are peroformed in TRNSYS.
34
Chapter 3. Numerical modeling of solar thermal system
3.2.1 Collector model – Flat plate collector
For the flat plate solar collector it is used the Type 539 from the TESS library. It is chosen this
model because compared to Type 1 from the TRNSYS library it takes into consideration the
influence of the capacitance effect to the temperature change.
The operation of most solar energy systems is inherently transitient. There is no such thing as
steady-state operation when one considers the transitient nature of the driving forces. This
observation has led to numerical studies by Klein, Wijeysundera and others on the effects of
collector heat capacity on collector performance. The effects can be regarded in two distinct parts.
One part is due to the heating of the collector from its early morning low temperature to its final
operating temperature in the afternoon. The second part is due to intermittent behavior during the
day whenever the driving forces such as solar radiation and wind change rapidly.
The rate of useful gain for a flat- plate collector according the Duffie and Beckaman :
Qu = AC FR [S − U L (Ti − Ta )]+
(7)
Where
- useful gain from the collector
Qu [kJ/h]
Ac [m2]
- collector aperture area
2
S [kJ/h m ]
- solar radiation absorbed per collector unit area
UL [kJ/h m2 K] - overall collector heat loss coefficient
- fluid inlet temperature
Ti [K]
FR [-]
- collector heat removal factor based on collector inlet temperature
where the + sign implies the presence of a controller and that only positive values of the term in
the brackets should be used. Operation of a forced – circulation collector will not be carried out
when Qu < 0 (or when Qu < Qmin where Qmin is a minimum level of energy gain to justify pumping
the fluid through the system). In real systems, this is accomplished by comparing the temperature
of the fluid leaving the collector ( i.e. the top header) with the temperature of the fluid in the exit
portion of the storage tank and running the pump only when the difference in temperatures is
positive and energy can be collected.
The rate of useful gain is also given by:
•
Qu = m⋅ C p (To − Ti )
(8)
•
Where m is the output of the pump circulating through the collector.
If the storage unit is fully mixed sensible heat unit, its performance is given by equation:
dTs
= Qu − LS − (UA) S (TS − Ta )
(9)
dt
The equivalent equations for stratified water tank storage systems, pebble bed exchangers or
(mC p ) S
heat of fusion systems are used in lieu of Equation 66(9) as appropriate. These equations are the
35
Chapter 3. Numerical modeling of solar thermal system
basic equations to be solved in the analysis of systems such as simple solar water heater with
collector, pump, controller and storage tank. The rate of energy removal to meet all or part of a load
is LS and is time-dependent ; S and Ta are also time dependent.
If the collector is discretized into isothermal temperature nodes, the governing differential
equation for node j can be expressed as:
Cj
dT j
dt
•
= F ' ( S j − A jU L (T j − Ta )) − m C p (T j − Tin , j )
(10)
Where:
Cj =
C
A
S
; Aj =
; Sj =
# Nodes
# Nodes
# Nodes
(11)
S = (τα ) n IAM ⋅ A ⋅ I T
(12)
Tin, j = T j −1 where To = Tin
(13)
The dependence of (τα) on the angle of incidence of radiation on the collector varies from one
collector to another and the standard test methods include experimental estimation of this effect.
Collector tests are generally performed on clear days at normal incidence so that the transmittance absorptance product (τα) is nearly the normal incidence value for beam radiation, (τα)n. The
intercept efficiency, FR (τα)n, is corrected for non-normal solar incidence by the factor (τα)/( τα)n.
By definition, (τα) is the ratio of the total absorbed radiation to the incident radiation. Thus, a
general expression for (τα)/( τα)n i.e. the incident angle modifier (IAM) is written as:
IAM =
(τα )
=
(τα ) n
I bT
(τα ) dg
(τα ) b
(τα ) ds
+ I dsT
+ I dgT
(τα ) n
(τα ) n
(τα ) n
IT
(14)
where:
(τα)
(τα)n
(τα)b
(τα)ds
(τα)dg
IT
IbT
IdsT
IdgT
surface)
[0…1]
[0…1]
[0…1]
[0…1]
[0…1]
[kJ/h m2]
[kJ/h m2]
[kJ/h m2]
[kJ/h m2]
- Product of the cover transmittance and the absorber absorptance
- (τα) at normal incidence
- (τα) for beam radiation (depends on the incidence angle θ)
- (τα) for sky diffuse radiation
- (τα) for ground reflected radiation
- Total radiation incident on the solar collector (Tilted surface)
- Beam radiation incident on the solar collector
- Sky diffuse radiation on the solar collector (tilted surface)
- Ground-reflected diffuse radiation on the solar collector (tilted
A general expression has been suggested by Souka and Safwat (1966) for angular dependence
of IAM for collectors with flat covers as:
1
IAM = 1 + b0 (
− 1)
(15)
cosθ
36
Chapter 3. Numerical modeling of solar thermal system
where θ is the incidence angle and b0 is a constant called the incidence angle modifier
coefficient – this equation follows the ASHRAE 93-77 convention, and b0 is generally negative
number. At larger angles of incidence, the linear relationship no longer applies, but most of the
useful energy absorbed in a collector system will be at times when θ is less than 60 °.
For flat-plate collectors, (τα)b/(τα)n can be approximated:
(τα ) b
1
1
= 1 − bo (
− 1) − b1 (
− 1) 2
(τα ) n
cosθ
cosθ
(16)
The incidence angle modifiers for both sky, (τα)ds/(τα)n, and ground diffuse, (τα)dg/(τα)n, are
determined by defining equivalent incidence angles for beam radiation that give the same
transmittance as for diffuse radiation.
In this component model 5 (five) optical modes can be selected in order to input the IAM data:
•
Optical mode 1: perfect IAM (τα)/( τα)n=1 for any incidence angle
•
Optical mode 2: the user specifies the values of b0 and b1 in Equation 16
•
Optical mode 3: values of (τα)b/( τα)n versus θ are supplied in an external data file but
the
collector is assumed to be symmetrical so only one direction is provided in the data file
•
Optical mode 5: values of (τα)b/( τα)n versus θ are supplied in an external data file for
both
the longitudinal and transversal directions, this mode is usually used to simulate
evacuated collectors
The effective incidence angles for sky diffuse and ground reflected radiations are:
θ sky = 59.68 − 0.1388β + 0.001497 β 2
θ gnd = 90.00 − 0.5788β + 0.002693β 2
An estimate of the loss coefficient and transmittance-absorptance product at normal incidence
can be made if the parameters of the collector efficiency test are known.
A general equation for solar thermal collector efficiency can be obtained from the HottelWhillier equation at steady-state conditions as:
•
m C p (To − Ti )
Q
T −T
η= U =
= FR (τα ) n − FRU L i a
A ⋅ IT
AI T
IT
(17)
But today the collector efficiency equation has different form which is used in the collector
efficiency certificates:
η = η 0 − a1
(T − Ta ) 2
Tm − Ta
− a2 m
IT
IT
(18)
where:
37
Chapter 3. Numerical modeling of solar thermal system
η0
a1
a2
IT
[-]
[W/m2 K]
[W/m2 K]
[W/m2]
Optical efficiency (conversion factor)
Heat transfer coefficient
Temperature depending heat transfer coefficient
Solar radiation at which the measurement is performed ( usually 1000 W/m2)
These parameters are available for collectors tested according to ASHRAE standards and rated
by SRCC (ASHRAE, 2003; SRCC, 1995), as well as for collectors tested according to the recent
European Standards on solar collectors (CEN, 2001).
Correction to the ideal efficiency curve
Analytical corrections are applied to the collector parameters to account for:
•
Operations at flow rates different from the test conditions
•
NS identical collectors mounted in series
•
Non-normal solar incidence
In order to account for conditions when the collector is operated at a flow rate other than the
value at which it was tested, both FR (τα) and FRUL’ are corrected to account for changes in FR. The
ratio, r1, by which they are corrected is given by:
•
r1 =
FRU ' L
FRU ' L
=
USE
FR (τα ) n
USE
FR (τα ) n
TEST
=
•
mC p 
− AF 'U L / m C p 
1 − e
 USE
AF 'U L 

•
(19)
m test C p 
− AF 'U L / m C p 
1 − e
 TEST
AF 'U L 

That quantity can be calculated from the test conditions:
TEST
•
F 'U L = −
•


 F U' A
ln1 − R• L 
A

m C p 

mCp
(20)
For liquid collectors, F'UL calculated from the test conditions is approximately equal to F'UL at
use conditions and can be used in both the numerator and denominator of Equation 19.
Calculation of the collector outlet temperature at steady-state conditions at normal incidence
can be made with the equation:
To = Ti +
A
•
( FRτα N I T − FRU L (Ti − Ta ))
mC p
(21)
However, we rarely have steady-state conditions in a real system and there are many times
where the mass of the collector impacts the collector outlet temperature. Equations 6610 and 12
may be combined to give the general differential equation for a node of the solar collector.
38
Chapter 3. Numerical modeling of solar thermal system
dT j
(22)
= F ' (τα ) n ⋅ IAM ⋅ A j ⋅ I T − AJ F ' U L (T j − Ta ) − mC p (T j − Tin , j )
dt
Without the collector heat removal factor, it becomes impossible to calculate the isolated
Cj
collector fin efficiency factor (F’), the collector transmittance-absorptance product at normal
incidence (ταn) and the collector heat loss coefficient (UL) required solving the collector node
energy balance.
Knowing the heat removal factor FR the other terms can be calculated as:
•


 F AU L 
LOG1 − R•
(23)

AU L

m
C
P 

The collector efficiency results can be presented as a function of the average collector
F '= −
mtest C p ,test
temperature or the outlet collector temperature and not as a function of the collector inlet
temperature.
η=
QU
(T − T )
(T − T )
= Fav (τα ) n − FavU L av a = F0U L o a
AI T
IT
IT
(24)
3.2.1.1 Collector model – Evacuated Tube Collector
Evacuated tube collectors (ETCs) have demonstrated that the combination of selective surface
and an effective convection suppressor can result in good performance at high temperatures. In the
component model for the vacuum tube collector Type 538 is used from the TESS library.
The vacuum envelope reduces convection and conduction losses, so the collector can operate at
higher temperatures than flat-plate collectors. Like flat-plate collectors they collect both direct and
diffuse radiation. However their efficiency is higher at low incidence angles. This effect tends to
give evacuated tube collectors an advantage over flat-plate collectors in terms of daylong
performance.
With the Type 538 is modeled an evacuated tube solar collector with a quadratic efficiency
curve and the off-normal radiation effects which can be treated with bi-axial incidence angle
modifiers and the user has the option of controlling the flow rate through the collector to maintain a
desired outlet temperature. The capacitance (mass) of the collector is not accounted for in this
model; steady state conditions are assumed. The total collector array may consist of collectors
connected in series and in parallel. The thermal performance of the total collector array is
determined by the number of modules in series and the characteristics of each module.
Regarding the mathematical description of the evacuated tube collectors the only difference in
the equations is in the IAM i.e. this type of collectors has biaxial Incidence Angle Modifiers.
39
Chapter 3. Numerical modeling of solar thermal system
The transversal incidence angle is measured in a plane that is perpendicular to both the collector
aperture and the longitudinal plane. The corresponding IAM is referred to as the transversal IAM,
or azimuthal modifier.
Figure 19. Transversal and longitudinal directions
The collector tube represented in Figure 19 is in the most common orientation, along a NorthSouth axis (assuming the collector faces due South). If the collector was tested in a different
configuration or if it is mounted with the tubes along an East-West axis, the IAM data obtained
from a collector test may have to be adapted (by switching the longitudinal and transversal
directions). In TRNSYS, "longitudinal" and "transversal" always refer to the plane of the collector
and the sun as described here below, not to the tubes.
In most cases, collector test reports provide the transversal IAM for different transversal
incidence
angle values (and longitudinal angle = 0) and longitudinal IAM for different longitudinal incident
angles (and transversal angle = 0). The data file requires the IAM for non-zero longitudinal and
transversal angles.
3.2.2 Thermal energy storage model
Thermal energy storage is needed if a solar heating system is to provide heat overnight or
during cloudy periods of the day. A variety of materials can be used for heat storage and the
desirable characteristics are the following: easy to transfer heat from the heated fluid to the storage
material; readily recoverable and supplied to the heating system; subject to little internal losses or
losses to the environment; inexpensive; requiring minimal floor area or volume.
As is generally well known, offsets and intermittence make achieving the potential of solar
technologies relatively complex. The bulk of energy production occurs at midday, if the sky is
clear; and during summer. Meanwhile, consumption is higher during winter, especially at morning
and night when space heating loads tend to be at their peak and occupants use more hot water. The
diurnal offset is relatively easy to compensate for with water tanks (buffers) and other short term
40
Chapter 3. Numerical modeling of solar thermal system
storage methods like the use of the building’s thermal mass. At the seasonal scale however,
solutions are more complex and expensive.
Seasonal storage systems are much larger than short-term ones. Braun [36] evaluated that
storage capacities per unit of collector area must be two to three orders of magnitude (100–1000
times) larger for seasonal storage than for overnight storage. Nevertheless, Fisch et al. reported
investment costs per square meter of solar collector for large scale solar plants only twice as high
for systems with seasonal storage than for systems with short-term storage. According to Braun et
al., significant reductions in solar collector requirements for heating could be achieved by using
seasonal storage at northern latitudes, where seasonal variations are large, and in cold climates,
where DHW loads are much smaller than space heating loads. Since solar collectors tend to be
expensive, there is definitely potential in developing more economical storage systems in order to
obtain higher solar fractions for these heating tasks.
In the analyzed system for the storage tanks is used the Type 60c from the TRNSYS library
which represents a stratified fluid storage tank with internal heaters and internal heat exchangers.
Water tank may operate with significant degrees of stratification, that is with the top of the tank
hotter than the bottom. Many stratified tank models have been developed. They fall into either of
two categories. In the first the multimode approach a tank is modeled as divided into N nodes
(sections) with energy balances written for each section of the tank. The result is set of N
differential equations that can be solved for temperatures of N nodes as function of time. In the
second, the plug flow approach segments of liquid at various temperatures are assumed to move
through the tank in plug flow and the models are essentially bookkeeping methods to keep track of
the size temperature and position of the segments.
The used tank component it is modeled by assuming that the tank consists of N (N ≤ 100) fullymixed equal volume segments, as shown in
Figure 20
Figure 20. Stratified Fluid Storage Tank
41
Chapter 3. Numerical modeling of solar thermal system
The tank component locations are entered as heights, measured from the floor up, rather than
node numbers. These components include: inlet and outlet flows, auxiliary heaters, thermostats, and
heat exchangers. In the model there are two inlet modes available, one where the inlet height is
variable i.e. the flow stream enters the node with the closes temperature. With sufficient nodes, this
permits a maximum degree of stratification. The second mode the flows inlets are predefined by
the user i.e. this mode is used in the analyzed model of this work. The temperature inversions that
will occur during the simulation are eliminated by mixing of the appropriate adjacent nodes. The
number of tank nodes is defined by the user and the tank is automatically divided by the program.
In this model are included two electric resistance heaters for which can be defined position in
the tank and thermostat temperature limit. The heater can operate in two modes. Mode 1 allows the
bottom heating element to be enabled only when the top element is satisfied. In this mode it is
impossible the both heaters to be on simultaneously. It is usually used in the DHW systems. Mode
2 allows the opposite from mode 1 i.e. the both heaters can be on in same time. In the heaters
control it is assumed dead band, The heater is enabled if the temperature of the node containing the
thermostat is less than (Tset -∆Tdb) or if it was on for the previous time step and the thermostat
temperature is less than Tset.
Regarding the tank insulation tank may not be uniformly insulated. Also A pressure relief valve
has been added to the storage tank to account for boiling effects. The user must specify the boiling
temperature of the fluid; venting will release sufficient energy to keep the tank at the boiling
temperature.
This Type 60c allows to be modeled three internal heat exchangers. There can be specified the
dimensions of the heat exchangers, as well as the temperature and flow rate of the inside fluid. The
inside fluid can be water, water/propylene glycol mixture or water/ethylene glycol mixture. Two
wall conductivity parameters are used: one for the conductivity of the heat exchanger material
itself, and one for the conductivity of the heat exchanger wall, which allows for contact resistance.
An energy balance about the ith tank segment is presented on Figure 21
42
Chapter 3. Numerical modeling of solar thermal system
Figure 21. Graphical representation of energy flows into a node [37]
If all of the energy flows are combined in one equation, then the temperature for the
temperature of node i is expressed as:
(M iC p )
(k + ∆k ) Ac ,i
dTi (k + ∆k ) Ac ,i
=
(Ti +1 − Ti ) +
(Ti −1 − Ti ) + (U tan k + ∆U i ) As ,i (Tenc − Ti )
dt
∆xi +1→i
∆xi +1→i
•
•
•
•
•
UA flue ,i (T flue − Ti ) + m down C p (Ti −1 ) − m up C p (Ti −1 ) − m down C p (Ti ) − m up C p (Ti +1 ) + γ htr1 Q aux1 +
•
γ htr 2 Q aux 2 + +UAhx1 (lmtd1 ) + UAhx 2 (lmtd 2 ) + UAhx 3 (lmtd 3 ) + m1in C pT1in − m1out C pTi +
•
(25)
•
+ m 2in T2in − m 2out C pTi
Where the γ htr 1 is the heater control signal. The temperatures of each of the N tank segments
are determined by the integration of their time derivatives expressed in the above equation. At the
end of each time step, temperature inversions are eliminated by mixing appropriate adjacent nodes.
3.2.3 Auxiliary heater
Practically sized solar system requires auxiliary heating to supply the portion of the heating
load that the solar system cannot supply. System controls are commonly designed to use solar
energy for heating and to relay on the auxiliary heating system whenever solar energy is not
available and storage has been depleted. Even if the control strategy included provisions to share
the load between the solar and auxiliary heating system during delivery, there is no assurance that
an appropriate share of solar energy would always be available. The auxiliary heater must therefore
be full capable of delivering heat at a rate sufficient to maintain comfort conditions during the
coldest days.
In the simulation for the additional heating the fluid entering the absorption chiller or building
heating system it is used model of proportionally controlled fluid heater. For the control is used
external proportional control (an input signal between 0 and 1) is in effect as long as a fluid set
43
Chapter 3. Numerical modeling of solar thermal system
point temperature is not exceeded, so If the set point is exceeded, the proportional control is
internally modified to limit the fluid outlet temperature
Figure 22. Energy balance for he auxiliary heater model
The outlet temperature when the heater is off from the energy balance equation can be
calculated the outlet temperature:
Tin
+ UATamb
2
Tout =
(26)
•
UA
mCp +
2
This model also calculates the heater losses to the environment with the following equation:
mC pTin − UA
•
Q loss = UA(
•
Tout − Tin
− Tamb )
2
(27)
•
Q fluid = m C p (Tout − Tin )
(28)
With the heater on and fluid flowing through the device the outlet temperature can be calculated
Tin
+ UATamb + γ htrη htr CAPhtr
2
=
(29)
•
UA
m Cp +
2
is the heater control signal having a value between 0 and 1, and ηhtr is the heater
mC pTin − UA
Tout
Where γhtr
efficiency. The auxiliary energy imparted to the fluid by the heating element is given by equation
•
•
Q aux =
m C p (Tout − Tin ) − UA(Tavg − Tamb )
(30)
η htr
3.2.4 Controller model - Differential controller
The basic control strategy for space and water heating solar system is to maximize electrical
energy collection and utilization and to minimize electrical energy use for collection and
distribution. While a theoretically optimum strategy for each system with a specific load in a
particular climate can be prescribed (Winn, Jonson and Moore 1974), practical control devices are
generally not available to implement the optimum strategy for each system. Fortunately, even a
simple controller performs reasonably well without inflicting severe penalties on the system. The
44
Chapter 3. Numerical modeling of solar thermal system
simples controller is a differential thermostat that starts or stops an electric motor that operates a
pump or valve.
The used differential controller in the simulations is the component Type 2 from the TRNSYS
library. This controller generates a control function γ0 that can have values of 0 or 1. The value of
γ0 is chosen as a function of the difference between upper and lower temperatures, TH and TL,
compared with two dead band temperature differences, ∆TH and ∆TL. the new value of γ0 is
dependent on whether γi = 0 or 1. The controller is normally used with γ0 connected to γi giving a
hysteresis effect. For safety considerations, a high limit cut-out is included with the TYPE 2
controller. Regardless of the dead band conditions, the control function will be set to zero if the
high limit condition is exceeded.
Mathematically, the control function is expressed as follows:
If the controller was previously on
If γi = 1 and ∆TL ≤ (TH - TL), γo = 1
(31)
If γi = 1 and ∆TL > (TH - TL), γo = 0
(32)
If the controller was previously off
If γi = 0 and ∆TH ≤ (TH - TL), γo = 1
(33)
If γi = 0 and ∆TH ≤ (TH - TL), γo = 1
(34)
However, the control function is set to zero, regardless of the upper and lower dead band
conditions, if TIN > TMAX. This situation is often encountered in domestic hot water systems where
the pump is not allowed to run if the tank temperature is above some prescribed limit.
3.2.5 Absorption chiller model
Steady-state absorption chiller models are based on the internal mass and energy balances in all
components, which depend on the solution pump flow rate and on the heat transfer between
external and internal temperature levels. Several problems are associated with a single
characteristic equation, which calculates all internal enthalpies only for the design conditions: if
bubble pumps are used, the solution flow rate strongly depends on the generator temperature. Also
if the external temperature levels differ significantly from design conditions, the internal
temperature levels change and consequently the enthalpies. Therefore, in the current work, the
internal energy balances were solved for each simulation time step as a function of the external
entrance temperatures, so that changing mass flow rates can be considered in the model.
In TRNSYS library for the absorption chiller there is the model Type107 which uses a catalog
data lookup approach to predict the performance of a single effect, hot water fired absorption
chiller. the calculation routine has been developed as Type107 in TRNSYS-15 during the work for
45
Chapter 3. Numerical modeling of solar thermal system
IEA TASK25 to improve simulations with absorption chillers in solar assisted cooling systems. In
this design, the heat required to desorb the refrigerant is provided by a stream of hot water. Since
this component rely on catalog data, the performance of the machine can be predicted and
interpolated within the range of available data but cannot be extrapolated beyond the range.
Because of the catalog data lookup approach, Type107 is not applicable over every range of inlet
conditions. In this component in the software catalog data are given but the lowest driving
temperature is 108.89 which is unrealistically high and would produce reflect on the end results.
Thus it was used the model Type 177 which is improvement from the Type 107 and it was develop
in “Technische Universität Berlin” by the Jan Albers.
