POLITECNICO DI MILANO
Department of Energy
MSc Energy Engineering
Hybrid CCHP: vapor-compression chiller
integrated with absorption sub-cooling
Supervisor: Prof. Paolo Silva
Co-Supervisor: Ing. Nicola Fergnani
Graduation thesis of:
Mauro Corti Student ID 877714
Academic year 2017/2018
III
EXTENDED ABSTRACT
Introduction
The 2030 climate and energy framework sets three main objectives to be achieved
within the indicated year: a reduction of at least 40% of greenhouse gas emission
(compared to 1990 levels), an improvement of at least 27% in energy efficiency and a
share of at least 27% of renewable energy [2]. Energy efficiency is one of the central
objectives for 2030 as well as a key factor in achieving our long-term energy and
climate goals. In this context, Combined Heat and Power (CHP) and Combined
Cooling Heat and Power (CCHP) systems could represent an effective solution to
reduce costs and emissions. As concerned the Italian territory, if those technologies
are recognized as “Cogenerazione ad Alto Rendimento” (CAR), benefits are
guaranteed by the Italian regulatory framework. To meet the ambitious goal imposed
by energy policies is not only important to raise the awareness on new technologies,
but also to study ways to revamp or enhance older system previously installed. For this
reason, the aim of this work is to provide a study about the coupling of a pre-existent
vapor-compression chiller with CCHP technology, in order to create a hybrid CCHP
refrigeration system. In particular, the objective is to figure out if this technology
allows to improve the energy-efficiency of vapour-compression chillers for industrial
refrigeration, and if so, also considering technical feasibility and economic
profitability on case studies. To achieve these tasks, software with increasing levels of
features and complexity have been developed, starting from a model to represent the
vapor-compression section, up to a complete model on the overall hybrid CCHP
system.
IV
Hybrid CCHP refrigeration components
In order to figure out the technical feasibility and economic profitability of the new
proposed solution, after an analysis of the various technologies available, it was
necessary to study the single subsystems used in the novel hybrid CCHP. The
technologies chosen, and introduced in the simulation model are presented in the
following list:
- Vapor-compressor refrigerator
It is the usual technology present in refrigeration cells, and is the component,
to which, the CCHP has to be coupled. To figure out the magnitude of the heat
that can be extracted through vapor-absorption refrigeration, this is the first
component that has been modeled. Its model has been validated through the
comparison with outputs of a design software provided by Bitzer [4]. As
depicted in Figure I, varying the refrigerant and condensation temperature, the
magnitude of the sub-cooling, with respect to the nominal refrigeration power,
assumes different importance. In particular, if R507A or R404A are employed
with usual condensation temperature, among 30-40°C, the sub-cooling heat
assumes an interesting share, around 15-25% of the refrigeration capacity.
V
Figure I: Heat exchanged for sub-cooling (from condensing temperature to 10°C)
with respect to refrigeration power, as function of Tcond
- Prime mover
Among the different prime mover technologies analysed that could be
employed in the present application, Internal Combustion Engine (ICE) has
been selected. Actually, the real choice was among the two most popular
technologies in this field, MGT and ICE. The reasons of the final decision are
the ICE’s lower specific cost, as can be seen in Figure II, the coupling with a
Br-Li absorption chiller and, finally, the high flexibility of the load. Concerning
the match with the absorption machine, since H2O-LiBr pair allows also low
temperature heat input, employing an ICE, it is also possible to use cooling
fluids, rather than more valuable flue gases, that eventually, could be exploited
for other operations in the plant.
0,00
0,10
0,20
0,30
0,40
0,50
0,60
10 20 30 40 50 60 70
Qsc
/Qre
frig
Tcond[°C]
R404A
NH3
R407F
R134A
R407C
R22
R507A
VI
Figure II: Specific cost comparison between ICE and MTG technologies for small size CHP [21]
Analysing technical datasheets and trendlines about engine’s characteristics
provided by Casati and Chinetti [21], it was possible to study the operation
even at partial load condition and, in the end, have a complete characterisation,
in the terms needed for the simulation, of the prime mover.