The part load performance of thermally driven heat pumps or chillers is dominated by the
external driving temperatures of hot, chilled and cooling water. Thus heat transfer calculations are
done for constant flow rates, normally. Under these conditions the transmission of the heat
exchangers (i.e. their UA-values) can be assumed as constant and the characteristic equation
method can be applied easily to describe the part load behavior of the heat pump taking varying
external temperatures into account. With varying flow rates the assumption of constant heat
transmission is not valid anymore. Nevertheless, heat transmission ratios can be derived, which
implement the heat transmission variation into the method of characteristic equations. In addition,
the temperature difference ratio z of logarithmic to arithmetic mean temperature differences is
influenced by variable flow rates. Together with other well-known dimensionless numbers, such as
the dimensionless heat transmission (NTU) or the ratio of heat capacity flow rates, they are
implemented into an extended method of characteristic equations, which accounts for variable heat
transmissions and loss parameters.
There are four Modes in the Type 177, and each of them is describe as follows:
•
Mode 1 Standard mode which uses user supplied characteristic parameters. For some
absorption chillers the parameters are given in a table.
•
Special mode for the (old) Yazaki WFC-10 with a thermosyphon desorber. The (new)
Yazaki WFC-10 SE cannot be simulated with this mode.
•
Special mode for absorption chiller with variable solution flow rate
•
Physical mode which uses user supplied values for the heat transmissions (i.e. UAvalues) of all main heat exchangers (including the solution heat exchanger) and a fixed
solution flow rate
The internal temperatures of the four heat exchangers (Desorber, Absorber, Condenser,
Evaporator) can be combined by using Dühring’s rule for the saturation temperatures of
sorbens and sorbate at condenser and evaporator pressure.
46
Chapter 3. Numerical modeling of solar thermal system
(TD – TA) = (TC-TE) B
(35)
Assuming equal heat fluxes inside and outside the ab-or adsorption chiller (i.e. adiabatic heat
exchangers are assumed) the internal saturation temperatures TX can be expressed as a function of
the mean external temperatures tX.
TD = tD – QE KD/YD – QD,x/YD
(36)
TE = tE – QE KE/YE
(37)
TA = tA – QE KA/YA – QA,x/YA
(38)
TC = tC – QE KC/YC
(39)
The coefficients KX in equation (32-35) hold for the internal specific enthalpy differences at the
corresponding heat exchangers related to the specific enthalpy difference at the evaporator. YX
are the heat transmissions of the heat exchangers (YX = UX·AX).
Defining a characteristic temperature difference (∆∆t):
∆∆t := (tD – tA) – (tC – tE) · B
(40)
The characteristic equation for the cooling capacity:
QE = sE · ∆∆t – sE · ∆∆tminE
(41)
If the slope (sE – which contains the enthalpy and heat transfer coefficients) and axis interval
(sE·∆∆tminE) are constant, the part load of an ab- or adsorption chiller can be expressed as a linear
function of ∆∆t only. It has been shown that under steady state conditions and constant solution
flow rates or constant cycle times the assumption of constant parameters is permitted at many
times.
It will be presented the mathematical reference for the Mode 3 of this component since only
that was used in the simulations.
In mode 3 a modification of the extended method of characteristic equations described in [38]
is used. it is based on fitted characteristic parameters from measurements or manufactures data
sheets. The external flow rates of hot, chilled and cooling water have to be fixed at their rated
values during the simulation. The solution flow rate may vary if the parameters are given for
variable solution flow rates also. Two Pseudo-Dühring-Parameters (B*E1, B*E2) are used to define a
characteristic temperature difference.
∆∆t*E = tDi + (1 + B*E1) · tACi + B*E2 · tEi
The characteristic equation for the cooling capacity
47
(42)
Chapter 3. Numerical modeling of solar thermal system
QE = qE2 ·( ∆∆t*E )² + qE1 · ∆∆t*E + qE0
(43)
can be used either in quadratic or linear form. For the latter case parameter qE2 is set to zero. For the
driving heat the following equations apply:
∆∆t*D = tDi + (1 + B*D1) · tACi + B*D2 · tEi
(44)
QD = qD2 ·( ∆∆t*D )² + qD1 · ∆∆t*D + qD0
(45)
Basically the Pseudo-Dühring-Parameters in Equation (38) and (39) are different. Nevertheless
mean values can be used (carefully) in both equations to simplify the fit procedure (e.g. of
manufacturers data). The parameter description of the model are given in Appendix D.
B*1 = (B*E1 + B*D1) / 2
(46)
B*2 = (B*E2 + B*D2) / 2
(47)
3.2.6 Developing building numerical model
Energy requirements for space heating or service water heating can be calculated from basic
principles of energy conservation. One of the most significant barriers in precise determing the
buildings energy consumption is the lack of knowledge about the factors determining the energy
use. Building energy consumption is mainly influenced by six factors: (1) climate, (2) building
envelope, (3) building services and energy systems, (4) building operation and maintenance, (5)
occupant activities and behavior and (6) indoor environmental quality provided. The latter three
factors, related to human behavior, can have an influence as great as or greater than the former
three. The user related aspects and behavior effects can be seen from the large spread in energy use
for similar or identical buildings, but a distinction between the building-related and the user-rela
related energy consumption cannot be established.
For the building simulation is used the Type 56 from the TRNSYS library. Type 56 describes a
building with multiple thermal zones, i.e. rooms. The model uses data from wall and window
materials and thicknesses. Each room has a homogenous temperature, and radiation heat between
the rooms is based on the room area. Heat addition from solar direct and diffuse radiation is
calculated for each room depending on window and heat transfer properties. Type 56 models the
thermal behavior of a building divided into different thermal zones. In order to use this type, a
separate pre-processing program must first be executed. The TRNBuild program reads in and
processes a file containing the building description and generates two files (described later) that
will be used by the TYPE 56 component during a TRNSYS simulation.
The level of detail of this type is compliant with the requirements of ANSI/ASHARE standard 1402001. The level of detail of Type 56 also meets the general technical requirements of the European
Directive on the Energy Performance of Buildings. During the last two decades TRNSYS is widely
48
Chapter 3. Numerical modeling of solar thermal system
employed in building energy simulations [39-41]. There exist systematic studies comparing the
performance of this software against experimental results, as well as comparing the results from
TRNSYS to other industry standards for building energy simulation.
In modeling the air conditioning equipment in the Type 56 model , the "energy rate" method can be
used as a simplified model. The temperatures are set in advance for heating and cooling, set points
for humidity control, maximum cooling and heating rates. These specifications can be different for
each zone of the building. If it is required a more detailed model of the heating and cooling
equipment, a "temperature level" approach is required. In this case, separate components are
required to model the heating and/or cooling equipment. The outputs from the TYPE 56 zones are
used as inputs to the equipment models, which in turn produce heating and cooling inputs to the
TYPE 56 zones.
Type 56 needs a great amount of building data to calculate the thermal behavior of the building,
these include geometry data, wall construction data, windows data,…etc. in In the project
initialization firstly are entered some general information about the project, like orientations of
walls and windows required by the described building. Next step is defining zones in the building
with the boundary surfaces (walls) materials, windows types thus defining the heat transfer
coefficients. additional to weather data information such as: Radiation, ambient temperature,
humidity,…etc. furthermore, it needs information such as SCHEDUALE which may define the
gain from the occupants during the day with intervals representing the time being occupant from
the building owners
Short mathematical description of Type 56
All defined surfaces have thermal air nodes for which the calculations are performed. The system
boundary for this energy balance includes the inside surface node of all surfaces of the zone. This
balance deals with radiative and convective heat flow into and out of the airnode. The convective
heat balance is determined by the equation:
Qi = Qsurf ,i + Qinf,i + Qvent + Qg ,c ,i + Qcp lg,i
(48)
where:
Qsurf,I convective heat gain from inner surface of zone (because of temperature difference between
airnode temperature and surface temperature)
Qinf,I
infiltration gains (airflow from outside only)
Qvent,I ventilation gains ((air flow from a user-defined source, like an HVAC system)
Qg,c,I
internal convective gains (by people, equipment, illumination, radiators, etc.)
Qcplg,I gains due to (connective) air flow from airnode I or boundary condition, where
49
Chapter 3. Numerical modeling of solar thermal system
Q cplg,i = V ⋅ ρ ⋅ c p ⋅ (Tzone,i − Tair )
Reagarding the radiative heat fluxes to the airnode graphical presentation is given on Figure 23
Figure 23. Radiative energy flows considering one wall
This balance is determined by the equation:
Qr ,wi = Qg ,r ,i , wi + Qsol ,wi + Qlong ,,wi + Qwall − gain
(49)
Where:
Qr,wi
radiative gains for the wall surface temperature node,
Qg,r,i,wi radiative airnode internal gains received by wall,
Qsol,wi the solar gains through zone windows received by walls,
Qlong,I the long-wave radiation exchange between this wall and all other walls and
windows (εi =1)
Qwall-gain the user-specified heat flow to the wall or window surface.
The walls are modeled according to the transfer function relationships of Mitalas and
Arseneault defined from surface to surface (from outer to inner surface), which consider the wall
as a black box. For any wall, the heat conduction at the surfaces are:
⋅
nbs
ncs
nbs
⋅
q s ,i = ∑ b T − ∑ c T − ∑ d sk q s ,i
k
s
k
s ,o
k =0
⋅
nas
k
s
k
s ,i
k =0
ncs
(50)
k =0
nbs
⋅
q s ,o = ∑ a T − ∑ b T − ∑ d sk q s ,o
k =0
k
s
k
s ,o
k =0
k
s
k
s ,i
(51)
k =0
These time series equations in terms of surface temperatures and heat fluxes are evaluated at equal
time intervals. The superscript k refers to the term in the time series, and it specified by the user
within the TRNBUILD description. The coefficients of the time series (a’s, b’s, c’s, and d’s) are
determined within TRNBUILD program using the z-transfer function routines of literature.
50
Chapter 3. Numerical modeling of solar thermal system
A window is thermally considered as an external wall with no thermal mass, partially
transparent to solar, but opaque to long-wave internal gains. In the energy balance calculation of
the TYPE 56, the window is described as a 2-node model shown in Figure 24. Equation 50 is valid
for
a so = bso = cso = d so = U g , s
a sk = bsk = csk = d sk = 0
for k>0
Figure 24. Two-node window model used in th TYPE56 energy balance equation
For the star network approach a zone is restricted to a single airnode. The long-wave radiation
exchange between the surfaces within the airnode and the convective heat flux from the inside
surfaces to the airnode air are approximated using the star network. This method uses an artificial
temperature node (Tstar) to consider the parallel energy flow from a wall surface by convection to
the air node and by radiation to other wall and window elements.
A detailed window model has been incorporated into the TYPE 56 component using output data
from the WINDOW 4.1 program developed by Lawrence Berkeley Laboratory, USA [42]. This
window model calculates transmission, reflection and absorption of solar radiation in detail for
windows with up to six panes. External and internal shading devices and an edge correction for
different glazing spacer types are considered. Each glazing absorbs and reflects a part of the
incoming solar radiation depending on the glazing material and the incidence angle. In the program
WINDOW 4.1, the detailed calculation of reflection between the individual panes and the
absorption and transmission of each pane is performed hemispherically for diffuse radiation and in
steps of 10° incidence angle for direct solar radiation. Together with the thermal properties of the
gas fillings and the conductivity and emissivity of the glazings, the optical data for the window is
written to an ASCII file by the WINDOW 4.1 program. This output file has a standard format,
which makes the results available for TRNSYS.
The optimum capacity of an energy storage system depends on the expected time dependence
of solar availability, the nature of loads to be expected on the processes, the degree of reliability
needed for the processes, the manner in which auxiliary energy is supplied and an economic
51
Chapter 3. Numerical modeling of solar thermal system
analysis that determines how much of the annual load should be carried by solar and how much by
the auxiliary energy source.
The two principal materials used for storage are water for liquid-based systems and rock pebbles
for air-based systems. Water has a high heat capacity and rocks have one-fifth as much (one third
per unit volume) and both are inexpensive. The thermal energy storage (TES) system is one of the
most appropriate methods of correcting the mismatch that occurs between the supply and demand
of energy. Heat can be stored in sensible/latent form, and by thermo chemical techniques. In
sensible heat storage (SHS), thermal energy is stored by raising the temperature of a solid or liquid.
SHS systems utilize the heat capacity and change in temperature of the material during the process
of charging and discharging. SHS characterized by temperature variation is a simpler technique, but
occupies a larger volume.
The latent heat storage is based on heat absorption or release when a phase change material (PCM)
undergoes a phase change. The latent heat storage by PCM in comparison with SHS, possesses a
greater density of stored energy and operates in a narrower operational temperature range. Zalba
have reviewed various aspects of latent heat storage systems, such as PCMs heat transfer and
applications. PCMs are advantageous for the dynamic and static storage of thermal energy as they
absorb and release large amounts of energy at specific temperature [34].
Storage mechanisms have been researched quite intensively in the frame of Task 32 [13] of the EA
(International Energy Agency) SHC (Solar Heating and Cooling) program. A new joint IEA project
[14], involving SHC Task 42 [15] and ECES (Energy Conservation through Energy Storage),
concentrates on the storage materials involved.
3.3 Modeling solar thermal assisted air-conditioning system
Assessment of thermal performance of the solar air-conditioning system is performed through a
dynamic simulation model with transient behavior implemented via thermal and mass storage terms
as well as delay times. The model ie analyzed system generally consists of four main subsystems
shown in Figure 25, as follows:
1. First subsitem composed of solar collectors with complete hydraulic fittings and control
- differential controllers, plate heat exchangers ie this system is represented the source
of thermal energy for heating or thermal enrgy for driving the cooling the absorption
machine
2. Second is the subsystem for hot and cold storage which includes the storage tanks for
hot / cold water that actually represents the connection between the heating system in
the building ie absorpcionata cooling machine and the source of heat.
52
Chapter 3. Numerical modeling of solar thermal system
3. The heating system introduced with heating / cooling devices, hydraulic armature heat
exchangers and cooling absorption machine and eventually existing conventional
sources of heat and / or cooling energy.
The fourth subsystem is the consumer of thermal energy ie the building . This system is
represented by the thermal characteristics of the object, ie its orientation in space .
Figure 25. Subsystems definition of solar cooling system
Figure 26. Functional scheme presenting inter conections between components of the
system moddeled in TRNSYS
The next picture shows the functional scheme according to which is made numerical modeling
of the analyzed system in TRNSYS using previously described and validated components.
On Figure 26 is presented the analyzed solar assisted air-conditioning system. The main system
components are: the solar collector array, two storage tanks, auxiliary heater, absorption chiller and
the energy consumer i.e. the building which also incorporates the heating/cooling system
components.
53
Chapter 3. Numerical modeling of solar thermal system
The working fluid from the solar collectors indirectly through heat exchangers is used to heat
the domestic hot water in tank 3 or heat the fluid in the storage tank 4 further used as part of the
heating energy in the building or part of the driving heat for the absorption chiller in summer. The
circulation of the solar collectors working fluid for the storage tanks 3 and 4 is done by two
separate circulating pumps P1 and P2, controlled by two differential controllers having mutual
predefined control function further explained. First condition for the pumps to be switched on is the
temperature difference between the collector outlet temperature and the fluid temperature in storage
tank (3 or 4) to be greater than the set upper dead band. The control logic for switching between the
two tanks is solved using the two controllers Type 2b (K1 and K2) one flow diverter Type 11f . The
advantage has the controller K2 of the tank 4 i.e. the initial input control signal (on/off) for the
controller of the DHW tank K1 is received from the controller K2 i.e. when the controller K2 is on,
then the controller K1 is off.
The collector(s) thermal efficiency in the simulation is determined using the equation
component from the TRNSYS library. The equation considers ratio between the useful energy gain
from the all od the collectors transferred to the fluid and the total tilted radiation for the collector
surface. The data for the quantity of useful energy gain and total radiation in the equation is read
from the quantity integrator which integrates these values in the predefined period defined from the
required value period thermal efficiency and energy i.e. daily, weekly, monthly, yearly or any other
time interval.
Sub-system 1 – solar collectors
The solar collectors are connected in array where the number of serial and parallel connections
is subject for further analysis in this work. As mentioned previously two collector types are used:
flat plate collectors produced in Camel Solar type CS Full Plate 2.0-4 with technical characteristics
given in Table 1, and the other type are the vacuum tube collectors with technical data given in
Table 2.
Solar collectors parameters are varied and analyzed in order to estimate their impact on the total
collected energy overall thermal efficiency and the solar fraction. Parameters analyzed in this work
regarding solar collectors are:
•
mass flow rate
•
azimuth (orientation) of solar collectors ( 0° south, 90° West, -90° East one and two axis
tracking, by azimuth and/or altitude )
•
solar collector area – gross area of the array used in simulations 16 m2, 32 m2, 64 m2
•
slope (tilt) angle of collector installation
54
Chapter 3. Numerical modeling of solar thermal system
Table 1. Technical data for collector type Camel Solar Flat plate 2.0-4
Dimensions L x W x H
Ansorber aperture area
Absorptance, α
Emmitance, ε
Transmitance
Conversion factor of the beam
irradiance, F'(τα)en
Factor to determine the incidence
angle modifier of the beam
irradience, bo
Optical efficiency, ƞ o
Heat transfer coefficient a1
Temperature depending heat
transfer coefficient a2
Incidence angle modifier diffuse
radiaton Kθd
Incidence angle modifier Kθ = 50°
Area related heat capacity c
Volume flow rate,
Peak power per collector unit
2
G=1000 W/m
mm
m²
-
2005 x 1005 x 85
1.83
0.95
0.05
0.92
-
0.795
-
0.138
-
0.791
2
4.176
W/m K
2
W/m K
2
0.008
-
0.988
0.935
2
kJ/m K
l/m²h
13.19
72
W
1448
The differential controller settings are analyzed for their influence over the thermal efficiency,
pump energy consumption and the solar fraction. In the controller, with the settings are considered
the values for upper and lower dead band, parametrically analyzed in order to estimate their
influence over system efficiency, electrical energy consumption and solar fraction. Since the
storage tanks are modeled as stratified another parameter of interest is the position of the
controllers probe (height) measuring the tank temperature.
All of the circulating pumps are modeled as constant flow pumps.
Second sub-system are the storage tanks as mentioned one is for DHW (3), other is for the
heating system and driving the absorption chiller. The tank for DHW through all of the simulations
has constant volume of 200l. The volume of storage tank 4 has been varied in the simulations in
order to account for its influence on the system thermal efficiency and solar fraction. The storage
volumes used in simulations are: 500l, 1000 l, 1500 l and 2000 l all of them are modeled with
internal heat exchangers .
Technical data for the storage tanks used as input values in the model, like dimensions, heat
exchanger data etc. are given in Appendix A
55
Chapter 3. Numerical modeling of solar thermal system
Another component is the auxiliary heater which adds energy to the fluid if the set temperature
is not reached with the solar heat. Type 6 is used as auxiliary heater for the DHW while for the
storage tank 4 is used Type 659 from the Tess library. Input parameters are the rated capacity,
specific heat of the fluid and set point temperature while the heat loss to the surroundings is
neglected. In most of the analyzed simulation scenarios the position of the auxiliary heater is
outside of the storage tank, only one analysis is performed to assess the influence of the position of
the auxiliary heater (inside or outside the storage tank) regarding the energy consumption.
The cooling energy is delivered by the absorption chiller modeled with Type 177 i.e. it is
modeled LiBr/H2O absorption chiller product of Sonenklima Suninverse 10, technical data given in
Table 4. Cooling tower is modeled with component Type 510 closed circuit cooling tower, used to
cool a liquid stream by evaporating water from the outside of coils containing the working fluid.
The working fluid is completely isolated from the air and water in this type of system. Technical
data are from cooling tower product of Baltimore Aircoil Company model details given in Anexx 1.
In the last sun-system is presented the energy consumer i.e. the building. Multi zone Type 56
is used to model the thermal behavior of a building having multiple thermal zones. Within the same
model are defined heating/cooling transfer devices such as underfloor heating and ventilation for
the cooling part. Also for as cooling heat transfer device is used heat exchanger water-air modeled
with Type 508a.
Internal (room) temperature in the building is maintained with thermostat Type 1502. The
thermostat controller has commands for first stage heating at cool fluid temperatures, second stage
heating at cooler fluid temperatures, and third-stage heating at even lower fluid temperatures. There
is option to disable first stage heating during second stage and third stage heating, and to disable
second stage heating during third stage heating.
In many heating applications, a desired fluid temperature may depend on the time of day or the
day of the week. This variation of the heating set point temperatures are modeled here using an
optional set back control function and a setback temperature difference.
56
Chapter 4
4. System and Components validation
4.1 Solar circuit component validation
In this part are presented and compared the results between the measurements of preassembled
experimental solar collector system presented on
Figure 27. Scheme of the experimental installation and the results from the same system are
modeled and simulated in TRNSYS. The validation is performed for the solar collector, storage
tank and differential controller components.
The system has one flat plate solar collector connected with the internal heat exchanger of the
storage tank. Control is provided by differential controller which is set to turn the circulation pump
on when the temperature difference between the collector outlet temperature and the tank
temperature is greater than five. The water from the storage tank is not discharged and the electric
heater is turned off during the measurements. The fluid (water) flow rate is set to 7,5 lit/min.
Figure 27. Scheme of the experimental installation
The measurements are made on an hour interval for the fluid inlet T1 and outlet T2
temperatures from the solar collector, tank fluid temperature T3 and the solar radiation measured
57
Chapter 4. System and Components validation
with the pyranometer S. The experimental setup of the analysed solar thermal system is placed in
Skopje, R.Macedonia northen latitude of 42° and 21.43° east longitude.
The solar collector is evacuated tubular direcrt flow product of Camel Solar type Vacumm CS
15 Solar KeyMark certified. It is placed under tilt angle of 45° with south orientation i.e. azimuth
angle of 0°.
The collector thermal performance test results made according EN 12975 are
presented in Table 2
Table 2. Technical data for collector type Camel Solar Vacumm tube SC 15
Dimensions L x W x H
Number of asborber tubes
Absorptance, α
Emmitance, ε
Conversion factor of the beam
irradiance, F'(τα)en
Factor to determine the incidence
angle modifier of the beam
irradience, bo
Optical efficiency, ƞ o
mm 1990 x 1180 x 158
15
0.92-0.96
0.04-0.06
0.695
-
0.138
-
0.738
2
Heat transfer coefficient a1
Temperature depending heat
transfer coefficient a2
Incidence angle modifier diffuse
radiaton Kθd
Incidence angle modifier Kθ = 50°
1.725
W/m K
2
W/m K
2
0.01
-
1.203
0.935
2
kJ/m K
l/m²h
m²
Area related heat capacity c
Volume flow rate,
Apperture area per collector unit
Peak power per collector unit
58.4
72
1.42
W
2
G=1000 W/m
-
-
1048
⋅
The area based collector power q was modeled according the equation:
⋅
q = F ' (τα ) en Kθb (θ l ,θ t )Gb + F ' (τα ) en Kθd (θ )G d −c1 (Tm − Ta ) − c2 (Tm − Ta ) 2 − c5
With
K θb (θ l ,θ t ) = K θb (θ l ,0) ⋅ K θb (0,θ t )
Where
Gb , Gd
W/m2
beam and diffuse solar irradiance
c1
W/m2K
heat transfer coefficient
c2
W/m2K
temperature depending heat transfer coefficient
c5
W/m2K
temperature depending heat transfer coefficient
58
dTm
dt
(52)
Chapter 4. System and Components validation
Tm
K
mean fluid temperature inlet/outlet solar collector
Ta
°C
ambient temperature where the collector is installed
Storage tank technical specification is presented in the Table 3
Table 3. Storage tank technical details
Capacity
Height
Diameter
Insulation, rigid PU
Coil capacity
l
H, mm
D, mm
mm
l
Heat exchanger surface
Prolonged power according
DIN 4708 80/60/45
NL-power coeficient at 60°C
Coil outlet
Cold water inlet
Sensor sleeve for thermostat
Coil inlet
Hot water outlet
m
kW
m³/h
L, mm
A, mm
G, mm
K,mm
E,mm
2
150
1210
560
50
4.56
0.74
25
0.61
2.5
202
202
822
592
868
The TRNSYS model components with their interconnections schematically are given on
Figure 28.
Figure 28. TRNSYS model for the experimental installation of solar thermal collector system
In the TRNSYS model for the solar collector model is used the collector Type 538 from the
Tess library modeled with the technical data given in Table 2. The storage tank is modeled with
the Type 60d including the internal heat exchanger for which are supplied data from Table 3. Type
2b-2 is used for the differential controller with upper dead band of 5 and lower dead band 2, the
high limit cut-off temperature is set to 100 °C. Between the solar collector and storage tank is
connected pipe Type 31 modeled with internal diameter 0.0025 m, length of 10 m and loss
coefficient of 0,3 W/m2K to account for the heat losses. The pipe Type 31 beside to account for the
59
Chapter 4. System and Components validation
heat losses in the pipes also is used in order to increase the thermal capacity of the system and thus
increase the simulation stability. Also for the circulating pump is used the Type 3d with mass flow
rate 450 kg/h i.e. 7,5 l/min same as in the experimental setup.
Measurements are performed starting from date 18.09.2013 until 28.03.2014 and in parallel are
measured two systems with same capacity storage tank of 150l but different type of collectors i.e.
flat plate and vacuum tube solar collectors. In the validation process are used the data for the
vacuum tube collector and the results from only one day period (18.09.2014) with collection time
interval ranging between 20min and 45min interval, starting from 10:40 until 16:05 h.
There is possibility in the TRNSYS software to be inputted specific measured values of the
system as the solar radiation but in this case there was no possibility since the measured values for
the sun radiation did not provide data for the total the beam and diffuse components which are
required by the model. Thus, in the simulations for the solar radiation data was used the weather
component from the TRNSYS library the Type 15 which supplies input data in the solar collector
numerical model: ambient temperature, beam, sky and diffuse radiation for the tilted surface
(calculated regarding the tilt angle of the collector), solar zenith and solar azimuth angle. The
weather data in this model are generated in a so called referent year which contains data based on
stochastic methods, interpolations where the data for temperatures, wind speed are for the period
between 1961 – 2009, while for the sun radiation are for the period 1985 – 2005. In determining
this kind of reference year, the typical range of meteorological measurements at hourly intervals are
required for a period of several years, a process which results in a complete picture of the climatic
conditions that govern the examined area. But this does not mean simply determing the average of
all years, because it does not adequately predict the changes that may occur, but is selected
representative month for this area. The procedure is as follows: for each month is determined
average solar radiation over the entire period of measurement and individual monthly average
radiation for each year within the period considered. Monthly value to the average radiation closest
or equal to the global monthly average over the period of measurement is chosen as a representative
month for typical reference year. This process is repeated for each month of the year where then
grouped the selected month and provides hourly average values over the year.
As mentioned previously measurements are performed for the inlet/outlet fluid temperature
from the solar collector, inside storage tank temperature and the total solar irradiation.
In the following graphs are presented the result from the comparison between the
measurements and simulation data:
60
Chapter 4. System and Components validation
65
Temperature, °C
60
55
50
45
Measure T1
40
Simulated T1
35
30
11:15 12:20 13:10 13:45 14:25 14:13 16:05
Time [h:min]
Temperature, °C
Figure 29. Measured and simulated temperatures for the collector inlet
70
65
60
55
50
45
40
35
10:43
12:00
12:30
13:30
14:05
14:43
15:35
Time [h:min]
Measured T2
Simulated T2
Figure 30. Measured and simulated temperatures at the collector outlet
65
60
55
50
45
40
35
30
11:15
12:20
13:10
13:45
Measured T3
14:25
14:13
16:05
Simulated T3
Figure 31. Measured and simulated temperatures inside storage tank
61
Global solar radiation, W/m²
Chapter 4. System and Components validation
1200
1000
800
600
400
200
0
Time [h:min]
Measured radiation S1
Simulated radiation S1
Figure 32. Hourly measured and simulated solar radiation for the specific day
From the above presented data with the diagrams can be concluded that there is acceptable
match between the measured and simulated results. The discrepancies that appear between the
temperatures of the experimental and simulated results are expected since firstly the solar radiations
are measured and simulated which differences clearly appear on Figure 32.
. Another influencing factor are the uncertainty of the measurements error and last but not the
least we should consider the transition nature of the solar thermal systems.
The resulting simulations reveal the individual thermal behavior of the solar collector, storage
tank, differential controller and circulating pump as well as their assembled thermal behavior.
These results were very close to their corresponding experimental data and this fact validates these
models for future application in the heating/cooling system.
4.2 Absorption chiller validation
Validation for the absorption chiller is made for the TRNSYS component Type 177, further in
detail described.. This component type offers four numerical modes of absorption chiller, and in
this simulation is used the mode “a” i.e. Type177a which is standard mode using user supplied
characteristic parameters. Since in this thesis will be analyzed solar air-conditioning for residential
buildings, in Table 4 are given the technical data for several small absorption chillers. From the
presented absorption chillers for the need of the simulation it is selected the absorption chiller
H2O/LiBr produced by Sonnenklima type Suninverse 10.
Numerical modeling for the selected chiller Suninverse is performed by providing the
characteristic parameters in the TRNSYS Type 177a. The required characteristic parameters for
modeling the absorption chiller within the Type 177a are given in Appendix D with data for several
other commercial chillers.
62
Chapter 4. System and Components validation
Table 4. Technical data for different market available small absorption chillers
Company
Yazaki
EAW
Sonnenklima
Rotarica
WFC-SC5,
Product name
chillii WFC 18 Wegral SE 15
Suninverse 10
Solar 045
Technology
Absorption
Absorption
Absorption
Absorption
Working pair
H2O/LiBr
H2O/LiBr
H2O/LiBr
H2O/LiBr
Absorption chiller
image
Cooling capacity, kW
Heating temperature,
°C
Recooling temperature,
°C
Cold water
temperature, °C
COP
Dimensions (WxDxH),
m
Weight, kg
17.6
15
10
4.5
88 / 83
90/80
75/65
90/85
31 / 35
30/35
27/35
30/35
12.5 / 7
0.70
0.60 x 0.80 x
1.94
420
17/11
0.71
1.75 x 0.76 x
1.75
660
18/15
0.77
1.13 x 0.80 x
1.96
550
72
300
120
13/10
0.67
1.09 x 0.76 x
1.15
290
1200
(incl.ventilator)
Electrical power, W
In the component Type 177a as input parameters are taken the values for Suninverse provided
in Table 4 for which with the simulation as output cooling power is obtained value of 10,1 kW
which corresponds with the factory value. Thus it can be concluded that this model of absorpbtion
chiller provides reliable results and can be used further in simulations.
Validation exists for the Type 177 mode “d” performed by Albers and Ziegler (2011) using the
measurement results from Kühn.
The supplied characteristic parameters in the validation of Type 177d are:
Symbol description:
γ
a
- flow rate ration
- ration of heat transfer coefficient
⋅
Q
- heat power
kW
⋅
m
- mass flow rate
Y
- effective heat transmition
T,t
- internal, external temperature
∆∆t
- characteristic temperature function
Sub- superscript
63
kg/s
kW K-1
°C
K
Chapter 4. System and Components validation
A
C
D
X
-
absorber
condenser
desorber
placeholder for component
ext, int
i,o
E
- external, internal
- inlet, outlet
- evaporator
An extensive measurement campaign has been carried out by Kühn et al. (2005, 2007) for the
investigation of a 10 kW chiller. Here two groups of measurements are used for the verification of
the extended calculation method including variable external flow rates:
Measurements under design conditions, i.e. inlet temperatures and flow rates meet the rated values
(presented in Table 5and Table 7. Legend of the performance indicators and framed circles in
Figure 34). Also measurements are performed at design inlet temperatures but with non-design
external flow rates. The range of flow rate variation was between 0.3 < γD < 2.3 of the rated value
for hot water and between 0.5 and 1.5 for chilled and cooling water (γE and γAC) (see circles in
Figure 34).
The first group of approx. 20 independent measurements of Kühn (2005) at rated conditions has
been used to determine the ratios aX = αext,X,0 /αint,X,0: First the external (i.e. tube inside) heat
transfer coefficients αext,X,0 have been calculated according to Gnielinski (2006). Then measured
values of QX and ∆Tlog,X have been used together with geometric data of tube bundles to determine
αint,X,0. The resulting ratios of aX are given in Table 5 and depicted in variation of inlet temperature
and flow rate of hot water (left) and cooling water (right)
Figure 33 as open symbols in together with their uncertainty which is not larger than ±5% to ±10%.
Table 5. Averaged measured performance data by Kühn (2005) of a 10kW absorption chiller
X
mX,ext,0
tXi
tXo
TX
QX
YX,0
αX
D
kg/s
0.33
°C
74.9
°C
65.1
°C
62.8
kW
13.4
kW/K
1.9
3.8
E
0.81
18.0
14.9
12.9
10.5
3.0
1.8
C
0.72
31.0
34.9
36.0
11.7
3.7
0.8
A
0.72
27.0
31.0
36.6
12.1
1.7
5.7
S
0.07
31.5
60.8
52.1
4.6
0.8
-
Unit
64
Chapter 4. System and Components validation
Figure 33. Measured ratios of external to internal heat transfer coefficients
Figure 34. Measured cooling capacity for design and variable flow rates, by Kühn et al.
2005&2007.
Accordnig the previous presented from the Kühn can be concluded that the Type 177d
modelled according the data in Table 5 provides reliable and acceptable output results, further
enabling this model to be used in simulation of solar air-conditioning systems.
65
Chapter 5
5. Performance evaluation of solar air-conditioning system
It has become recognized that, however, solar heating is the product of a collection of
components comprising a system and needs to be studied such. Because of the interactions of
components, optimal system performance occurs under conditions different from those for optimal
behavior of each component. For example optimal collection efficiency would not necessarily be
coupled with least auxiliary energy.
In this chapter are presented two schemes for quantification of performance level of a solar
heating and cooling plant.
Many different hydraulic schemes are designed which makes difficult to compare the
installations performances [43]. Methods used to determine solar heating and/or cooling energy
requirements for both active and passive/hybrid systems are described by Feldman and Merriam,
Hunn, Nowag and other. In the frame of IEA-TASK 38 a unified monitoring procedure has been
developed. For thermally driven systems the scheme on Figure 35 is used to identify main
components and energy flows of the system. On Figure 35 is presented small scale system for
family houses, small multi dwellings, using a small size packaged ab/adsorption solar system. This
configuration is an adaptation of the solar combi system including the cooling function, also called
SSC + Solar Combi.
66
Chapter 5. Performance evaluation of solar air-conditioning system
Figure 35. Solar cooling system variables [44]
The description of each variable of the system is presented in Table 6
Table 6. Energy and thermal flows of solar cooling systems
Label
E1
E2
E3
E4
E5
Electricity consumer, kWh
Heating system
pump collector field (primary loop)
pump collector field (secondary loop)
pump boiler hot-storage (including internal boiler consumption)
pump hot-storage to space heating (SH)
pump hot-storage to domestic hot water (DHW)
E6
E7
Cooling system
pump hot-storage to cooling machine
pump cooling machine (ACM) to cooling tower
E8
E9
pump cooling machine (ACM) to cold storage
pump cold storage to cold distribution
E10
E11
pump back-up source -cold storage
absorption/adsorption cooling machine (ACM)
E12
E13
E14
E15
compression chiller (back-up system)
pump compression chiller to fan (back-up system)
fan, cooling tower
fan of compression chiller (back-up system)
Water treatment System
E20
water treatment for wet cooling tower
Thermal flows, kWh
Qsol
solar irradiation on total collector aperture area
Q0
collector solar thermal output
67
Chapter 5. Performance evaluation of solar air-conditioning system
Q1
solar thermal output to hot storage
Q1s
heat output from hot storage
Q2S
boiler thermal output (fossil) into storage
Q2D
fossil boiler thermal input bypassing hot storage (directly used)
Q3a
space heating (SH) consumption (conventional)
Q3b
space heating (SH) consumption (ventilation system)
Q4
domestic hot water consumption (DHW)
Q6a
hot storage input to cooling machine (ACM)
Q6b
hot storage input to DEC-system (sorption regeneration)
Q7
cold output ACM to cold storage
Q8
cold output back-up chiller or free cooling to cold storage
Q10a
cold storage output to cold-distribution
5.1 Performance indicators
There are four generally accepted measures of solar system performance:
1. Collector efficiency applies to the performance of the solar energy collection subsystem. It is the
energy collected, divided by the radiation incident upon the collectors. The radiation may be the
total radiation, or it may be only that incident while the collection subsystem is operating.
2. System efficiency, or solar heating performance factor is the solar heat delivered to the load
divided by the total radiation incident upon the collector. It is similar to the collector efficiency but
also takes into account heat loss from the pipes and storage. The net system efficiency is the solar
heat delivered less the electrical inputs to the system, divided by the incident solar radiation. This
figure must be used however with the caution that the delivered heat may not have the same
economic value as the equivalent energy in the form of electricity.
3. Solar fraction is the fraction of the total heat requirement that is met by solar energy. The figure
relates the output of the solar system to the size of the system as well as on its efficiency.
4. Electrical coefficient of performance is the solar heat delivered to the load, divided by the
electrical energy used to operate the system.
Each solar system operates at characteristic efficiency level resulting from the interaction of the
subsystems, environmental conditions and system configurations. The net savings per square meter
of solar collector indicate the relative performance of each of these systems.
The five categories of system-level design parameters that limit solar system performance. These
are:
1. Solar resource assessment. This category represents the solar reference weather data values used
by the solar design community
68
Chapter 5. Performance evaluation of solar air-conditioning system
2. Collection subsystem. This category represents the solar collection sub-system, including devices
used to capture incoming solar radiation
3. Storage subsystem. This category deals with all aspects of the system effects caused by storage
components.
4. Controls. This category refers to equipment and methods for controlling solar components within
the solar system.
5. Load. This category deals with the types and magnitude of the heat requirements in the
buildings.
Also in order to calculate the different performance indicators, some necessary values are defined
: the useful solar energy (ESU), the parasitic electricity demand of the whole system (Eaux) and of
the solar part (Eaux sol), the thermal losses of the hot and cold storage (Qloss HS and Qloss CS),
the thermal losses of the hot storage due to the heating backup system (Qloss HB), the thermal
losses of the cold storage due to the cooling backup system (Qloss CB), the final energy
consumption of the heating backup system (ConsHB) and of the cooling backup system (ConsCB).
Each of this quantities are defined in Table 7.
Table 7. Legend of the performance indicators
Nomenclature
Unit
Qsol
Global irradiation on collector area
kWh
kWh
Q1 - Q10
E1 at E14
Thermal energy defined according to FIGURE
Auxiliary electrical consumptions defined according to Figure
kWh
kWh
V1
Eaux & Eaux sol
Qloss HS & Qloss CS
Water consumption of the heat rejection system
Parasitic electricity demand of the whole system and of the solar part
Thermal losses of the hot and cold storages
m3
kWh
kWh
Qloss HB & Qloss CB
Thermal losses of the storages due to the hot and cold backup system
kWh
ESU
Useful solar energy
kWh
RgHB & RgCB
Generation efficiency of the hot and cold backup system
-
ConsHB & ConsCB
Hot and cold backup system energy consumption
kWh
ηHS & ηCS
COPth
Hot and cold storage efficiency
Thermal coefficient of performance of the sorption chiller -
-
PER
Primary Energy Ratio
-
εx
Primary energy conversion factors of the energy X
-
€X
Cost per kWh of the energy X
€/kWh
Rcoll & Rsol
PSU
Collector thermal yield and solar thermal efficiency of the system
Useful solar thermal productivity
kWh/m²
COPelec sol
Electrical coefficient of performance for solar energy
-
WCspe
Specific water consumption of the system
l/kWh
69
Chapter 5. Performance evaluation of solar air-conditioning system
kWhcost
Operation cost of the system
€/kWh
Iconf, Idata & Ifct
Comfort, monitoring data lost and functioning indicators
%
Thot, Tcold
Reference hot and cold storage temperatures
°C
Tamb, Tav, Text
Reference ambient, collector average and external temperatures
°C
Nmonth, Ncool
ENS
Number of months in functioning and cooling mode
Average annual daily value of the irradiation on the collector kWh
Minimum, maximum, value and percentage value of the considered
indicator
kWh
Imin, Imax, I, I%
-
In some cases the solar fraction can be used as performance index defined as a fraction of
required heat for a certain application which is delivered by the solar system. However this
parameter is difficult to judge in some cases since it does not reflect the full picture of the energy
balance. Particularly for solar cooling systems in which different energy sources may serve as a
back-up, it may be difficult to define the solar fraction properly. Therefore since estimation of
primary energy is the main goal it is recommended to use the corresponding parameter to quantify
the energy performance of a solar air-conditioning assisted system. For complete system
performance assessment it is necessary to consider the energy consumption for the entire year.
5.3.1 Thermal efficiency indicators
Losses are inferred from energy flow balances computed for system components. Thus for
example the reflection and back-radiation loss from the collector array is calculated as the
difference of the solar energy incident on the array and the collected solar energy. Any losses that
result in space heating will be indicated in
The thermal efficiency indicators describe the main thermal losses of the system through the hot
(1) or (2) and cold (3) storage and the thermal coefficient of performance of the chiller (4).
5.3.2 Global performance indicators
The global performance indicators represent the overall system performances and take into
account the solar energy use as well as the heating and cooling backup energy use. The global
performance indicator defined is the primary energy ratio. The primary energy savings per unit
solar collector area give an indication of the contribution of each square meter of collector field to
the energy saving of the entire system.
In the Table 7 are summarized and presented the calculation methods for different performance
indicators.
70
Chapter 5. Performance evaluation of solar air-conditioning system
Table 8. Calculation procedures for solar thermal system performance indicators
Indicators
Calculation method
ηhs = 1- QlossHS / (Q1+ Q2s), with Qloss HS = Q1- Q3a- Q4- Q6a
(53)
ηcs = 1- QlossHS / Q1, with Qloss HS = Q1+ Q2s- Q3a- Q4 - Q6a
(54)
COPth = Q7 / Q6a
(55)
PER = (Q10a + Q3a + Q4) / (Eaux + εelec + ConsCB x εCB + ConsCB x εCB), with
ConsCB = Q2 / RgHB and ConsCB = Q8 / RgCB
(56)
Rcoll = Q1 / Qsol
(57)
Rsol = ESU / Qsol
(58)
PSU = ESU / Scoll
(59)
COPelec sol = ESU / Eaux sol
(60)
WCspe = V1 / Q7
(61)
kWhcost = (Eaux x €elec + ConsSCB x €CB + ConsSHB x €HB + V1 x €water) / (Q10a + Q3a +
+ Q4)
(62)
ESU = Q3a + Q4 – Q2s + QlossHB + (Q10 – Q8 + QlossCB) / COPth ; with
QlossCB = QlossCS x Q8 / (Q8 + Q7) and QlossHB = QlossHS x Q2s / (Q1 + Q2s)
(63)
Eaux = E1 + E2 + E3 + E4 + E5 + E6 + E7 + E8 + E9 + E10 + E13
(64)
Eaux sol = E1 + E2 + ∑cool (E6 + E7 +E11 + E14) x Q1 / (Q1 + Q2a)
(65)
71
Chapter 6
6. Simulation results and analysis
In this chapter are analyzed multiple scenarios based on which are derived and adopted
conclusions on the impact of individual parameters on the system performance, useful energy yield,
solar fraction and consumption of primary energy. The analysis can be generally divided into two
parts: analysis of the system operating in heating mode and second part in cooling mode.
6.1 Reference building modeling and simulation
Building as energy consumer has a major impact on the overall efficiency of the solar system
i.e. can be freely said that the building itself is one of the leading parameter in sizing the system.
Since the analyzes are made for climatic conditions in Macedonia also the thermal performance of
buildings must be in accordance with the Regulations on energy efficiency in Macedonia.
Furthermore the analysis is taken into account the impact of the specific consumption of heating /
cooling energy of the building kWh/m2 a to the response and the performance of the solar collector
system. Main indicators based on which system comparison is base are: thermal efficiency of solar
collectors, solar fraction and power consumption for the auxiliary devices.
Thus in the simulation / analysis systems are analyzed in conjunction with a reference building
with three different specific energy consumption i.e. three types of building are defined.
The building has one floor with a total conditioned area of 150 m2. In Table 10 are given data
for. surfaces and orientation of exterior walls, windows, floor, roof and coefficients of heat
transfer.
In Table 9 are listed three types of the building i.e. the dimensions and orientations are
unchanged only the insulation thickness is varied in order to obtain different values for specific
annual consumption of thermal energy. The main motive for variations in the thickness of the
insulation is to analyze the influence of the thermal performance of buildings on the economic
viability of the use of solar thermal systems in air-conditioning.
72
Chapter 6. Simulation results and analysis
Table 9. Reference building physical and thermal performance data
Surface
Out.wall 1
Windows 1
Out.wall 2
Windows 2
Out.wall 3
Windows 3
Out.wall 4
Windows 4
Floor
Roof
Window type
Windows solar heat
gain coefficient;g-value
Building I Building II Building III
Orientation
Area, m²
U value, W/m²K
North
42
0.58
0.33
0.18
North
3
1.40
1.40
1.40
East
25.5
0.58
0.33
0.18
East
4.5
1.40
1.40
1.40
West
25.5
0.58
0.33
0.18
West
4.5
1.40
1.40
1.40
South
42
0.58
0.33
0.18
South
3
1.40
1.40
1.40
150
0.33
0.33
0.24
150
0.54
0.42
0.35
Double glazed TRNSYS library (w4-lib data)
0.589
Insulation Insulation Insulation
5 cm
10 cm
20 cm
Out.wall construction 2 x Plaster 2cm, brick 25cm
Granite tile 6cm, cement
Insulation Insulation Insulation
mortar 5cm , concrete slab
10 cm
10 cm
15 cm
20cm
Concrete slab 20cm, hydro
Insulation Insulation Insulation
Roof
isolation, cement mortar 5cm
15 cm
20 cm
25 cm
Outside convective heat transfer
αout = 25 W/m²K
coefficient
Inside convective heat transfer
α in = 7,7 W/m²K
coefficient
Floor
Constant value of 0.3 1/h is defined for the infiltration of outdoor air, while for the summer
when cooling is required in the building is envisaged/modeled mechanical ventilation defined with
air mass flow and temperature entered through \the models of fan and heat exchanger air-water
which is directly connected with the cooling absorption machine.
Regarding the thermal comfort, in the heating mode the inside temperature is defined to be 20
°C from 05:00 – 22:00 and for the rest is defined setback temperature of 16 °C, for the cooling
mode is defined constant inside temperature of 26 °C.
Calculation of energy consumption in the building is obtained directly as output size of the
numerical model of the object in kJ / h value which further is integrated for the required period with
the quantity integrator. Also as output parameters of the model is the output temperature of floor
heating, the temperature of the air entering the fan to the heat exchanger and air-water and the
delivered energy from the underfloor heating system into the building.
Monthly analysis is performed for the building heat energy consumption regarding different
heat transfer coefficients i.e. different wall, floor and roof isolation thickness thus defining three
types of Building I, II and III. as presented in Table 9.
73
Chapter 6. Simulation results and analysis
3500.00
3000.00
kWh
2500.00
2000.00
Building I
1500.00
Building II
1000.00
Building III
500.00
0.00
1
2
3
4
5
6
7
8
9 10 11 12
Month
Figure 36. Monthly energy consumption for the three building “types”
Analyzing the presented simulation results on Figure 36 can be noticed that as expected the
Building III has the smallest heat consumption i.e. regarding specific annual energy consumption,
Building I has 90 kWh/m2a , Building II with 70 kWh/m2a and Building III has 57 kWh/ m2a.
Comparing the energy consumption Building III has 42% lower than Building I and 19% than
Building II.
On Figure 37 are presented the frequency hourly values of the outside dry bulb temperatures for
Skopje, R.Macedonia. All of the meteorological data used in simulations such as temperatures, sun
radiation, relative humidity etc. , are obtained from the base of Meteonorm software which is a
comprehensive meteorological reference. Those values are used as inputs in the simulation to
determine the building heat energy consumption.
Ambient/outside temperature
45
Temperature, °C
35
25
15
5
-5
-15
0
730 1460 2190 2920 3650 4380 5110 5840 6570 7300 8030 8760
Hour of the year
Figure 37. Hourly dry bulb ambient temperatures for Skopje, R.Macedonia
74
Chapter 6. Simulation results and analysis
The station data used in Meteonorm is supplemented by surface data from five geostationary
satellites. This data is available on a global grid with a horizontal resolution of 8 km (3 km in
Europe and Northern Africa). Usually, measurement data can only be used in the vicinity of a
weather station. Elsewhere, the data has to be interpolated between different stations. The
sophisticated interpolation models inside Meteonorm allow a reliable calculation of solar radiation,
temperature and additional parameters at any site in the world. From the monthly values (station
data, interpolated data or imported data), Meteonorm calculates hourly values of all parameters
using a stochastic model. The resulting time series correspond to “typical years” used for
system design [45]. Analysis of the solar thermal system in heating mode for the building
In scenario 1 it is analyzed the influence of the heating system type regarding the heat transfer
elements (underfloor heating or radiators) regarding the efficiency of solar collectors, solar fraction
and the total consumption of heating energy. In the first case analyzed case are set as radiators
heating elements. In modeling of the radiator heating system it is defined that radiative part of
energy transferred equals to 0.99 which is defined only heating.
This fraction of the heater power is supplied as internal radiative gains and distributed to the
walls of the zone. As the set temperature for the heating equipment is related to the air temperature
of the zone, the radiative fraction of the heating power RRAD cannot be higher than 0.99 in order
to have a convective part remaining to ensure stable control of the heating equipment. The radiators
are modeled with Type 1231 from the Tess Library. The heating radiator model is based on the
ASHRAE method outlined in the 2004 ASHRAE Handbook - HVAC Systems and Equipment.
Second case is when the heating system i.e. the heat is transferred through underfloor
heatingThus the modeling is performed so that the building model that defines the active layers
within the floor area. In this case the floor area is divided into seven active layers in a single area of
21.43 m2 in which they are set by the floor heating pipes whose characteristics are defined in the
table.
Table 10. Active layer TYPE data
Data description
Unit
m
Pipe spacing center to center dx
0.2
m
Pipe outside diameter δ
0.02
m
Pipe wall thickness
0.002
kJ/h m K
Pipe wall conductivity
1.26
kJ/kg K
Specific heat of the fluid
4.187
The cross-section of the underfloor heating i.e. the active layers configuration and the pipe
disposition is presented in Figure 38
75
Chapter 6. Simulation results and analysis
Figure 38. Cross section of the active layer for the underfloor heating
The fluid in the storage tank 4 indirectly is heated by solar collectors which further with the
circulating pump mass flow rate of 2000 kg/h is transferred through the auxiliary heater to the
underfloor heating. The auxiliary heating temperature i.e. the output fluid temperature is
maintained on 50 ° C if the heat transfer elements are radiators while for the underfloor heating is
40 ° C. The system is planned heating and DHW tank with a volume of 200 l of internal heat
exchanger and an external electric heater with power of 9 kW. The volume of the storage tank 4 is
considered as a parameter in the analysis with values of 1000 l, 1500 and 2000 l. Each of these
storage tanks are modeled with an internal heat exchanger which technical characteristics as given
in Annex 1. Auxiliary heater power is 12 kW.
Return i.e. exit temperature of the fluid from the heating system is a dynamic variable that
depends on several parameters and is output size of the numerical model of radiators i.e. floor
heating. If the return temperature of the working fluid is higher than the temperature in the storage
tank 4 measured at the highest point 1, it is then redirect directly to the heated auxiliary heater.
The domestic hot water tank 3, differential controller upper band is set to five, lower dead band
to two, while for the storage tank 4 both of the values are three.
In Table 12 and Table 13 are presents results from the heating system simulation. As previously
mentioned main parameter in this analysis is the type of heating i.e. the left half of the table refers
to system with radiators while the right is for a system with underfloor heating. Leading parameters
regarding which comparison is made are: solar fraction for the heating system and for the DHW,
performance and "real" performance of collectors. Solar fraction is calculated as the ratio between
the thermal energy storage tank (calculated with Equation 66) and the sum of the thermal energy
storage tank and auxiliary heater.
⋅
QSH = m sh c p (t o − t i ) , kJ/h
(66)
where:
QSH, kJ/h - extracted heat energy from the storage tank
76
Chapter 6. Simulation results and analysis
⋅
m sh , kg/h
- mass flow rate of working fluid from storage tank to system (heating or
DHW)
to, ti , K
- working fluid outlet and inlet temperatures
The collector thermal efficiency is calculated as ratio between the sum of the useful collector
energy and the total tilted solar radiation on the collector surface.
ηth =
∑ Qcu
∑G
(67)
t
ηth
- thermal efficiency of the collector array
∑Q
- sum of the useful collector array energy transferred to the working fluid
∑G
- sum of the total tilted radiation on collector surface
CU
t
Another parameter in the analysis is so called “real” collector thermal efficiency calculated with
which is again ration between the sum of collector useful energy and the sum of the total tilted
radiation for the collector surface which in this case are only summed the values of radiation when
the circulation pump is on i.e. when the collector is active.
This parameter is inserted in the analysis since it serves as indicator to give sense how long the
collector is in function i.e. the bigger the difference between these two efficiencies means that the
collector more of the time is in stagnation and vice versa.
The mass flow rate through the solar collectors is selected to be 50 kg/h m2 i.e. for the 16 m2 is
800 kg/h, for 32 m2 is 1600 kg/h and for 64 m2 is 3200 kg/h.
The simulations are done with time step of 7.5 min and the result were integrated on monthly
basis. Presented results are averaged values from the monthly values and the period of analysis are
the heating season months for Skopje i.e. October – April. Analyzing the presented results in Table
11 and Table 12 can be noticed that for different collector area and/or storage tank volume
parameters value are the same like the solar fraction, but in the monthly data can be noticed the
discrepancies. Thus must be noted that data average is performed since presenting all of the
monthly value will be too excessive.
77
Chapter 6. Simulation results and analysis
Table 11. Parametric analysis of solar assited heating system with radiators
Radiator heating system
Collector array area, m²
1500 L
1500 L
1500 L
2000 L
BUILDING III
1000 L
2000 L
BUILDING II
1000 L
2000 L
BUILDING I
1000 L
Storage tank
volume
Parameter
Sol.fraction.Heat
Eff_Monthly
Eff.real
Sol.frac.DHW
Sol.fraction.Heat
Eff_Monthly
Eff.real
Sol.frac.DHW
Sol.fraction.Heat
Eff_Monthly
Eff.real
Sol.frac.DHW
Sol.fraction.Heat
Eff_Monthly
Eff.real
Sol.frac.DHW
Sol.fraction.Heat
Eff_Monthly
Eff.real
Sol.frac.DHW
Sol.fraction.Heat
Eff_Monthly
Eff.real
Sol.frac.DHW
Sol.fraction.Heat
Eff_Monthly
Eff.real
Sol.frac.DHW
Sol.fraction.Heat
Eff_Monthly
Eff.real
Sol.frac.DHW
Sol.fraction.Heat
Eff_Monthly
Eff.real
Sol.frac.DHW
16
32
64
0.08
0.25
0.55
0.75
0.07
0.26
0.53
0.74
0.06
0.27
0.53
0.73
0.07
0.24
0.55
0.76
0.06
0.25
0.53
0.75
0.06
0.25
0.54
0.75
0.08
0.23
0.55
0.76
0.07
0.25
0.53
0.76
0.06
0.26
0.53
0.75
0.18
0.18
0.45
0.79
0.17
0.19
0.41
0.79
0.16
0.19
0.39
0.78
0.17
0.17
0.45
0.80
0.15
0.18
0.4
0.8
0.16
0.19
0.40
0.79
0.17
0.17
0.45
0.80
0.16
0.18
0.40
0.80
0.17
0.19
0.40
0.80
0.28
0.13
0.36
0.81
0.28
0.14
0.31
0.81
0.29
0.15
0.30
0.81
0.27
0.12
0.36
0.82
0.27
0.13
0.31
0.82
0.28
0.13
0.30
0.82
0.28
0.11
0.36
0.82
0.28
0.12
0.31
0.82
0.29
0.13
0.30
0.82
In the next Table 12 is given analysis for the same parameters as in Table 11 but with difference
that as heat transfer devices it is used underfloor heating and the fluid supply temperature is 40 °C.
78
Chapter 6. Simulation results and analysis
Table 12. Parametric analysis of solar assited underfloor heating system
Underfloor heating system
Collector array area, m²
1500 L
1500 L
1500 L
2000 L
BUILDING III
1000 L
2000 L
BUILDING II
1000 L
2000 L
BUILDING I
1000 L
Storage tank
volume
Parameter
Sol.fraction.Heat
Eff_Monthly
Eff.real
Sol.frac.DHW
Sol.fraction.Heat
Eff_Monthly
Eff.real
Sol.frac.DHW
Sol.fraction.Heat
Eff_Monthly
Eff.real
Sol.frac.DHW
Sol.fraction.Heat
Eff_Monthly
Eff.real
Sol.frac.DHW
Sol.fraction.Heat
Eff_Monthly
Eff.real
Sol.frac.DHW
Sol.fraction.Heat
Eff_Monthly
Eff.real
Sol.frac.DHW
Sol.fraction.Heat
Eff_Monthly
Eff.real
Sol.frac.DHW
Sol.fraction.Heat
Eff_Monthly
Eff.real
Sol.frac.DHW
Sol.fraction.Heat
Eff_Monthly
Eff.real
Sol.frac.DHW
16
32
64
0.19
0.25
0.55
0.69
0.21
0.28
0.52
0.66
0.23
0.29
0.51
0.64
0.22
0.25
0.54
0.83
0.24
0.27
0.51
0.68
0.27
0.28
0.51
0.66
0.24
0.24
0.54
0.72
0.26
0.26
0.52
0.70
0.29
0.28
0.51
0.68
0.30
0.17
0.45
0.75
0.34
0.19
0.41
0.73
0.32
0.21
0.41
0.71
0.31
0.16
0.44
0.78
0.35
0.18
0.40
0.75
0.40
0.20
0.40
0.73
0.33
0.15
0.45
0.79
0.35
0.17
0.40
0.77
0.42
0.19
0.40
0.75
0.39
0.11
0.38
0.81
0.45
0.13
0.31
0.78
0.44
0.14
0.31
0.77
0.38
0.10
0.37
0.70
0.44
0.12
0.31
0.81
0.49
0.13
0.31
0.80
0.32
0.10
0.37
0.84
0.37
0.11
0.31
0.82
0.52
0.12
0.30
0.81
In order to have better insight over the compared systems from the results in Table 11 and
Table 12 it is derived diagram presentation of the results given on Figure 39
79
Chapter 6. Simulation results and analysis
0.50
0.45
Solar fraction
0.40
0.35
Radiator/1
0.30
Underfloor/1
0.25
Radiator/1.5
0.20
Underfloor/1.5
0.15
0.10
Radiator/2
0.05
Underfloor/2
0.00
16
32
64
Collector area, m²
Figure 39. Solar fraction for Building I, radiator and underfloor heating in regard of
collector array area and strorage volume
The number that is written in the “series” name in the diagram after Radiator and Underfloor
indicates the storage tank volume in m3. It’s easily noticeable on the results from Figure 39 that
there are differences between the solar fractions i.e. bigger differences are occurring for the
combination of lower collector areas and bigger storage volumes. This is because first the analyzed
Building I has lowest thermal insulation i.e. highest heat energy consumption which results in high
frequency heat energy discharge of the storage tank. The underfloor heating has bigger solar
fraction since it uses the thermal mass of the floor which buffers the temperature fluctuations in the
building i.e. tank discharge frequencies thus allowing more time the tank to be reheated with solar
energy and the second important influencing factor is that the underfloor heating has 10°C lower
design driving temperature compared to radiator system.
Another possible analysis is to gain insight for the influence of the specific building energy
consumption on the solar fraction. On Figure 40 are presented results from this analysis from which
firstly it’s obvious that underfloor heating has bigger solar fractions ranging from 24% up to 73%.
This solar fraction differences trend between radiator and underfloor solar assisted heating
increases if the building lowers the specific heat energy consumption and lowers the collector area.
This analysis is done for constant volume of storage tank 4, 1500l and internal heat exchanger
with parameters given in Appendix A.
80
Chapter 6. Simulation results and analysis
0.50
0.45
Solar fraction
0.40
0.35
Radiator/I
0.30
Underfloor/I
0.25
Radiator/II
0.20
Underfloor/II
0.15
0.10
Radiator/III
0.05
Underfloor/III
0.00
16
32
64
Collector array area, m²
Figure 40. Solar fraction for heating with storage tank 1500l regard of collector array area
building type
According to the presented results can be concluded that for buildings with specific heat
consumption from 90 kWh/ m2 a up to 57 kWh/ m2 a, with 0,1 m2/m2conditioned specific collector area
per conditioned surface can be achieved between 20 – 25 % solar fraction, with 0,2 m2/m2conditioned
range round 35% , and with 0,4 m2/m2conditioned maximum 50 %. It should be noted that the solar
fraction also strongly depends from the storage tank volume and for the radiator heating system
increasing the storage volume results in decrease of solar fraction while at the underfloor heating its
vice versa. It is recommended the storage volume to be in the range 50-60 l/m2 collector area in
order to optimize between the solar fraction and collector efficiency.
Also during the simulations monthly average space temperature was tracked in order to see if
the basic comfort conditions are satisfied and the results are presented on Figure 41.
Zone temperature
Temperature, °C
28.00
26.00
24.00
22.00
20.00
18.00
0
1
2
3
4
5
6
7
8
9
10 11 12
Month
Figure 41. Monthly average zone temperature, only heating analysed
81
Chapter 6. Simulation results and analysis
6.3 Analysis of the solar thermal system in cooling mode for the building
This part analyses the performance of solar thermal cooling system i.e. solar driven absorption
chiller. As parameters in the analysis are considered: collector type (flat plate and vacuum tube)
collector area tilt(slope) and azimuth angle and storage tank volume in regard of solar cooling and
DHW fraction, efficiency “ real” efficiency and electrical consumption of the cooling tower fan and
the system fan. As mentioned previously for the absorption chiller it is used the component model
Type 177a in which input parameters are inserted the data from the LiBr/H2O Suninverse 10 chiller
product of Sonneklima with cooling power of 10 kW with heat driving temperature of 75 °C,
cooling water 27 °C and chilled water outlet temperature set at 15°C .
In general are considered two scenarios, first is with flat plate collectors and the second is with
vacuum tube collectors.
6.3.1 Scenario 1
The system is simulated as shown on Figure 26 i.e. the solar collectors are used to heat the fluid
in storage tank 4 which is connected to the generator of the absorption chiller. In serial connection
after the storage tank outlet is installed auxiliary heater used to add heat in periods when the
working fluid has lower temperature that the set one. This scenario considers flat plate collectors
with characteristic given in Table 1 and the Building II type, which thermal performance described
in Table 9. The building internal heat gains consider the lighting power density 5 W/m2 and the
home appliances with specific power of 2 W/m2. The absorption chiller condenser is connected to
the wet cooling tower product of Baltimore AirCoil type PF2-0406AA-31-3 (technical details given
in Appendix E). It’s used Type510 model from Tess library, a closed circuit cooling tower which
cools the liquid stream by evaporating water from the outside of coils containing the working fluid.
The working fluid is completely isolated from the air and water in this type of system. Data used in
the numerical model are given in Table 13.
The control signal of the cooling tower fan is set to have the tower try to maintain the desired
outlet water temperature of 26 °C and the fluid flow rate with the circulation pump is set to 2600
kg/h. Also to the circulation of the cooling water are modeled pipe network with diameter 0.04 m
and length of 15 m heat transfer coefficient for thermal losses 3 kJ/h m2 K which accounts for the
heat losses to the environment and also increases the system thermal capacity affecting the
simulation stability. Values of the inlet parameters for the cooling tower such as ambient air
temperature and relative humidity are read from the weather component respectively for the
simulated time and period of the year.
82
Chapter 6. Simulation results and analysis
The cooling water temperature is parameter which is variable depending from the absorption
chiller working conditions.
Table 13. Cooling tower design parameters
Parameter
Design inlet temperature
Design outlet fluid temperature
Design fluid flow rate
Fluid specific heat
Design ambient air temperature
Design wet bulb temperature
Design air flow rate
Rated fan power
Unit
°C
°C
kg/s
kJ/kg K
°C
°C
kg/s
kW
Value
35
29.4
1.7
4.19
35
25.5
8
0.56
The cooling system in the building is modeled using the ventilation air distribution system.
Combination among the chilled water from the absorption chiller and the ventilation air is provided
with heat exchanger water-air modeled Type 508a which is a cooling coil modeled using a bypass
approach in which the user specifies a fraction of the air stream that bypasses the coil. The
remainder of the air stream is assumed to exit the coil at the average temperature of the fluid in the
coil and at saturated conditions. The two air streams are remixed after the coil. Chilled water flow
from the absorption chiller to the cooling coil is set to 2900 kg/h and the air flow rate to the
building is 4000 kg/h.
The auxiliary heater power is modeled 12 kW and the outlet temperature is 80 °C, which is the
absorption machine driving temperature.
The storage tank 4 in this case is used to store the heat for driving the absorption chiller. With
the thermostat Type 108 is regulated the space temperature in the building set to 26 °C, which
control signal is directly regulating the function of the circulation pump from the chilled water
absorption chiller and the fan distributing the conditioned air.
Simulation time step is 15 min and as cooling period are considered the months from MaySeptember.
The collector efficiencies are calculated same as described in Chapter 5. Solar fraction is
calculated as ratio between the useful collector energy transferred to the working fluid and the total
radiation on the collector surface. Also there is the DHW storage tank which solar heating control
is same as described before in the heating system analysis i.e. the advantage of the solar energy is
given to heat the storage tank 4. In this analysis are considered the electrical consumptions of the
circulation pumps and fans.
83
Chapter 6. Simulation results and analysis
Table 14. Monthly average solar fractions and efficiency in regard of collector areaa and
storage volume
Flat plate, Tilt 40°, Azimuth 0° - flow rate 50 l/h m²
1000
Collector area m²
1500
2000
Sol.fraction
16
0.28
32
0.50
64
0.67
16
0.24
32
0.49
64
0.69
16
0.19
32
0.48
64
0.70
Avg_Efficiency
Avg_Eff_Real
Sol.DHW
0.26
0.48
0.98
0.19
0.36
0.99
0.15
0.32
0.99
0.26
0.47
0.98
0.20
0.37
0.99
0.15
0.33
0.99
0.27
0.46
0.98
0.21
0.37
0.99
0.16
0.30
0.99
The comparison results between the solar fractions and efficiencies in regard of different
collector areas and storage volumes are presented in Table 14. The mass flow rate is set constant
according the collector area i.e. it is 50 kg/h m2. Solar collectors are tilted on 40° with south
orientation i.e. azimuth is 0°. As can be seen the solar fractions increase with the increase of
collector area and storage volume and varies in the range between 20% up to 70%. Also the thermal
efficiency should not be neglected which for the analyzed cases is in the range between 15% up to
27% monthly averages. Solar fraction for the DHW is almost in every case 100% since the daily
consumption is very low compared to the available energy from the solar collectors.
According the above presented results can be concluded that with solar energy regarding the
specific collector areas can be covered: 0,1 m2 / m2conditioned can cover almost 30% , 0,2 m2 /
m2conditioned covers 50% and 0,4 m2 / m2conditioned can cover 70% of the total required heating energy
for driving the absorption chiller. Analyzing the solar fraction for one constant specific collector
area and changing only the storage volume can be noticed that biggest fractions are for specific
volume per collector area of 30 l/m2.
Another analysis is done to estimate the influence of collector interconnections (number in
parallel and/or serial) regarding the solar fraction and thermal efficiency. The main idea for this
analysis is to check if the higher fluid outlet temperature induced by serial collector connections
will increase the solar fraction.
Table 15. Monthly avrage solar fractions and thermal efficiency in regard of colelctor array
interconnections
1000l, tilt 30° azimuth 0°
16/1 - 800 kg/h 16/4 - 400 kg/h 32/1-1600 kgh
32/2-800 kgh 64/1 - 3200 kg/h 64/2 - 1600 kg/h
Sol.fraction
0.30
0.27
0.52
0.55
0.69
0.69
Avg_Efficiency
0.26
0.27
0.20
0.18
0.15
0.15
Avg_Eff_Real
0.50
0.50
0.37
0.38
0.33
0.32
1500l, tilt 30° azimuth 0°
16/1 - 800 kg/h 16/2 - 400 kg/h 32/1-1600 kgh 32/2-800 kgh
64/1 - 3200 kg/h 64/2 - 1600 kg/h
Sol.fraction
0.24
0.24
0.51
0.51
0.69
0.69
Avg_Efficiency
0.26
0.26
0.20
0.20
0.15
0.15
Avg_Eff_Real
0.47
0.47
0.38
0.37
0.33
0.32
2
2
* 16/2-400 kg/h - total collector array has area of 16 m , each collector module area is 2 m and two modules each with
four collectors are in serial connection as presented on Figure 42.
84
Chapter 6. Simulation results and analysis
16/2 - 400 kg/h
2
2m
2
2m
2
2
2
2m
2m
2
2m
2
2m
2
2m
2m
Figure 42. Scheme of the collector array conection 16/2
It is considered that all of the collectors have tilt of 30 ° directed toward south i.e. with azimuth
of 0 °. The specific mass flow rate is kept constant with value 50 l/h m2.
In Table 15 are presented the results from this simulation of the absorption cooling system
where parametric analysis is done varying the number of collector in parallel and serial connection.
Regarding the storage tank 4 the analysis is done for 1000 l and 1500 l.
In general analyzing the results can be concluded that serial connection between the collector
modules doesn’t affect too much on the solar fraction and thermal efficiency i.e. only slight
increase at the collector array with 32 m2 and storage tank of 1 500 l. Thus recommendation
collectors to be in parallel connected since there are no or slight improvements in solar fraction but
on the other hand there is decrease in thermal efficiency, increase in “real” efficiency indicating
that collector longer stays not in function i.e. it is in stagnation and the last disadvantage is the
increase in pressure loss and thus there is increase in pump energy consumption.
Further analyzed the influence of the collector tilt angle with and without azimuth tracking
system (one axis-vertical tracking system) to the solar fraction and thermal efficiency. Simulation
results are presented in Table 16 .
Table 16. Solar fractions and thermal efficiency for solar assited cooling system in regard of
collector orientation i.e. azimuth
Storage tank 1000l - Collector mass flow rate 50 l/h m²
Collector area
m² / tilt °
Sol.fraction
Avg_Efficiency
Avg_Eff_Real
16/1/40
16/1/30
0.28
0.26
0.48
0.30
0.26
0.50
16/1/30/T
32/1/40
32/1/30
32/1/30/T
0.44
0.53
0.55
0.67
0.24
0.18
0.18
0.19
0.50
0.38
0.38
0.41
Storage tank 1500l - Collector mass flow rate 50 l/h m²
64/1/3200/40 64/1/3200/30 64/1/3200/30/T
0.67
0.15
0.32
0.69
0.15
0.33
0.77
0.16
0.35
Collector area
16/1/40
16/1/30
16/1/30/T
32/1/40
32/1/30
32/1/30/T
64/1/3200/40 64/1/3200/30 64/1/3200/30/T
m² / tilt °
Sol.fraction
0.22
0.24
0.38
0.49
0.51
0.66
0.67
0.69
0.81
Avg_Efficiency
0.26
0.26
0.26
0.20
0.20
0.20
0.15
0.15
0.15
Avg_Eff_Real
0.46
0.47
0.47
0.37
0.38
0.39
0.32
0.33
0.36
*64/1/3200/30/T - (64) m² collector array , (1) all in parallel connected, (3 200 ) kg/h mass flow rate, (30) tilt angle, (T) tracking azimuth
85
Chapter 6. Simulation results and analysis
In order to have better visibility of the parametric influence of the collector tilt angle and
azimuth influence, on Figure 43 are presented monthly values for solar fractions from JuneSeptember
0.50
0.45
Solar fraction
0.40
0.35
16/1/30
0.30
16/1/40
0.25
16/1/30/T
0.20
0.15
0.10
June
July
August
September
Figure 43. Monthly values of solar fraction for different tilt angles and tracking azimuth for
solar assited cooling system
The presented results indicates that the solar fractions are bigger in average for 2% for
collectors tilted on 30° compared to collectors tilted 40°. Much bigger differences can be noticed
for collectors with azimuth tracking system, ranging between 10% - 14%.
Another case is analyzed i.e. compared the system presented on Figure 26 with and without
cold storage marked with number 8. The analysis is in regard of the solar fraction and cooling
tower fan electrical energy consumption since it’s the largest energy consumer in the absorption
chiller assembly. Analyzed system has 32 m2 flat plate solar collector with tilt of 30° south
oriented-azimuth 0°, 1500 l hot storage tank and 500 l cold storage tank. Type 534 is used to model
the cold storage tank which is divided into isothermal temperature nodes. The degree of
stratification is defined through the specification of the number of "nodes" which in this case is five
nodes. The building is type II according the description in Table 9 and the control strategy and
internal heat gains are in accordance with the previously described.
86
0.65
80
0.60
70
0.55
60
50
0.50
40
0.45
30
0.40
20
0.35
10
0.30
Fan cooling tower
energy, kWh
Solar fraction
Chapter 6. Simulation results and analysis
0
June
July
August
September
With tank
Without tank
Fan/With tank
Fan/Without tank
Figure 44. Solar fraction and fan cooling tower energy consumption for system with and
without storage tank for solar assited cooling system
According the presented results on Figure 44, the conclusion is that installing cold storage tank
between the absorption chiller and building cooling system will cause decrease in the solar fraction
but also will decrease the fan cooling tower electrical consumption. The decrease in solar fraction
can be explained with the fact that the stored chilled water causes decrease in charging/discharging
the hot storage tank thus collector more of the time is in stagnation. Graphical presentation of
81.00
1400
80.00
1200
79.00
1000
78.00
800
77.00
600
76.00
400
75.00
200
74.00
Collector energy yield, kWh
Temperature, °C
previously explanation is given with the diagram on Figure 45.
0
June
July
August
September
With cold tank
Without cold tank
Collector yield/with
Colector yield/without
.
Figure 45. Monthly average storage tank temperture and solar collector yield in regard of
system with/without cold storage tank for solar assited cooling system
6.3.2 Scenario 2
This scenario considers the same system as described in Scenario 1 only the collectors are
vacuum type. The analysis is directed toward comparison between the flat plate and vacuum tube
collectors regarding the solar fraction and thermal efficiency in solar cooling system.
87
Chapter 6. Simulation results and analysis
Vacuum tube collector are product of Camel Solar CS15, with technical data given in Table 2.
The reference building is the type II as described in Table 9 and two cases with hot water storage
tanks volumes of 1000 l and 1500 l.
Table 17. Average solar fraction and thermal efficiency of solar cooling system in regard of
collector area and strage tank volume for solar assited cooling system
Vacuum tube collector, tank 1 500l, specific flow rate 50 kg/h m², tilt 30°, azimuth 0°
Solar collector area, m²
16
18
10
14
0.66
0.70
0.42
0.60
Sol.fraction
0.41
0.39
0.52
0.44
Avg_Efficiency
Vacuum tube collector, tank 1 000l, specific flow rate 50 kg/h m², tilt 30°, azimuth 0°
Solar collector area, m²
10
14
16
18
Sol.fraction
0.46
0.59
0.66
0.71
Avg_Efficiency
0.48
0.46
0.43
0.41
22
0.77
0.35
22
0.78
0.37
Simulations were performed for storage tanks with volumes 1 000 l and 1500 l taking the collector
area as parameter with areas of 10, 14, 16,18 and 22 m2 where aperture area one collector module is
1,42 m2 which has 15 vacuum tubes. Solar fractions as presented in Table 17 range between 42%
for collector array of 10 m2 up to 78 % with collector array of 22 m2. As expected, for same
collector areas thermal efficiency decreases with the decrease of the storage tank volume. Although
with 22 m2 can be achieved solar fractions of 78% limiting factor is the economic viability of the
system which should have reasonable payback meaning it should correspond i.e. lower than the
Solar fraction/Avg_Efficiency
lifetime period of the system .
1.00
0.80
0.60
0.40
0.20
0.00
10
14
16
18
22
Vacuum collector area, m²
SF
EE
Figure 46. Solar fraction and thermal efficiency in regard of collector area for solar assited cooling
system
Table 18. Solar fraction and thermal efficiency for different specific vacuum collector areas
storage volumes for solar assited cooling system
Vacuum tube collector, tilt 30 °, azimuth 0°
Specific coll.area
Solar fraction
Avg_efficiency
0.1 m² /m²condit. 70 l/m²
0.1 m² /m²condit. 105 l/m²
0.62
0.41
0.60
0.44
88
0.1 m² /m²condit. 142 l/m²
0.59
0.46
Chapter 6. Simulation results and analysis
In Table 18 are results for the analyzed cases with specific collector area of 0.1 m2 per m2
conditioned area varying the specific storage tank volume per collector area used to determine the
solar fractions and thermal efficiency.
According the presented results in Table 18 and Table 19, optimizing between values of the
solar fraction and thermal efficiency, results recommendation of 60 l storage volume per square
meter vacuum tube collector area. Vacuum tube collectors have bigger specific storage volume
compared to the flat plate collectors (30 l/m2) since they have higher efficiency at higher working
fluid temperatures.
6.4 Techno-economic analysis
An important part of the development of any project is the evaluation of its profitability. At the
first stage of evaluation the consequences of project financing are normally not considered. Thus,
the consequences of loan interest, taxes, grants, subsidies etc. are not taken into account in the
project profitability calculations.
The economic viability of a solar energy system depends on two factors: (1) the value of
conventional energy displaced (fuel cost savings) and (2) the capital cost of hardware necessary to
achieve those savings. There are two principal approaches to economic analysis that are related
payback analysis and cost of delivery energy.
There are many factors that affect the decision to adopt a new technology including the real or
perceived risks, the economic attractiveness as expressed by some figure of merit and the status as
an innovator or trendsetter. The economic evaluation of a solar application includes factors such as
the capacity cost of delivering solar energy, the optimum sizing of collectors and other equipment
the costs of competing technologies and financial analyses. There figures of merit that are used to
accept or reject particular solar application including simple payback, cash flow , capital cost per
unit of energy saved, life cycle cost, net present value and levelized energy cost.
To establish cost goals for a future technology the analysis of the economics of active solar
systems can be turned around to ask the question: Based on the value on energy savings, how much
can be spet on the system? This value becomes the cost goal for the system. For a range of
economic assumptions, an analysis considering a 5-year or 7 year simple payback is assumed to be
an adequate figure of merit to establish cost goals for active solar cooling and heating technologies.
For residential applications, a payback period of 5 - 7 years may consider as acceptable.
Relatively short payback periods required for market acceptance of a new technology, fuelescalation rates do not greatly change the required cost-goal multiplier. Certainly, for estimating the
approximate cost goals for future technologies, the uncertainties generated by assuming a simple
89
Chapter 6. Simulation results and analysis
payback period are smaller than the uncertainties of future costs of collectors and other
components. The actual cost of energy at the time when future systems are proposed and built has a
significant impact on the future economic attractiveness of solar options.
6.4.1 Evaluating cost effectiveness
To evaluate the cost effectiveness of system that have not yet been built requires: (1) estimating
the load to be served and calculating the contribution of the active solar system to meet the load, (2)
calculating the contribution of an efficient conventional system to meet the same load and there by
establish the energy and fuel cost savings attributable to the solar system, (3) estimating the cost of
active solar system with full back up and the cost of conventional system, (4) comparing the energy
cost savings to the incremental solar system cost with suitable economic assumptions.
In the design of a solar system, the performance of a specific application is calculated by use of
simplified design methods or it is modeled in detail.
The performance of future conventional space-conditioning systems affects the economic
potential of active solar systems. The performance and cost of today’s conventional heating,
cooling and domestic hot water system can be readily determined, but conventional heating and
cooling technology is constantly improving.
As mentioned previously in the text, first should be determined the building heating/cooling
loads. In the analysis it is considered the reference Building (type II) with specific energy
consumption of 70 kWh/m2 a for which with the simulation are determined the heating and cooling
loads. The time step used in the simulations is 7,5 min and the heating and loads are integrated on
hourly basis. On Figure 47 are presented results from the simulation of solar assisted heating with
flat plate collectors varying their total area 16 m2, 32 m2, 64 m2 , mass flow rates are 50 kg/h m2
and constant heat storage tank of 1000 l. Collectors are tilted 40° toward south – azimuth 0° also is
installed 200l DHW storage tank heating with the same collector array only in period when the
heating storage tank is charged or the condition for the circulation pump is not satisfied.
90
Chapter 6. Simulation results and analysis
0
360
720
1080
1440
1800
2160
2520
2880
3240
3600
3960
4320
4680
5040
5400
5760
6120
6480
6840
7200
7560
7920
8280
8640
Heat energy, kWh
160
140
120
100
80
60
40
20
0
Hour of the year, h
Buidling heating loads
Collector yield 16/1
Collector yield 32/1
Collector yield 64/1
Collector yield 64/2
Figure 47. Hourly heating loads and useful heat yield for different collector array areas and
constat storage tank of 1000 l and one case for 64 m2 with 2000 l
It can be seen that there is no big difference between the solar collector yield of 32 m2 and 64
m2 since the storage tank capacity is too small to accept the 64 m2 collector array solar/heat yield,
which results in decrease in solar fraction and stagnation of the collectors.
Since the diagram on Figure 46 is very dense because of the hourly values and cannot be easily
noticed the differences, therefore on Figure 47 are also presented collector yields but only for ten
days.
140
Heat energy, kWh
120
100
80
60
40
20
0
576
600
624
648
672
696
720
Collector yield 16/1
Hour of the year, h
Collector yield 32/1
Heat building loads
Collector yield 64/2
744
768
792
Collector yield 64/1
Figure 48. Hourly heating loads and collector energy yield for different areas ten day
period
With the obtained building energy heat consumption next step is the LCCA (Life cycle cot
analysis) in regard of: auxiliary heat source, collector areas and storage tank. Common for the
analyzed systems is that in each of them the heat is distributed through the underfloor heating with
91
Chapter 6. Simulation results and analysis
flow rate of 2000 kg/h, solar collector array mass flow rate depending from the collector area i.e. 50
kg/h m2 , auxiliary heat energy is provided by heaters located at the fluid tank outlet with capacity
of 12 kW for the heating system and 9 kW for the DHW.
Comparison is performed between different combinations of solar thermal systems and
auxiliary heating devices in regard of conventional heating system with electrical boiler. The
analyzed solar systems have total solar collector area of 16 m2, 32 m2 and 64 m2 combined with
storage tanks of 1000 l, 1500 l and 2000l , and auxiliary heating energy provided by electric heater
or heat pump air-water with E.V.I compressors with nominal capacity of 15 kW product of Hidros
model Lzti 10.
In Table 19 are presented data for delivered heating energy, annual heating energy cost, system
price and the environmental indicator - CO2 emissions for conventional heating systems with heat
sources: electrical energy, wood pellets and heat pump. For the heat pump COP is assumed
averaged yearly value of 2,5.
Table 19. Annual energy balance, energy and system costs, CO2 emissions for conventional
heat source systems obtained with simulations, specific heat energy consumption 70 kWh/m2 a
Parameter
Heat power
Annual delivered heat energy
Average thermal
efficiency/COP
Annual consumed energy
System electrical energy
consumption (circ.pumps)
Specific energy price
Total energy price
Heat source device/system
price
Annual CO2 emissions
Unit
Electrical
boiler
kW
kWh
kWh
Pellet boiler Heat pump
12
13103
0.99
13235
0.91
14399
2.5
5241
144
kWh
eur/kWh
eur
0.09
1203
0.05
667
0.09
472
eur
kg/year
800
12177
2000
58
5000
1730
The CO2 emissions and primary energy consumption factors are provided from the standard
“Energy performance of buildings - Overall energy use and definition of energy ratings” EN
15603_2008 [46].
Regarding the energy prices, the electrical energy price (euro/kWh) is provided from the
Energy Regulatory Commission of the Republic of Macedonia valid from 01.07.2014 [47] , wood
pellet energy price is obtained by taking the average market price of 0,2 euro/kg with lower heating
value Hd = 4,3 kWh/kg.In Table 20 and Table 21 are given the results for the analyzed solar thermal
systems in regard of heat energy consumption, annual energy and system costs and the
92
Chapter 6. Simulation results and analysis
environmental impact indicator presented by the value for the annual CO2 (kg/year) annual
emissions.
Table 20. Solar thermal system energy performance, costs, parameters and CO2 emissions,
primary energy consumption for building - heat energy consumption 70 kWh/m2 a, – part I
Ref.building 70 kWh/m² a
Parameter
Auxiliary energy - heating system
Auxiliary energy - DHW
System electrical energy
consumption (circ.pumps)
Specific electrical energy price
Total energy price
Heat source device/system price
Primary energy consumption
Annual CO2 emissions
Annual average solar fraction - SF
Unit
kWh
kWh
Solar thermal system - Area/Storage tank volume - EH (electric heater) ; HP (heat pump) 16/1000-EH 16/1000-HP 32/1000-EH 32/1000-HP 64/1000-EH 64/1000-HP 32/1500-EH
8550
8550
7420
7420
6466
6466
6996
750
562
428
639
kWh
eur/kWh
eur
eur
kWh
kg/year
144
144
114
114
90
90
130
850
3600
28777
3069
348
7000
12790
1275
729
6000
26798
2672
0.09
300
10500
10946
1091
629
10000
23117
2305
256
14700
9426
940
640
7000
25702
2562
-
0.19
0.30
0.39
0.34
Table 21. Solar thermal system energy performance, costs, parameters and CO2 emissions, primary
energy consumption for building with specific heat energy consumption 70 kWh/m2 a, – part II
Parameter
Auxiliary energy - heating system
Auxiliary energy - DHW
System electrical energy
consumption (circ.pumps)
Specific electrical energy price
Total energy price
Heat source device/system price
Primary energy consumption
Annual CO2 emissions
Annual average solar fraction - SF
Unit
kWh
32/1500-HP 64/1500-EH 64/1500-HP 32/2000-EH 32/2000-HP
6996
5830
5830
7250
7250
639
490
64/2000-EH 64/2000-HP
5921
5921
kWh
eur/kWh
eur
eur
kWh
kg/year
130
103
103
138
138
111
111
287
11500
10539
1051
578
12000
21260
2120
237
16500
8709
868
0.09
665
8000
24454
2438
273
12500
10056
1003
543
12800
19966
1991
223
16700
8207
818
-
0.34
0.45
0.32
0.44
Further is analyzed how the building energy consumption affects the solar system performance i.e.
same systems are applied to building with lower specific heat energy consumption. Thus it is
analyzed solar thermal system installed on Building III (as described in Table 9) with specific heat
energy consumption on annual basis of 57 kWh/m2 a
In order to obtain better insight for the solar thermal system viability, the method of life cycle cost
analysis (LCCA) is used to find the cost-effective optimal combination between the solar collector
area, storage tank volume and auxiliary heater type. It is obvious that better and feasible
combination with heat pump instead of electrical heater although the starting investment is much
bigger.
.
93
Chapter 6. Simulation results and analysis
Table 22. Solar thermal system energy performance, costs, parameters and CO2 emissions,
primary energy consumption for building with specific heat energy consumption 57 kWh/m2 a,
– part I
Ref.building 57 kWh/m² a
Parameter
Auxiliary energy - heating system
Auxiliary energy - DHW
System electrical energy
consumption (circ.pumps)
Specific electrical energy price
Total energy price
Heat source device/system price
Primary energy consumption
Annual CO2 emissions
Annual average solar fraction - SF
Solar thermal system - Area/Storage tank volume - EH (electric heater) ; HP (heat pump) Unit
16/1000-EH 16/1000-HP 32/1000-EH 32/1000-HP 64/1000-EH 64/1000-HP 32/1500-EH
kWh
6708
6708
5762
5762
5418
5418
6192
kWh
713
527
402
595
kWh
139
eur/kWh
eur
eur
kWh
kg/year
139
680
3600
22664
2449
280
7000
10285
1025
109
109
86
86
115
576
6000
21177
2111
0.09
239
10500
8687
866
532
10000
19549
1949
217
14700
7990
797
567
7000
22846
2278
0.28
0.33
0.22
0.39
Table 23. Solar thermal system energy performance, costs, parameters and CO2 emissions,
primary energy consumption for building with specific heat energy consumption 57 kWh/m2 a,
– part II
Ref.building 57 kWh/m² a
Parameter
Auxiliary energy - heating system
Auxiliary energy - DHW
System electrical energy
consumption (circ.pumps)
Specific electrical energy price
Total energy price
Heat source device/system price
Primary energy consumption
Annual CO2 emissions
Annual average solar fraction - SF
Unit
kWh
kWh
Solar thermal system - Area/Storage tank volume - EH (electric heater) ; HP (heat pump) 32/1500-HP 64/1500-EH 64/1500-HP 32/2000-EH 32/2000-HP 64/2000-EH 64/2000-HP
6192
5600
5600
5074
5074
4042
4042
595
457
630
488
kWh
eur/kWh
eur
eur
kWh
kg/year
115
94
94
118
118
97
97
255
11500
9367
934
554
12000
20360
2030
227
16500
8331
831
0.09
524
8000
19271
1921
216
12500
7943
792
416
12800
15315
1527
172
16700
6319
630
0.28
0.36
0.41
0.53
In the following analysis will be considered collector areas of 16 m², 32 m² and 64 m², storage tank
volumes of 1000 l, 1500l and 2000l, and for auxiliary heater are considered electrical heater and
air-water heat pump.
6.4.2 Life cycle cost analysis
Life Cycle Cost (LCC) analysis is an economic method of project evaluation in which all costs
occurring from building, owning, operating, maintaining and ultimately demolishing/disposing of
the project are taken into consideration. In a LCC analysis the costs occurring at different time is
discounted to their present value.
The investment includes all expenses connected to the realization of the project, normally the
following elements: Design/Planning, project Management / quality assurance, components
installation, control and testing, as-built document, commissioning, training, other expenses, taxes,
VAT.
94
Chapter 6. Simulation results and analysis
Technical life is the physical lifetime of the investment/equipment, i.e. for how long the
equipment technically can operate. Economic life is the practical lifetime for the
investment/equipment, i.e. the lifetime before it is profitable to change into new equipment.
If components/products are being replaced before they are worn out as a result of new and more
efficient components available on the market, then the economic lifetime is shorter than the
technical lifetime. Changes in standards and regulations, energy prices, comfort levels, etc. may
also lead to replacing equipment prior to the end of their technical life.
Inflation “b”, is defined as the yearly average price increase for all consumer goods. In this
analysis the inflation is considered to be 2 %.
A discount rate nr , is used to calculate the present value of, for instance, future energy savings,
adjusted for the cost of capital. Discount rates could be in nominal or real terms, where the real
discount rate is adjusted to eliminate the effects of expected inflation. The nominal discount rate
includes the expected general inflation. The real discount rate is the nominal rate corrected for
inflation, relative increase of energy price, and other possible relative price increases. In this case
the nominal discount rate is 6 % and the real discount rate calculated with Equation 68 is:
r=
nr − b
1+ b
(68)
6.4.3 Calculation of Profitability
There are a number of methods to calculate the profitability of investments:
• Payback
• Net Present Value
• Net Present Value Quotient
• Pay-Off
• Internal Rate of Return
The discounted value (present value) concept is the basis for several of these methods. The
following parameters are used for the calculations:
• Investment I0 [€]
• Annual net savings B [€/year]
• Economic lifetime n [year]
• Real discount rate r · 100 [%]
The simple payback is the time it takes to pay back the investment, based on equal annual net
savings. If the payback is longer than economic lifetime of the measure, the measure is not
95
Chapter 6. Simulation results and analysis
profitable. The Payback method is useful for quick calculations, but there are limitations: It should
be used when the real discount rate is low, if payback period is less than 4 - 5 years and if the
method ignores the value of annual savings after the payback period.
In order to summarize the discounted value of the future annual savings, it is necessary to
define a reference year, to which all investments and savings should be related. All in- and
outgoing payments should be related to the same reference year. Normally, the selected year is the
one in which the investments are made (year 0). The Net Present Value (NPV) of a measure or project
is today's value of all the future annual net savings during the economic life time minus the initial
investment. The NPV of a measure or project is today's value of all the future annual net savings during the
economic life time minus the initial investment. The NPV value must be positive in order the measure to be
profitable.
The Pay-off is the time it takes to repay the investment, considering the real discount rate. This means
the number of years that makes the NPV equal to 0.
The IRR is the interest rate that equates the net present value of future savings/cash flow over
the economic lifetime of the assets to the cost of the investment.
The Net Present Value Quotient, NPVQ, is the ratio between the Net Present Value and the total
investment. The highest NPVQ indicates the most profitable measure. The NPVQ could be used for internal
ranking of energy efficiency measures
The LCC method provides a better assessment of the long-term cost-effectiveness of the
project/measure than other economic methods which only focus on initial investment costs and
operation costs for the first years. At the same time, a LCC analysis requires more information than
other profitability evaluation methods. The lowest Life Cycle Costs indicates the most profitable
investment, measure or solution.
96
Chapter 6. Simulation results and analysis
Table 24. LCCA for solar thermal system assited buildign heating with specific .heat energy
consumption 70 kWh/m2 a
Real interest rate 3,9% ; Building 70 kWh/m2 a
Measures
Investment Net savings
EUR
EUR/Year
Pellet boiler
2000
506
Heat pump
5000
671
HP 16/1-1000
7000
814
HP 32/1-1000
9700
849
EH 16/1-1000
3600
312
HP 32/1-1500
10700
864
EH 32/1-1500
7000
511
EH 32/1-1000
6000
421
EH 32/1-2000
8000
487
HP 64/1-1000
14700
881
HP 64/1-1500
16500
902
HP 32/1-2000
16500
868
HP 64/1-2000
17300
896
EH 64/1-1000
10000
509
EH 64/1-1500
12000
560
EH 64/1-2000
12800
596
Lifetime
Year
20
20
20
20
20
20
20
20
20
20
20
20
20
20
20
20
PB
Year
4
7.5
8.6
11.4
11.5
12.4
13.7
14.3
16.4
16.7
18.3
19
19.3
19.7
21.4
21.5
PO
Year
4.4
9
10.7
15.5
15.7
17.3
20
21.3
26.9
27.6
32.9
35.6
36.8
38.3
47.7
48
IRR
%
25
12
10
6
6
5
4
3
2
2
1
0
0
0
0
0
NPV
EUR
4.93
4.18
4.14
1.919
670
1.125
-7
-239
-1.335
-2.643
-4.156
-4.621
-5.038
-3.034
-4.336
-4.644
NPVQ˟Max.Investment
EUR
Year
2.46
4.12
12
0.84
6.34
12
0.59
7.687
12
0.20
8.017
12
0.19
2.946
12
0.11
8.159
12
0.00
4.825
12
-0.04
3.975
12
-0.17
4.599
12
-0.18
8.319
12
-0.25
8.518
12
-0.28
8.196
12
-0.29
8.461
12
-0.30
4.806
12
-0.36
5.288
12
-0.36
5.628
12
PB=Payback, PO=Pay-off, IRR=Internal Rate of Return, NPV=Net Present Value, NPVQ=Net Present Value Quotient,
* Maximum investment [EUR] with [Years] pay-off
According the presented results several conclusions can be drawn. As most profitable measure
according the LCC analysis is the measure with the highest NPVQ. Thus, most profitable would be
installing pellet boiler as unique heating source with payback period of 4 years. Next profitable
measure would be solar system with total collector area of 16 m2 and heat pump as additional
auxiliary heat source. Further follows the systems with collector array areas of 32 m2 and 64 m2
with different storage tank volumes in combination with auxiliary heat device electric heater (EH)
or heat pump (HP). LCC analysis reveal that for same collector area and storage tank volume more
profitable is the combination solar collectors with heat pump instead with electric heater although
the starting investment is far lower for electric heater but energy savings contribute to have lower
payback period for the heat pump system.
It can be noticed from Table 24 that the net present values for all of the system combination
with 64 m2 collectors, have negative values which indicates that those systems are not profitable for
buildings with specific heat energy consumption of 70 kWh / m2 a.
Next is LCC analysis for the same solar thermal systems but applied to building with lower
specific heat consumption of 57 kWh/m2 a, represented with the Building type III described in
Table 9.
97
Chapter 6. Simulation results and analysis
Table 25. LCCA for solar thermal system assited buildign heating with specific .heat energy
consumption 57 kWh/m2 a
Real interest rate 3,9% ; Building 70 kWh/m2 a
Measures
Investment Net savings
EUR
EUR/Year
Pellet boiler
2000
314
Heat pump
5000
571
HP 16/1-1000
7000
702
EH 16/1-1000
3600
301
HP 32/1-1000
9700
730
HP 32/1-1500
10700
715
EH 32/1-1000
6000
393
HP 32/1-2000
12500
754
EH 32/1-1500
7000
393
EH 32/1-2000
8000
446
HP 64/1-1000
14700
768
HP 64/1-2000
17300
786
EH 64/1-1000
10000
454
HP 64/1-1500
16500
731
EH 64/1-2000
12800
542
EH 64/1-1500
12000
404
Lifetime
Year
20
20
20
20
20
20
20
20
20
20
20
20
20
20
20
20
PB
Year
6.4
8.8
10
12
13.3
15
15.3
16.6
17.8
17.9
19.1
22
22
22.6
23.6
29.7
PO
Year
7.5
10.9
12.9
16.5
19.1
23
23.7
27.3
31.2
31.6
36.1
51.7
51.8
56.3
67.7
99
IRR
%
15
10
8
6
4
3
3
2
1
1
0
0
0
0
0
0
NPV
EUR
2.297
2.814
2.607
519
290
-915
-622
-2.181
-1.622
-1.896
-4.19
-6.453
-3.787
-6.496
-5.383
-6.471
NPVQ˟Max.Investment
EUR
Year
1.15
2.965
12
0.56
5.392
12
0.37
6.629
12
0.14
2.842
12
0.03
6.893
12
-0.09
6.752
12
-0.10
3.711
12
-0.17
7.12
12
-0.23
3.711
12
-0.24
4.212
12
-0.29
7.252
12
-0.38
7.422
12
-0.38
4.287
12
-0.39
6.903
12
-0.42
5.118
12
-0.54
3.815
12
PB=Payback, PO=Pay-off, IRR=Internal Rate of Return, NPV=Net Present Value, NPVQ=Net Present Value Quotient,
* Maximum investment [EUR] with [Years] pay-off
In order to obtain better visibility of the results presented in Table 24 and Table 25 it is
constructed diagram Figure 49 between the heating system/source type and the simple payback
period for the two building types and only the system with positive NPV are taken into account.
14
Payback period, [year]
12
10
8
6
4
2
0
Pellet boiler
Heat pump HP 16/1-1000 EH 16/1-1000 HP 32/1-1000
Budilding 57
Building 70
Figure 49. Payback time for solar thermal systems for two building types regarding the
specific heat energy consumption
It’s interesting to be noticed that same solar thermal systems for the building with the higher
specific energy consumption have lower payback period since they have bigger energy
consumption thus genrating bigger energy differences between the basic scenario with electrical
boiler and the solar system. This is comparison between same systems but also shoul be noticed
that with same systems at the building with lower specific heat consumption are achieved bigger
98
Chapter 6. Simulation results and analysis
solar fractions. Another comparison is performed for the same solar thermal systems in reagrd of
Primary energy consumption, kWh
the primary energy consumption given on Figure 48.
35000
30000
25000
20000
15000
Building 70
10000
Building 57
5000
0
Figure 50. Primary energy consumption for solar thermal systems in regard of two
buildings with different specific heat energy consumption 57 kWh/m2 a and 70 kWh/m2 a
Differences in primary energy consumptions are noticeable for the same solar systems applied
in the two buildings i.e. logically much lower primary energy consumption and lower CO2
emissions for the building with lower specific heat energy but it’s interesting that especially the
differences are expressed at the smaller collector areas with electrical heater as auxiliary device.
Also from the presented LCC analysis can be concluded that for same collector areas it is more
cost-effective the auxiliary source to be heat pump instead of electrical heater although the
investment is higher for the HP but on longer period is more feasible. On Figure 51 are given
specific costs per kW cooling power for absorption chillers. The costs are from manufacturers
based on Germany and from a survey of the international energy agency. According the presented
costs for the analysis is chosen that the absorption chiller as ne assembly with the cooling tower
costs 1500 eur/ kW i.e. total 15 000 eur.
Figure 51. Specific absorption chiller costs per kW cooling power [48]
99
Chapter 6. Simulation results and analysis
In the next Table 26 are presented results from the comparison between the energy
consumption, system costs, primary energy consumption and CO2 emissions for the chiller
conventional system and the solar assisted absorption cooling.
Table 26. Solar thermal system energy performance, costs, parameters and CO2 emissions,
primary energy consumption for building with cooling energy consumption 12 kWh/m2a
Conventional chiller cooling average EER 3 - 600 kWh/year
Parameter
Auxiliary energy - cooling system
Auxiliary energy - DHW
Absorption electrcal energy
System electrical energy
consumption (circ.pumps)
Specific electrical energy price
Total energy price
Heat source device/system price
Primary energy consumption
Annual CO2 emissions
Annual average solar fraction - SF
Solar thermal system - Area/Storage tank volume = 1000 l
Unit
kWh
kWh
kWh
Chiller
600
835
kWh
293
eur/kWh
eur
eur
kWh
kg/year
156
6500
5720
570
%
10/1000-EH 14/1000-EH 16/1000-EH
324
246
204
0
461
18/1000-EH 22/1000-EH
174
132
90
0.09
79
22
2896
289
46
2638
263
59
18
15000
2499
249
66
16
12
2400
239
71
2261
225
78
The performance data for absorption cooling machine are given in Appendix D. It is
assumed that the chiller average cooling coefficient of performance is 3 and regarding the building
cooling loads in the simulations is used the Building type II for which the annual cooling load is
1800 kWh. The DHW in the case of conventional chiller is heated 100% by electrical heater while
in the case of absorption machine the area of solar collectors are sufficient to provide 100% solar
fraction for the DHW i.e. it is not used any additional auxiliary energy.
The cooling period considers four months starting from June until September and the inside
temperature is set to be constant 26 °C.
The primary energy and CO2 factors are the same as previously i.e. are used from the standard
“Energy performance of buildings - Overall energy use and definition of energy ratings” EN
15603_2008 [46].
According the presented results can be concluded that there is no need to be performed LCC
analysis since the simple payback period is bigger than 20 years. Main reasons for the nonfeasibility of these solar-assisted air-conditioning systems mainly is the low electricity price of 0.09
eur/kWh but also is the low cooling energy demand since it is analyzed residential building which
usually have small internal gains. Nevertheless according the results in Table 26 the solar – assisted
cooling systems provides big possibilities for decrease in primary energy consumption and CO2
emissions.
100
Chapter 7
7. Summary and recommendations for further work
A solar system can be designed to satisfy any particular space and water heating application.
Technically considered it is feasible to design a system which can satisfy 100% of the heating
needs of a building, but generally it is economically not profitable solution. In practice solar
heating systems are designed to displace up to about 50% of conventional fuel needs and require
auxiliary heating systems that are fully capable of supplying the total heating load when no solar
energy is being collected and when solar energy has been depleted.
When solar heating is designed for space heating systems it is economical to include solar
heating of domestic hot water (DHW) in the system. In the residential systems during the summer,
when there is no space heating load, the entire solar system can be devoted to water heating so that
solar can supply a substantial portion if not all of the DHW heating needs. However, for large
commercial systems, summer operation to provide a relatively small DHW demand may not be
worthwhile.
The size of a solar system (primarily the storage volume and collector area) for a particular
building depends on the portion of the total load of the system is expected to provide. Size is also
strongly dependent from the climate and location
In this thesis were analyzed the thermal performances of solar assisted heating and cooling
systems. Main leading parameters for assessment of the system are the solar fraction of the heating
and DHW systems i.e. energy provided by solar collectors and auxiliary heat energy and the
collector efficiencies. Thus for the needs of assessment, simulation model of solar assisted airconditioning system was developed in TRNSYS. The analyzed system main components are: solar
collectors, storage tank, auxiliary heater building and hydraulic components.
First analysis was performed to examine the influence of the heating system type (radiator and
underfloor heating) to the system thermal performance i.e. to the solar fraction, collector efficiency
and “real” efficiency. In the analysis also were considered different: collector areas ( 16 m2, 32 m2
101
Chapter 7. Summary
and 64 m2), storage tank volumes (1000 l, 1500l, 2000 l) and three types of buildings (I, II, III) with
different specific heat consumptions 90 kWh/m2 a, 70 kWh/m2 a and 57 kWh/m2 a respectively.
Results from the simulation showed that for building type I i.e. specific heat consumption of 90
kWh/m2a, 16 m2 collector area, storage tank 1000 l and radiator heating solar energy can cover 8%
from the annual heat energy needs while the maximum solar fraction of 52% is achieved for
Building III (57 kWh/m2a) with underfloor heating system 64 m2 collector area and 2000l storage
tank. with 0,1 m2/m2conditioned specific collector area per conditioned surface can be achieved
between 20 – 25 % solar fraction, with 0,2 m2/m2conditioned range round 35% , and with 0,4
m2/m2conditioned maximum 50 %. It should be noted that the solar fraction also strongly depends from
the storage tank volume and for the radiator heating system increasing the storage volume results in
decrease of solar fraction while at the underfloor heating its vice versa. It is recommended the
storage volume to be in the range 50-60 l/m2 collector area in order to optimize between the solar
fraction and collector efficiency.
Further, analysis is done for solar assisted cooling system with vacuum tube collectors and
same volumes of storage tanks and building types as described previously. Obtained results show
that solar fractions increase with the increase of collector area and storage volume and vary in
range between 20% up to 70%. Also the thermal efficiency should not be neglected which in the
analyzed cases is in the range between 15% up to 27% monthly averages. Solar fraction for the
DHW is almost in every case 100% since the daily consumption is very low compared to the
available energy from the solar collectors.
Specific indicators for solar fraction in regard of the ration between solar collector area and
conditioned area are: 0,1 m2 / m2conditioned can cover almost 30% , 0,2 m2 / m2conditioned covers 50%
and 0,4 m2 / m2conditioned can cover 70% of the total required heating energy for driving the
absorption chiller. Analyzing the solar fraction for one constant specific collector area and
changing only the storage volume can be noticed that biggest fractions are for 30 l/m2 specific
storage tank volume per collector area for vacuum tube collectors.
Installing cold storage tank between the absorption chiller and building cooling system will
cause decrease in the solar fraction but also will decrease the fan cooling tower electrical
consumption. The decrease in solar fraction can be explained with the fact that the stored chilled
water causes decrease in charging/discharging the hot storage tank thus collector more of the time
is in stagnation.
The analysis is separated in two parts: first is solar collectors only for heating and second
analysis is solar collectors for cooling with absorption chillers. As referent system for comparison
102
Chapter 7. Summary
is selected electrical boiler i.e. all of the savings and payback periods are calculated against
electrical energy costs. Also in the comparison are considered pellet boiler and heat pump.
Results show that most profitable would be installing pellet boiler as heating source 100%
covering the heating needs, with payback period of 4 years. Second profitable measure would be
solar system with total collector area of 16 m2 and heat pump as additional auxiliary heat source.
Further follows the systems with collector array areas of 32 m2 and 64 m2 with different storage
tank volumes in combination with auxiliary heat device electric heater (EH) or heat pump (HP).
LCC analysis reveal that for same collector area and storage tank volume more profitable is the
combination solar collectors with heat pump instead with electric heater although the starting
investment is far lower for electric heater but energy savings contribute to have lower payback
period for the heat pump system.
Solar cooling is not yet feasible solution for the residential sector. Main reasons are the low
specific cooling energy consumption, low electricity price and high investment cost for the
absorption chiller. However, solar energy showed that can be achieved big primary energy savings
resulting with more that 50% compared with the conventional energy sources i.e. electrical energy.
Thus can be said that solar energy is environmental friendly also decreasing to a high extent the
greenhouse gasses emissions.
Further recommendation is to be researched the technical and economic feasibility of
implementation solar assisted air-conditioning systems in commercial buildings since they have
high internal loads thus offering potential for big energy savings.
103
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APPENDICIES
106
Appendix A
Technical data solar thermal storage, type: 15/9 S2, 12/8 S2, 13/7 S2
107
Dimensions
mm
a
b
c
d
e
f
g
h
i
j
k
l
m
n
p
ΦC
ΦD
15/9S2
2000
1927
1827
1287
497
2399
2263
1875
1560
1537
1380
1244
587
420
90
2565
1300
1100
12/8S2
1500
1768
1666
1168
468
2193
2061
1378
1081
579
421
90
2361
1200
1000
Legend
R
TS1
TS2
EE
T
TR
CW
IS2
OS2
IS1
OS1
HW
AV
108
Recirculation
Thermo pocket 1
Thermo pocket 2
Electric heating element
Thermometer
Thermoregulator
Inlet cold water
Inlet heat exchanger 2
Outlet heat exchanger 2
Inlet heat exchanger 1
Outlet heat exchanger 1
Outlet hot water
Air ventilation
Appendix B
Technical data flat plate solar collector
Camel Solar, type CS Full Plate 2.0-4
Flat plate collector "Camel Solar" CS full plate 2.0-4
Dimensions L x W x H
Absorber
Material
2005 x 1005 x 85 mm
Aluminum sheet and copper piping
Additional aluminum sheet (0.2mm)
ultrasonic welded on back side, two
times with absorber sheet one time and
once with riser tube
Joint absorber risers
Thickness
Surface treatment
Absorptance
Emittance
Heat transfer fluid content
Flow pattern
Dimension absorpber tubes
Number of absorber tubes
Distance between absorber tubes
Dimensions of the header
Number of connections
Transparent cover
Number
Outer diamter glass tube
Material
Transmittance
Thickness
Insulation
Material
Thermal conductivity
0.5mm
TiNOX
0.95
0.05
1.5 liter
Harp
8 x 0.4 mm
10
94 mm
22 x 0.8 mm
2
1
High borosilicate 3.3 glass
Tempered low iron glass
0.92
3.2mm
Rockwool
0.035 W/mK
109
Density
Thickness of insulation
Limitations
Stagantion temperature
Max. pressure
Allowed heat transfer fluid
Nominal flow rate per collector
Thermal performance data
Conversion factor of the beam irradiance, F'(τα)en
Factor to determine the incidence angle modifier of
the beam irradience, bo
Optical efficiency, ƞo
Heat transfer coefficient a1
Temperature depending heat transfer coefficient a2
Incidence angle modifier diffuse radiaton Kθd
Incidence angle modifier Kθ = 50°
Area related heat capacity c
Volume flow rate,
Peak power per collector unit G=1000 W/m2
110
100 kg/m³
back side 50 mm, sidewards 20 mm
197 °C
10 bar
Glycol/water mixture
90 kg/h
0.795
0.138
0.791
4.176 W/m²K
0.008 W/m²K²
0.988
0.935
13.19 kJ/m²K
72 l/m²h
1448 W
Appendix C
Technical data vacuum tube collector
“Camel Solar”, type CS 15
Dimensions L x W x H
Absorber
Material
1990 x 1180 x 158 mm
Glass
Aluminium heat transfer
sheet
1.5mm
CU/SS-ALN(H)/SS-ALN(L)ALN
0.92-0.96
0.04-0.06
2.95 Liter
Parallel
8 x 0.4 mm (U pipe)
15
76 mm
22 x 0.8 mm
2
Joint absorber risers
Thickness
Surface treatment
Absorptance
Emittance
Heat transfer fluid content
Flow pattern
Dimension absorber tubes
Number of absorber tubes
Distance between absorber tubes
Dimensions of the header
Number of connections
Transparent cover
Number
Outer diameter glass tube
Material
Transmittance
Thickness
Insulation
Material
Thermal conductivity
Density
Thickness of insulation
1
58 mm
High borosilicate 3.3 glass
0.92
1.5mm
Rockwool
0.035 W/mK
100 kg/m³
20 cm
111
Limitations
Stagnation temperature
Max. pressure
Allowed heat transfer fluid
Nominal flow rate per collector
Conversion factor of the beam irradiance, F'(τα)en
Factor to determine the incidence angle modifier of
the beam irradiance, bo
Optical efficiency, ƞo
Heat transfer coefficient a1
Temperature depending heat transfer coefficient a2
Incidence angle modifier diffuse radiation Kθd
Incidence angle modifier Kθ = 50°
Area related heat capacity c
Volume flow rate,
Aperture area per collector unit
Peak power per collector unit G=1000 W/m2
112
250 °C
10 bar
Glycol/water mixture
90 kg/h
0.695
0.138
0.738
1.725 W/m²K
0.01 W/m²K²
1.203
0.935
58.4 kJ/m²K
72 l/m²h
1.42 m²
1048 W
Appendix D
Parameters value for the absorption chillers
Numerical model Type 177 for different commercial available absorption chillers
Par.
Description
1
Unit
kJ/h
kJ/h K
kJ/h K
Calculation mode
2
rD,0
Zero order parameter for axis interval (desorber)
3
rD,I
First order parameter for axis interval (desorber)
4
s D,0
Zero order parameter for slope calculation (desorber)
5
s D,I
First order parameter for slope calculation (desorber)
6
rE,0
Zero order parameter for axis interval (evaporator)
7
rE,I
First order parameter for axis interval (evaporator)
8
s E,0
Zero order parameter for slope calculation (evaporator)
9
s E,I
10
B
kJ/h K
Dühring parameter
Minimum cooling water inlet temperature (i.e. range of validity for parameter
°C
Maximum cooling water inlet temperature (i.e. range of validity for parameter 2 to 13)
°C
Minimum hot water inlet temperature (i.e. range of validity for parameter 2 to 13)
°C
Maximum hot water inlet temperature (i.e. range of validity for parameter 2 to 13)
°C
Minimum chilled water inlet temperature (i.e. range of validity for parameter 2 to 13)
°C
Maximum chilled water inlet temperature (i.e. range of validity for parameter 2 to 13)
°C
Rated mass flow rate hot water (parameters 2 to 9 are valid in a range of +/- 10% )
kg/h
Rated mass flow rate chilled water (parameters 2 to 9 are valid in a range of +/- 10%)
kg/h
Rated mass flow rate cooling water (parameters 2 to 9 are valid in a range of +/- 10%)
kg/h
11
12
13
14
15
16
17
mD,0
18
mE,0
mAC,0
19
2
7
-47405
-4039
32134
-27
10233
-420
1710
-60
-1855
55
2384
2
140
1.2
26
47
74
116
5
13
36935
65641
89137
-14
1.2
22
42
49
106
7
22
1171
2898
2590
-41
1.2
26
37
59
96
5
25
1940
1999
4975
kJ/h K
kJ/h
kJ/h K
kJ/h K
First order parameter for slope calculation (evaporator)
113
YIA 1A1 Suninverse Wegra 15
1
1
1
-214163
-756
-15654
4443
398
978
47555
2359
3203
Appendix E
Cooling tower performance data – Baltimore aircoil model PF2-0406AA-31-3
114
Appendix F
- TRNSYS model of the analyzed solar assisted air-
conditioning system
115
Appendix G
-
Draft proposal for scientific paper
116
Techno-economic feasibility analysis for application of small solarassisted air conditioning systems in Macedonia
Igor Sheshoa*1, Laurent Georgesb, Vojislav Novakovicb,1, Done Tashevskia
a
University “SS.Cyril and Methodius University of Skopje”, Faculty of Mechanical Engineering Skopje, Karpos II bb, 1000 Skopje, R.Macedonia
Norwegian University of Science and Technology, Department Energy and Process Engineering, Kolbjørn Hejes vei 1,Trondheim 7491 Norway,
b
Abstract
This paper has objective, due to climatic conditions in Macedonia, to estimate the thermal performance of solar assisted
air conditioning systems in regard of solar fraction, primary energy and perform life cycle cost analysis to assess the
feasibility of their implementation in residential sector. The analysis is performed using numerical methods within
TRNSYS software, since the solar processes are transitient in nature been driven by time dependent forcing functions
and loads represented with component models. Validation is made for the used TRNSYS components models i.e. the
simulation results are compared with experimental measurements where the modelled and measured solar collectors are
Macedonian product. Reference building is modelled representing the heating and cooling energy consumer. The
building is modelled in three types regarding the annual specific heating and cooling energy consumption. The obtained
results reveals that the solar fraction for heating and for driving absorption chiller as ratio between the total collector
and conditioned area. Solar collector systems applied only for heating and DHW with heat pump as auxiliary heat
source and building with specific heat consumption of 57 kWh/m2a, have payback period starting from 7,5 years for 16
m2 collector area and 1000l heat storage tank, while the solar assisted cooling system all of the NPV values are negative
indicating that economically without subsidies is not profitable measure. Solar assisted cooling is not feasible since the
electricity price in Macedonia has low price and also the absorption chiller has relatively high investment cost
offsetting the payback period to higher values. But it must be noted that regarding primary energy consumption solar
assisted air conditioning systems for conditions in Macedonia can provide more than 50 % savings compared to
conventional energy sources i.e. electrical energy.
Keyword:Solar assited air conditioning, primary energy, solar fraction, simulaton,TRNSYS
1. Introduction
In the past twenty years the uncontrolled or mildly said “demanded” industry development accompanied by the
increased thermal comfort demand and the limited energy resources naturally activated and imposed the need of the
forgotten term of energy efficiency. In the past, the cheap/affordable energy sources, high building energy consumption
and the modest technology development were the main limiting factors for the viability of utilization renewable energy
sources.
Undertaken measures started with limiting the energy consumption through different directives and regulations which
translated into actions meant: improving efficiencies of existing systems, decreasing the energy demand, up to
developing new technologies, and all of that with one purpose - doing more with less. Now with the implementation of
the buildings directives such as the Energy Performance Building Directive 2002/91/EC the EU 2020 strategy,
technology development enabled the renewable energy sources to be feasible for implementation in buildings airconditioning systems. European commission adopted communication “Energy 2020” that defines new strategy toward
2020 for a competitive, sustainable and secure energy [1]. A significant contribution to the primary energy
consumption of first and second world countries is being made by the rapidly increasing use of electrical airconditioning units worldwide. Worldwide sales of room air-conditioners of all types amount to approximately 50 Mio.
units p.a., with the U.S. China and Japan being the three main markets. In Europe, commercial air conditioning has a
share of 4% of the total annual electricity consumption while residential air conditioning accounts only for 0.4%.
Although the latter number is still comparatively low, Europe has seen a seven-fold increase of residential airconditioning sales between 1990 and 2004. The reasons for the growing use of air-conditioning are twofold. First, the
comfort demands from both building users and owners have increased. The standard of living of the present generation
is higher than in the past, especially in private buildings. . Second, the trend towards commercial buildings with large
glazed facades has increased the internal heat load to be removed by air-conditioning. Third, electricity prices are
1
Corresponding author. Tel: +38923099200,
e-mail address: [email protected] (I.Shesho)
1
comparatively low. The additional cost caused by the use of air-conditioning units is not in the order of magnitude to
influence the consumer behavior significantly.
The obvious consequence of this growing air-conditioning use is increased power consumption. Outside Europe,
another consequence of excessive air-conditioning are locally higher temperatures in metropolitan areas, commonly
referred to as heat islands. As a part, these inner-city temperature peaks are the result of heat conveyed from building
inside to outside, released at a temperature level above ambient temperature. Both these consequences are strong
arguments for alternative air-conditioning or cooling methods.
Priority is given to buildings and transport sector. Through directives to improve building energy efficiency (energy
performance of buildings directive or EPBD), the European Union recognized high potential for energy savings from
buildings and promote the installation of solar thermal systems in the building sector. Solar technologies can supply the
energy for all of the building’s needs—heating, cooling, hot water, light and electricity—without the harmful effects of
greenhouse gas emissions created by fossil fuels thus solar applications can be used almost anywhere in the world and
are appropriate for all building types.
Solar thermal systems for hot water production are already mandatory in new buildings according to solar ordinances
for example in Spain [2], Portugal, Italy, Greece and other European countries [3].
Systems combining production of domestic hot water (DHW) and space heating systems are well suited to middle and
high latitudes, due to significantly higher solar radiation in the transitional period around winter (September-October
and March-May) and the significant heating demand in these latitudes at that time [4].
Installations with large solar collector areas and small size heat storage capacity can cover around 50% of the total heat
demand. This percentage can be higher in some cases of large storage capacities and primary energy savings up to 80%
[5]. Simulations of central solar heating plants with seasonal storage (CSHPSS) have shown that the solar fraction of
such systems varies between 50% and 100% [6, 7]. The heat produced by the collectors may be stored in thermal
energy storages in order to provide domestic hot water and space heating when required [8].
2.
Developing simulation model of the analysed solar thermal assisted air-conditioning system
Assessment of thermal performance of the analyzed solar air-conditioning system is performed through a dynamic
simulation model with transient behavior implemented via thermal and mass storage terms as well as delay times. The
model i.e. analyzed system, generally consists of four main subsystems shown, as follows:
4.
5.
6.
7.
First subsitem is composed of solar collectors with complete hydraulic fittings and control - differential
controllers, plate heat exchangers ie this system is represented the source of thermal energy for heating or
thermal enrgy for driving the cooling the absorption machine
Second is the subsystem for hot and cold storage which includes the storage tanks for hot / cold water that
actually represents the connection between the heating system in the building ie absorpcionata cooling
machine and the source of heat.
The heating system introduced with heating / cooling devices, hydraulic armature heat exchangers and
cooling absorption machine and eventually existing conventional sources of heat and / or cooling energy.
The fourth subsystem is the consumer of thermal energy ie the building . This system is represented by
the thermal characteristics of the object, ie its orientation in space.
Figure 1. Functional scheme presenting inter conections between components of the system moddeled in TRNSYS
2
On Figure 1 is presented the analyzed i.e. modeled and simulated solar assisted air-conditioning system. The main
system components are: the solar collector array, two storage tanks, auxiliary heater, absorption chiller and the energy
consumer i.e. the building which also incorporates the heating/cooling system components.
The working fluid from the solar collectors indirectly through heat exchangers is used to heat the domestic hot water
in tank 3 or heat the fluid in the storage tank 4 further used as part of the heating energy in the building or part of the
driving heat for the absorption chiller in summer. The circulation of the solar collectors working fluid for the storage
tanks 3 and 4 is done by two separate circulating pumps P1 and P2, controlled by two differential controllers having
mutual predefined control function further explained. First condition for the pumps to be switched on is the temperature
difference between the collector outlet temperature and the fluid temperature in storage tank (3 or 4) to be greater than
the set upper dead band. The control logic for switching between the two tanks is solved using the two controllers Type
2b (K1 and K2) one flow diverter Type 11f . The advantage has the controller K2 of the tank 4 i.e. the initial input
control signal (on/off) for the controller of the DHW tank K1 is received from the controller K2 i.e. when the controller
K2 is on, then the controller K1 is off.
In the analyses are considers vaccum tube and flat plate collectors and the Building II type, with thermal
performance described in Table 4. The building internal heat gains consider the lighting power density 5 W/m2 and the
home appliances with specific power of 2 W/m2. The absorption chiller condenser is connected to the wet cooling
tower product of Baltimore AirCoil type PF2-0406AA-31-3. Numerical modelling of the cooling tower is provided by
the TRNSYS Type510 model from Tess library, a closed circuit cooling tower which cools the liquid stream by
evaporating water from the outside of coils containing the working fluid. The working fluid is completely isolated from
the air and water in this type of system.
The control signal of the cooling tower fan is set to have the tower try to maintain the desired outlet water
temperature of 26 °C and the fluid flow rate with the circulation pump is set to 2600 kg/h. Also to the circulation of the
cooling water are modeled pipe network with diameter 0.04 m and length of 15 m heat transfer coefficient for thermal
losses 3 kJ/h m2 K which accounts for the heat losses to the environment and also increases the system thermal capacity
affecting the simulation stability. Values of the inlet parameters for the cooling tower such as ambient air temperature
and relative humidity are read from the weather component respectively for the simulated time and period of the year.
The cooling water temperature is parameter which is variable depending from the absorption chiller working
conditions.
The cooling system in the building is modeled using the ventilation air distribution system. The combination among
the chilled water from the absorption chiller and the ventilation air is provided with heat exchanger water-air modeled
Type 508a which is a cooling coil modeled using a bypass approach in which the user specifies a fraction of the air
stream that bypasses the coil. The remainder of the air stream is assumed to exit the coil at the average temperature of
the fluid in the coil and at saturated conditions. The two air streams are remixed after the coil. Chilled water flow from
the absorption chiller to the cooling coil is set to 2900 kg/h and the air flow rate to the building is 4000 kg/h. The
auxiliary heater power is modeled 12 kW and the outlet temperature is 80 °C, which is the absorption machine driving
temperature.
The storage tank 4 in this case is used to store the heat for driving the absorption chiller. With the thermostat Type
108 is regulated the space temperature in the building set to 26 °C, which control signal is directly regulating the
function of the circulation pump from the chilled water absorption chiller and the fan distributing the conditioned air.
Simulation time step is 15 min and as cooling period are considered the months from May-September.
The collector(s) thermal efficiency in the simulation is determined using the equation component from the TRNSYS
library. The equation considers ratio between the useful energy gain from the all of the collectors transferred to the
fluid and the total tilted radiation for the collector surface. The data for the quantity of useful energy gain and total
radiation in the equation is read from the quantity integrator which integrates these values in the predefined period
defined from the required value period thermal efficiency and energy i.e. daily, weekly, monthly, yearly or any other
time interval.
2.1 System component models definition and validation
2.1.1 Solar system component validation
The validation is performed for the following solar thermal system components: solar collector, storage tank and
differential controller. The system consist one flat plate solar collector connected with the internal heat exchanger of
the storage tank. Control is provided by differential controller which is set to turn the circulation pump on, when the
temperature difference between the collector outlet temperature and the tank temperature is greater than five. The water
from the storage tank is not discharged and the electric heater is turned off during the measurements. The fluid (water)
flow rate is set to 7,5 lit/min.
The measurements are made on an hour interval for the fluid inlet T1 and outlet T2 temperatures from the solar
collector, tank fluid temperature T3 and the solar radiation measured with the pyrometer S. The experimental setup of
3
the analyzed solar thermal system is placed in Skopje, R.Macedonia, northen latitude of 42° and 21.43° east longitude.
The temperature measurements are performed with temperature data logger thermocouple probes type K.
The solar collector is evacuated tubular direct flow product of Camel Solar type Vacuum CS 15 Solar KeyMark
certified. It is installed under tilt of 45°, south orientated i.e. azimuth angle of 0°. The collector thermal performance
test results made according EN 12975 are presented in Table 1 while the storage tank technical data are given in Table
2
Table 1. Technical data for collector type Camel Solar Vacuum tube SC 15 ( “U” type)
Dimensions L x W x H
mm
1990 x 1180 x 158
Number of asborber tubes
-
15
Absorptance, α
-
0.92-0.96
Emmitance, ε
Conversion factor of the beam
irradiance, F'(τα)en
Factor to determine the incidence
angle modifier of the beam irradience,
bo
-
0.04-0.06
-
0.695
-
0.138
Optical efficiency, ƞ o
Heat transfer coefficient a1
Temperature depending heat transfer
coefficient a2
Incidence angle modifier diffuse
radiaton Kθd
Incidence angle modifier Kθ = 50°
Area related heat capacity c
Volume flow rate,
Apperture area per collector unit
Peak power per collector unit G=1000
2
W/m
-
0.738
2
1.725
W/m K
2
2
0.01
W/m K
-
1.203
-
0.935
2
kJ/m K
l/m²h
58.4
m²
1.42
72
W
1048
. Table 2. Storage tank technical details
l
150
Height
H, mm
1210
Diameter
D, mm
560
mm
50
Capacity
Insulation, rigid PU
Coil capacity
Heat exchanger surface
Prolonged power according DIN
4708 80/60/45
l
4.56
2
m
kW
0.74
m³/h
0.61
25
-
2.5
Coil outlet
L, mm
202
Cold water inlet
A, mm
202
Sensor sleeve for thermostat
G, mm
822
Coil inlet
K,mm
592
Hot water outlet
E,mm
868
NL-power coeficient at 60°C
In the TRNSYS model for the solar collector model is used the collector Type 538 from the Tess library modeled with
the technical data given in Table 1 The storage tank is modeled with the Type 60d including the internal heat
exchanger for which are supplied data from Table 2. Type 2b-2 is used for the differential controller with upper dead
band of 5 and lower dead band 2, the high limit cut-off temperature is set to 100 °C. Between the solar collector and
storage tank is connected pipe Type 31 modeled with internal diameter 0.0025 m, length of 10 m and loss coefficient of
0,3 W/m2K to account for the heat losses. The pipe Type 31 beside to account for the heat losses in the pipes also is
used in order to increase the thermal capacity of the system and thus increase the simulation stability. Also for the
circulating pump is used the Type 3d with mass flow rate 450 kg/h i.e. 7,5 l/min same as in the experimental setup.
Measurements are performed starting from date 18.09.2013 until 28.03.2014 and in parallel are measured two systems
with same capacity storage tank of 150l but different type of collectors i.e. flat plate and vacuum tube solar collectors.
In the validation process are used the data for the vacuum tube collector and the results from only one day period
(18.09.2014) with collection time interval ranging between 20min and 45min interval, starting from 10:40 until 16:05
h.
4
Temperature, °C
70
60
50
40
30
11:15
12:20
13:10
13:45
14:25
14:13
16:05
Time [h:min]
Measure T1
Simulated T1
Temperature, °C
Figure 2. Measured and simulated temperatures for the collector inlet
70
65
60
55
50
45
40
35
10:43
12:00
12:30
13:30
14:05
14:43
15:35
Time [h:min]
Measured T2
Simulated T2
Temperature, °C
Figure 3. Measured and simulated temperatures at the collector outlet
70
60
50
40
30
11:15
12:20
13:10
13:45
14:25
14:13
16:05
Time [h:min]
Measured T3
Simulated T3
Figure 4. Measured and simulated temperatures inside storage tank
5
Global solar radiation, W/m²
1200
1000
800
600
400
200
0
Time [h:min]
Measured radiation S1
Simulated radiation S1
Figure 5. Hourly measured and simulated solar radiation for the specific day
According above presented data i.e. diagrams can be concluded that there is acceptable match between the measured
and simulated results. The discrepancies that appears between the temperatures of the experimental and simulated
results are expected since the solar radiations have different values because one values are obtained with measurements
and the others are from Meteonorm database for the selected location. Another influencing factor is the uncertainty of
the measurements error and last but not the least it should not be neglected the transition nature of the solar thermal
systems.
The resulting simulations reveal the individual thermal behaviour of the solar collector, storage tank, differential
controller and circulating pump as well as their assembled thermal behaviour. These results were very close to their
corresponding experimental data and this fact validates these models for future application in the heating/cooling
system.
2.1.2 Absorption chiller component validation
Validation for the absorption chiller is made for the TRNSYS component Type 177. This component type offers four
numerical modes of absorption chiller, and in this simulation is used the mode “a” i.e. Type177a which is standard
mode using user supplied characteristic parameters. Since in this paper are considered only solar air-conditioning for
residential buildings, in Table 3 are given the technical data for several small absorption chillers. From the presented
absorption chillers, in the simulation is modeled the absorption chiller H2O/LiBr produced by Sonnenklima type
Suninverse 10. n the component Type 177a as input parameters are taken the values for Suninverse provided in Table 3
for which with the simulation as output cooling power is obtained value of 10,1 kW, corresponding with the factory
value. Validation exists for the Type 177 mode “d” performed by Albers and Ziegler [10], using the measurement
results from Kühn [11]. According to the this can be concluded that this numerical model of absorption chiller provides
reliable results and can be used further in simulations.
Table 3. Technical data for different market available small absorption chillers
Company
Product name
Technology
Working pair
Cooling capacity, kW
Heating temperature, °C
Recooling temperature, °C
Cold water temperature, °C
COP
Weight, kg
Electrical power, W
Yazaki
Sonnenklima
Rotarica
WFC-SC5,
chillii WFC 18
Absorption
Suninverse 10
Absorption
Solar 045
Absorption
H2O/LiBr
17.6
88 / 83
31 / 35
12.5 / 7
0.70
420
H2O/LiBr
10
75/65
27/35
18/15
0.77
550
72
120
H2O/LiBr
4.5
90/85
30/35
13/10
0.67
290
1200
(incl.ventilator)
6
2.1.3 Reference building model description
Building as energy consumer has a major impact on the overall efficiency of the solar system i.e. can be freely said
that the building itself is one of the leading parameter in sizing the system. Since the analyses are made for climatic
conditions in Macedonia also the thermal performance of buildings must be in accordance with the Regulations on
energy efficiency in Macedonia. Furthermore the analysis is taken into account the impact of the specific consumption
of heating / cooling energy of the building kWh/m2 a to the response and the performance of the solar collector system.
In Table 4 are listed three types of the building i.e. the dimensions and orientations are unchanged only the insulation
thickness is varied in order to obtain different values for specific annual consumption of thermal energy. The main
motive for variations in the thickness of the insulation is to analyze the influence of the thermal performance of
buildings on the economic viability of the use of solar thermal systems in air-conditioning. Constant value of 0.3 1/h is
defined for the infiltration of outdoor air, while for the summer when cooling is required in the building is
envisaged/modeled mechanical ventilation defined with air mass flow and temperature entered through \the models of
fan and heat exchanger air-water which is directly connected with the cooling absorption machine.
Regarding the thermal comfort, in the heating mode the inside temperature is defined to be 20 °C from 05:00 – 22:00
and for the rest is defined setback temperature of 16 °C, for the cooling mode is defined constant inside temperature of
26 °C.
Table 4. Reference building physical and thermal performance data
Building I Building II Building III
U value, W/m²K
Surface
Orientation
Area, m²
Out.wall 1
North
42
0.58
0.33
0.18
Windows 1
North
3
1.40
1.40
1.40
Out.wall 2
East
25.5
0.58
0.33
0.18
Windows 2
East
4.5
1.40
1.40
1.40
Out.wall 3
West
25.5
0.58
0.33
0.18
Windows 3
West
4.5
1.40
1.40
1.40
Out.wall 4
South
42
0.58
0.33
0.18
Windows 4
South
3
1.40
1.40
1.40
Floor
150
0.33
0.33
0.24
Roof
150
0.54
0.42
0.35
Double glazed TRNSYS library (w4-lib data)
Window type
Windows solar heat
0.589
gain coefficient;g-value
Insulation Insulation Insulation
Out.wall construction 2 x Plaster 2cm, brick 25cm
5 cm
10 cm
20 cm
Granite tile 6cm, cement
Insulation Insulation Insulation
Floor
mortar 5cm , concrete slab
10 cm
10 cm
15 cm
20cm
Insulation Insulation Insulation
Concrete slab 20cm, hydro
Roof
isolation, cement mortar 5cm
15 cm
20 cm
25 cm
Outside convective heat transfer
αout = 25 W/m²K
coefficient
Inside convective heat transfer
αin = 7,7 W/m²K
coefficient
4000.00
kWh
3000.00
2000.00
1000.00
0.00
1
2
3
4
5
6
7
8
9 10 11 12
Month
Building I
Building II
Building III
Figure 6. Monthly energy consumption for Building I, II and III
Analyzing the presented simulation results on Calculation of energy consumption in the building is obtained directly as
output size of the numerical model of the object in kJ / h value which further is integrated for the required period with
the quantity integrator. Also as output parameters of the model is the output temperature of floor heating, the
7
temperature of the air entering the fan to the heat exchanger and air-water and the delivered energy from the underfloor
heating system into the building.
Monthly analysis is performed for the building heat energy consumption regarding different heat transfer coefficients
i.e. different wall, floor and roof isolation thickness thus defining three types of Building I, II and III. as presented in
Figure 6.
It can be noticed that as expected the Building III has the smallest heat consumption i.e. regarding specific annual
energy consumption, Building I has 90 kWh/m2a , Building II with 70 kWh/m2a and Building III has 57 kWh/ m2a.
Comparing the energy consumption Building III has 42% lower than Building I and 19% than Building II.
As mentioned previously for the absorption chiller it is used the component model Type 177a in which input
parameters are inserted the data from the LiBr/H2O Suninverse 10 chiller product of Sonneklima with cooling power of
10 kW with heat driving temperature of 75 °C, cooling water 27 °C and chilled water outlet temperature set at 15°C .
3.
Simulation results and discussion
The comparison results between the solar fractions and efficiencies in regard of different collector areas (flat plate
collectors) and storage volumes are presented in Table 5. The mass flow rate is set constant according the collector area
i.e. it is 50 kg/h m2. Solar collectors are tilted on 40° with south orientation i.e. azimuth is 0°. As can be seen the solar
fractions increase with the increase of collector area and storage volume and varies in the range between 20% up to
70%. Also the thermal efficiency should not be neglected which for the analyzed cases is in the range between 15% up
to 27% monthly averages. Solar fraction for the DHW is almost in every case 100% since the daily consumption is very
low compared to the available energy from the solar collectors.
Table 5. Monthly average solar fractions and efficiency in regard of collector areaa and storage volume
Flat plate, Tilt 40°, Azimuth 0° - flow rate 50 l/h m²
1000
1500
2000
Collector area m²
16
32
64
16
32
64
16
32
64
Sol.fraction
0.28
0.50
0.67
0.24
0.49
0.69
0.19
0.48
0.70
Avg_Efficiency
Avg_Eff_Real
Sol.DHW
0.26
0.48
0.98
0.19
0.36
0.99
0.15
0.32
0.99
0.26
0.47
0.98
0.20
0.37
0.99
0.15
0.33
0.99
0.27
0.46
0.98
0.21
0.37
0.99
0.16
0.30
0.99
According the above presented results can be concluded that with solar energy regarding the specific collector areas can
be covered: 0,1 m2 / m2conditioned can cover almost 30% , 0,2 m2/m2 conditioned covers 50% and 0,4 m2/m2
conditioned can cover 70% of the total required heating energy for driving the absorption chiller. Analyzing the solar
fraction for one constant specific collector area and changing only the storage volume can be noticed that biggest
fractions are for specific volume per collector area of 30 l/m2.
Further analyzed the influence of the collector tilt angle with and without azimuth tracking system (one axis-vertical
tracking system) to the solar fraction and thermal efficiency. Simulation results are presented in
Table 6.Solar fractions and thermal efficiency for solar assited cooling system in regard of collector orientation i.e.
azimuth
Storage tank 1000l - Collector mass flow rate 50 l/h m²
Collector area
m² / tilt °
Sol.fraction
Avg_Efficiency
Avg_Eff_Real
16/1/40
16/1/30
0.28
0.26
0.48
0.30
0.26
0.50
16/1/30/T
32/1/40
32/1/30
32/1/30/T
0.44
0.53
0.55
0.67
0.24
0.18
0.18
0.19
0.50
0.38
0.38
0.41
Storage tank 1500l - Collector mass flow rate 50 l/h m²
64/1/3200/40 64/1/3200/30 64/1/3200/30/T
0.67
0.15
0.32
0.69
0.15
0.33
0.77
0.16
0.35
Collector area
16/1/40
16/1/30
16/1/30/T
32/1/40
32/1/30
32/1/30/T
64/1/3200/40 64/1/3200/30 64/1/3200/30/T
m² / tilt °
Sol.fraction
0.22
0.24
0.38
0.49
0.51
0.66
0.67
0.69
0.81
Avg_Efficiency
0.26
0.26
0.26
0.20
0.20
0.20
0.15
0.15
0.15
Avg_Eff_Real
0.46
0.47
0.47
0.37
0.38
0.39
0.32
0.33
0.36
*64/1/3200/30/T - (64) m² collector array , (1) all in parallel connected, (3 200 ) kg/h mass flow rate, (30) tilt angle, (T) tracking azimuth
In order to have better visibility of the parametric influence of the collector tilt angle and azimuth influence, on Figure
7 are presented monthly values for solar fractions from June-September
8
Solar fraction
0.50
0.40
0.30
0.20
0.10
June
16/1/30
July
August
16/1/40
September
16/1/30/T
Figure 7. Monthly values of solar fraction for different tilt angles and tracking azimuth for solar assited cooling system
3.1 Life cycle cost analysis of solar cooling system
The economic evaluation of a solar application includes factors such as the capacity cost of delivering solar energy, the
optimum sizing of collectors and other equipment the costs of competing technologies and financial analyses. There
figures of merit that are used to accept or reject particular solar application including simple payback, cash flow ,
capital cost per unit of energy saved, life cycle cost, net present value and levelized energy cost. For a range of
economic assumptions, an analysis considering a 5-year or 7 year simple payback is assumed to be an adequate figure
of merit to establish cost goals for active solar cooling and heating technologies. For residential applications, a payback
period of 5 - 7 years may consider as acceptable.
Comparison is performed between different combinations of solar thermal systems and auxiliary heating devices in
regard of conventional heating system with electrical boiler. The analyzed solar systems have total solar collector area
of 16 m2, 32 m2 and 64 m2 combined with storage tanks of 1000 l, 1500 l and 2000l , and auxiliary heating energy
provided by electric heater or heat pump air-water with E.V.I compressors with nominal capacity of 15 kW product of
Hidros model Lzti 10 [12].
In Table 7 are presented data for delivered heating energy, annual heating energy cost, system price and the
environmental indicator - CO2 emissions for conventional heating systems with heat sources: electrical energy, wood
pellets and heat pump. The heat pump COP is assumed with averaged yearly value of 2,5.
Table 7. Annual energy balance, energy and system costs, CO2 emissions for conventional heat source systems
obtained with simulations, specific heat energy consumption 70 kWh/m2 a
Parameter
Unit
Electrical boiler
Pellet boiler
Heat pump
Heat power
kW
12
Annual delivered heat energy
Average thermal efficiency/COP
Annual consumed energy
System electrical energy
consumption (circ.pumps)
Specific energy price
Total energy price
Heat source device/system price
Annual CO2 emissions
kWh
kWh
13103
0.99
13235
13103
0.91
14399
13103
2.5
5241
144
144
144
0.09
1203
800
12177
0.05
667
2000
58
0.09
472
5000
1730
kWh
eur/kWh
eur
eur
kg/year
12
12.5
The CO2 emissions and primary energy consumption factors are provided from the standard “Energy performance of
buildings - Overall energy use and definition of energy ratings” EN 15603_2008 [13].
Regarding the energy prices, the electrical energy price (euro/kWh) is provided from the Energy Regulatory
Commission of the Republic of Macedonia valid from 01.07.2014 , wood pellet energy price is obtained by taking the
average market price of 0,2 euro/kg with lower heating value Hd = 4,3 kWh/kg.
In Table 8 and Table 9 are given results from the analyzed solar thermal systems in regard of heat energy consumption,
annual energy and system costs and the environmental impact indicator presented by the value for the annual CO2
(kg/year) annual emissions.
9
Table 8. Solar thermal system energy performance, costs, parameters and CO2 emissions, primary energy consumption
for building with specific heat energy consumption 70 kWh/m2 a, – part I
Ref.building 70 kWh/m² a
Parameter
Auxiliary energy - heating system
Auxiliary energy - DHW
System electrical energy
consumption (circ.pumps)
Specific electrical energy price
Total energy price
Heat source device/system price
Primary energy consumption
Annual CO2 emissions
Annual average solar fraction - SF
Unit
kWh
kWh
Solar thermal system - Area/Storage tank volume - EH (electric heater) ; HP (heat pump) 16/1000-EH 16/1000-HP 32/1000-EH 32/1000-HP 64/1000-EH 64/1000-HP 32/1500-EH
8550
8550
7420
7420
6466
6996
6466
750
562
428
639
kWh
eur/kWh
eur
eur
kWh
kg/year
-
144
144
114
114
90
90
130
850
3600
28777
3069
348
7000
12790
1275
729
6000
26798
2672
0.09
300
10500
10946
1091
629
10000
23117
2305
256
14700
9426
940
640
7000
25702
2562
0.34
0.19
0.30
0.39
Table 9. Solar thermal system energy performance, costs, parameters and CO2 emissions, primary energy consumption
for building with specific heat energy consumption 70 kWh/m2 a, – part II
Parameter
Auxiliary energy - heating system
Auxiliary energy - DHW
System electrical energy
consumption (circ.pumps)
Specific electrical energy price
Total energy price
Heat source device/system price
Primary energy consumption
Annual CO2 emissions
Annual average solar fraction - SF
32/1500-HP 64/1500-EH 64/1500-HP 32/2000-EH 32/2000-HP
6996
5830
5830
7250
7250
639
490
Unit
kWh
64/2000-EH 64/2000-HP
5921
5921
kWh
eur/kWh
eur
eur
kWh
kg/year
130
103
103
138
138
111
111
287
11500
10539
1051
578
12000
21260
2120
237
16500
8709
868
0.09
665
8000
24454
2438
273
12500
10056
1003
543
12800
19966
1991
223
16700
8207
818
-
0.34
0.45
0.32
0.44
Table 10. Solar thermal system energy performance, costs, parameters and CO2 emissions, primary energy
consumption for building with specific heat energy consumption 57 kWh/m2 a, – part I
Ref.building 57 kWh/m² a
Parameter
Auxiliary energy - heating system
Auxiliary energy - DHW
System electrical energy
consumption (circ.pumps)
Specific electrical energy price
Total energy price
Heat source device/system price
Primary energy consumption
Annual CO2 emissions
Annual average solar fraction - SF
Unit
kWh
kWh
Solar thermal system - Area/Storage tank volume - EH (electric heater) ; HP (heat pump) 16/1000-EH 16/1000-HP 32/1000-EH 32/1000-HP 64/1000-EH 64/1000-HP 32/1500-EH
6708
6708
5762
5762
5418
6192
5418
713
527
402
595
kWh
eur/kWh
eur
eur
kWh
kg/year
139
139
109
109
86
86
115
680
3600
22664
2449
280
7000
10285
1025
576
6000
21177
2111
0.09
239
10500
8687
866
532
10000
19549
1949
217
14700
7990
797
567
7000
22846
2278
0.28
0.33
0.22
0.39
Table 11. Solar thermal system energy performance, costs, parameters and CO2 emissions, primary energy
consumption for building with specific heat energy consumption 57 kWh/m2 a, – part II
Ref.building 57 kWh/m² a
Parameter
Auxiliary energy - heating system
Auxiliary energy - DHW
System electrical energy
consumption (circ.pumps)
Specific electrical energy price
Total energy price
Heat source device/system price
Primary energy consumption
Annual CO2 emissions
Annual average solar fraction - SF
Unit
kWh
kWh
Solar thermal system - Area/Storage tank volume - EH (electric heater) ; HP (heat pump) 32/1500-HP 64/1500-EH 64/1500-HP 32/2000-EH 32/2000-HP 64/2000-EH 64/2000-HP
6192
5600
5600
5074
5074
4042
4042
595
457
630
488
kWh
eur/kWh
eur
eur
kWh
kg/year
115
94
94
118
118
97
97
255
11500
9367
934
554
12000
20360
2030
227
16500
8331
831
0.09
524
8000
19271
1921
216
12500
7943
792
416
12800
15315
1527
172
16700
6319
630
0.28
0.36
0.41
0.53
Further is analyzed the building energy consumption influence on the solar system performance i.e. same solar airconditioning system is applied to building with lower specific heat energy consumption. Analysing the presented
results is indicative that improved and feasible combination would be with heat pump instead of electrical heater which
10
is further verified with LCC analysis. The LCC method provides a better assessment of the long-term cost-effectiveness
of the project/measure than other economic methods which only focus on initial investment costs and operation costs
for the first years.
Same time, a LCC analysis requires more information than other profitability evaluation methods. The lowest Life
Cycle Costs indicates the most profitable investment, measure or solution
Table 12. Life Cycle Cost Analysis for solar thermal system assited buildign heating with specific .heat energy
consumption 70 kWh/m2 a
Real interest rate 3,9% ; Building 70 kWh/m2 a
Measures
Investment Net savings
EUR
EUR/Year
Pellet boiler
2000
506
Heat pump
5000
671
HP 16/1-1000
7000
814
HP 32/1-1000
9700
849
EH 16/1-1000
3600
312
HP 32/1-1500
10700
864
EH 32/1-1500
7000
511
EH 32/1-1000
6000
421
EH 32/1-2000
8000
487
HP 64/1-1000
14700
881
HP 64/1-1500
16500
902
HP 32/1-2000
16500
868
HP 64/1-2000
17300
896
EH 64/1-1000
10000
509
EH 64/1-1500
12000
560
EH 64/1-2000
12800
596
Lifetime
Year
20
20
20
20
20
20
20
20
20
20
20
20
20
20
20
20
PB
Year
4
7.5
8.6
11.4
11.5
12.4
13.7
14.3
16.4
16.7
18.3
19
19.3
19.7
21.4
21.5
PO
Year
4.4
9
10.7
15.5
15.7
17.3
20
21.3
26.9
27.6
32.9
35.6
36.8
38.3
47.7
48
IRR
%
25
12
10
6
6
5
4
3
2
2
1
0
0
0
0
0
NPV
EUR
4.93
4.18
4.14
1.919
670
1.125
-7
-239
-1.335
-2.643
-4.156
-4.621
-5.038
-3.034
-4.336
-4.644
NPVQ
2.46
0.84
0.59
0.20
0.19
0.11
0.00
-0.04
-0.17
-0.18
-0.25
-0.28
-0.29
-0.30
-0.36
-0.36
˟Max.Investment
EUR
Year
4.12
12
6.34
12
7.687
12
8.017
12
2.946
12
8.159
12
4.825
12
3.975
12
4.599
12
8.319
12
8.518
12
8.196
12
8.461
12
4.806
12
5.288
12
5.628
12
Table 13. Life Cycle Cost Analysis for solar thermal system assited buildign heating with specific .heat energy
consumption 57 kWh/m2 a
Real interest rate 3,9% ; Building 70 kWh/m2 a
Measures
Investment Net savings
EUR
EUR/Year
Pellet boiler
2000
314
Heat pump
5000
571
HP 16/1-1000
7000
702
EH 16/1-1000
3600
301
HP 32/1-1000
9700
730
HP 32/1-1500
10700
715
EH 32/1-1000
6000
393
HP 32/1-2000
12500
754
EH 32/1-1500
7000
393
EH 32/1-2000
8000
446
HP 64/1-1000
14700
768
HP 64/1-2000
17300
786
EH 64/1-1000
10000
454
HP 64/1-1500
16500
731
EH 64/1-2000
12800
542
EH 64/1-1500
12000
404
Lifetime
Year
20
20
20
20
20
20
20
20
20
20
20
20
20
20
20
20
PB
Year
6.4
8.8
10
12
13.3
15
15.3
16.6
17.8
17.9
19.1
22
22
22.6
23.6
29.7
PO
Year
7.5
10.9
12.9
16.5
19.1
23
23.7
27.3
31.2
31.6
36.1
51.7
51.8
56.3
67.7
99
IRR
%
15
10
8
6
4
3
3
2
1
1
0
0
0
0
0
0
NPV
EUR
2.297
2.814
2.607
519
290
-915
-622
-2.181
-1.622
-1.896
-4.19
-6.453
-3.787
-6.496
-5.383
-6.471
NPVQ˟Max.Investment
EUR
Year
1.15
2.965
12
0.56
5.392
12
0.37
6.629
12
0.14
2.842
12
0.03
6.893
12
-0.09
6.752
12
-0.10
3.711
12
-0.17
7.12
12
-0.23
3.711
12
-0.24
4.212
12
-0.29
7.252
12
-0.38
7.422
12
-0.38
4.287
12
-0.39
6.903
12
-0.42
5.118
12
-0.54
3.815
12
According the presented results several conclusions can be drawn. As most profitable measure according the LCC
analysis is the measure with the highest NPVQ. Thus, most profitable would be installing pellet boiler as unique
heating source with payback period of 4 years. Next profitable measure would be solar system with total collector area
of 16 m2 and heat pump as additional auxiliary heat source. Further follows the systems with collector array areas of
32 m2 and 64 m2 with different storage tank volumes in combination with auxiliary heat device electric heater (EH) or
heat pump (HP). LCC analysis reveal that for same collector area and storage tank volume more profitable is the
combination solar collectors with heat pump instead with electric heater although the starting investment is far lower
for electric heater but energy savings contribute to have lower payback period for the heat pump system.
It can be noticed from Table 12 that the net present values for all of the system combination with 64 m2 collectors,
have negative values which indicates that those systems are not profitable for buildings with specific heat energy
consumption of 70 kWh / m2 a.
11
Next is LCCA for the same solar thermal systems but applied to building with lower specific heat consumption of 57
kWh/m2 a, represented with the Building type III described in Table 4.
It’s interesting to be noticed that same solar thermal systems for the building with the higher specific energy
consumption have lower payback period since they have bigger energy consumption thus generating larger energy
differences compared with the scenario with electrical boiler and the solar system. This is comparison between same
systems but also should be noticed that with same systems at the building with lower specific heat consumption are
achieved bigger solar fractions.
Further is analyzed the building energy consumption influence on the solar system performance i.e. same solar airconditioning system is applied to building with lower specific heat energy consumption. Analysing the presented
results is indicative that improved and feasible combination would be with heat pump instead of electrical heater which
is further verified with LCC analysis. The LCC method provides a better assessment of the long-term cost-effectiveness
of the project/measure than other economic methods which only focus on initial investment costs and operation costs
for the first years. Another comparison is performed for the same solar thermal systems in regard of the primary energy
according the results presented in Table 10 and Table 11. Differences in primary energy consumptions are noticeable
for the same solar systems applied in the two buildings i.e. logically much lower primary energy consumption and
lower CO2 emissions for the building with lower specific heat energy but it’s interesting that especially the differences
are expressed at the smaller collector areas with electrical heater as auxiliary device. Also from the presented LCC
analysis can be concluded that for same collector areas it is more cost-effective the auxiliary source to be heat pump
instead of electrical heater although the investment is higher for the HP but on longer period is more feasible.
The specifc absorption chiller costs per kW cooling power are taken from [14]. The costs are from manufacturers based
on Germany and from a survey of the international energy agency. According the presented costs for the analysis is
chosen that the absorption chiller as ne assembly with the cooling tower costs 1500 eur/ kW i.e. total 15 000 eur.
In the next Table are presented results from the comparison between the energy consumption, system costs, primary
energy consumption and CO2 emissions for the chiller conventional system and the solar assisted absorption cooling.
It is assumed that the chiller average cooling coefficient of performance is 3 and regarding the building cooling loads in
the simulations is used the Building type II for which the annual cooling load is 1800 kWh. The DHW in the case of
conventional chiller is heated 100% by electrical heater while in the case of absorption machine the area of solar
collectors are sufficient to provide 100% solar fraction for the DHW i.e. it is not used any additional auxiliary energy.
According the presented results can be concluded that there is no need to be performed LCC analysis since the simple
payback period is bigger than 20 years i.e. bigger than the system lifetime. Main reasons for the non-feasibility of these
solar-assisted air-conditioning systems mainly is the low electricity price of 0.09 eur/kWh but also is the low cooling
energy demand since it is analyzed residential building which usually have small internal gains. Nevertheless according
the results in Table the solar – assisted cooling systems provides big possibilities for decrease in primary energy
consumption and CO2 emissions
Table 14. Solar thermal system energy performance, costs, parameters and CO2 emissions, primary energy
consumption for building with specific cooling energy consumption 12 kWh/m2 a
Conventional chiller cooling average EER 3 - 600 kWh/year
Parameter
Auxiliary energy - cooling system
Auxiliary energy - DHW
Absorption electrcal energy
System electrical energy
consumption (circ.pumps)
Specific electrical energy price
Total energy price
Heat source device/system price
Primary energy consumption
Annual CO2 emissions
Annual average solar fraction - SF
Solar thermal system - Area/Storage tank volume = 1000 l
Unit
kWh
kWh
kWh
Chiller
600
835
kWh
293
eur/kWh
eur
eur
kWh
kg/year
156
6500
5720
570
%
10/1000-EH 14/1000-EH 16/1000-EH
324
246
204
0
461
18/1000-EH 22/1000-EH
174
132
90
0.09
79
22
2896
289
46
2638
263
59
18
15000
2499
249
66
16
12
2400
239
71
2261
225
78
4. Conclusion
Results show that most profitable would be installing pellet boiler as heating source 100% covering the heating needs,
with payback period of 4 years. Second profitable measure would be solar system with total collector area of 16 m2 and
heat pump as additional auxiliary heat source. Further follows the systems with collector array areas of 32 m2 and 64
m2 with different storage tank volumes in combination with auxiliary heat device electric heater (EH) or heat pump
(HP). LCC analysis reveal that for same collector area and storage tank volume more profitable is the combination solar
collectors with heat pump instead with electric heater although the starting investment is far lower for electric heater
but energy savings contribute to have lower payback period for the heat pump system.
12
Solar cooling is not yet feasible solution for the residential sector. Main reasons are the low specific cooling energy
consumption, low electricity price and high investment cost for the absorption chiller. However, solar energy showed
that can be achieved big primary energy savings resulting with more that 50% compared with the conventional energy
sources i.e. electrical energy. Thus can be said that solar energy is environmental friendly also decreasing to a high
extent the greenhouse gasses emissions.
Further recommendation is to be researched the technical and economic feasibility of implementation solar assisted
air-conditioning systems in commercial buildings since they have high internal loads thus offering potential for big
energy savings.
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