- Vapor-absorption refrigerator
In the hybrid CCHP configuration, this device is fed by the heat recovered from
the prime mover and its effect is to cool down the fluid exiting the condenser
of the vapor-compression section, to provide a margin for additional
refrigeration capacity or, as in this case, to reduce the electrical energy input.
The choice about absorption chiller regards the working fluid pair, in particular
among the most common working fluids, NH3-H2O and H2O-LiBr. Lithium
Bromide absorption chiller is the type introduced in the model because, with
respect to NH3-H2O one, has better performances (COP in the range 0.7-0.8),
higher reliability (since ice formation is not possible in the circuit), it can
exploit also low temperature heat input and has lower specific cost, as depicted
VII
in Figure III; moreover referring to the application presented in the case study,
extra low temperatures at evaporator are not needed.
Figure III: Specific cost with respect to cooling capacity for NH3-H2O and H2O-LiBr pairs [21]
The above-mentioned components were included in the model which has been
developed. In the model it is possible to design the CCHP system and, so, ICE and
absorption chiller in different ways since different control strategies of the system are
possible. Those are respectively driven by electrical power consumed by compressor,
thermal power required by absorption chiller or maximization of cogenerator load. For
the purpose of analysing real cases, a manual design logic has been also introduced.
Hybrid CCHP simulation model
The idea is to install a cogeneration unit and an absorption chiller that will enhance
both performance and capacity of the vapor-compression refrigerator. The internal
combustion engine will generate electric energy that will be used by the vapor-
€-
€200,00
€400,00
€600,00
€800,00
€1.000,00
€1.200,00
€1.400,00
€1.600,00
- '200 '400 '600 '800 1'000 1'200 1'400 1'600
Spe
cifi
c co
st [
€/k
W]
Pcooling[kW]
LiBr (Single effect)
Ammonia-Water
VIII
compression refrigeration system and will produce, at the same time, useful heat that
will be exploited by an absorption refrigerator. The useful effect of absorption chiller
installed is used to sub-cool the refrigerant at condenser’s outlet of the vapor-
compression section providing margin, maintaining the same refrigeration power, to
reduce the electric energy consumed. Following, in Figure IV, is shown a simplified
block scheme of hybrid Combined Cooling, Heat and Power (CCHP) that it has been
studied.
Figure IV: Hybrid CCHP conceptual scheme
The developed software relies on the vapor-compression model validated through the
comparison with Bitzer’s software [5]. The vapor-compression model has been
developed on Microsoft Excel and Visual Basic, where, though the add-in RefProp
[6], it was possible to calculate properties of refrigerant point by point choosing fluid,
IX
evaporating and condensing temperature and cooling capacity. Through the
comparison of outlet results, the unknown parameter, such as efficiencies of machines,
have been calibrated. Then in the vapor-compression model has been introduced the
effect of the absorption chiller, the sub-cooling up to 10°C or at higher temperature
depending on the maximum capacity of the absorption chiller. Using this model, maps
have been constructed varying refrigeration load and condensation temperature. A
graphical example of these maps is provided in Figure V.
Figure V: Interpolation map of compressor’s electrical power
Relatively to the sizing of the ICE, it has been decided to evaluate different options.
The 4 approaches are following reported:
- 1 (For COMP): ICE sized to meet the rated electric power of the compressor,
corresponding to the annual average design conditions.
- 2 (For MA): ICE designed to meet the rated heat demand of the absorption
machine. This latter is in turn designed to provide a full sub-cooling of the
compression chiller’s refrigerant in annual average conditions.
- 3 (For % of loading): ICE size chosen as the highest among the 2 previous
approaches.
0
200
400
600
800
1000
1200
1400
10 20 30 40 50 60 70
Pe
l,co
mp
[kW
]
Tcond[°C]
100%
80%
60%
40%
20%
X
- 4 (Manual): it allows to select manually the nominal CHP’s electrical power
to find an optimal size for the technical-economic analysis.
In the annual simulation, it has been decided to develop the energy balance and
economic analysis on the base of two different configurations:
- continuous regulation: in which the vapor-compression chiller is able to follow
continuously the load thanks to the presence of an inverter;
- step regulation: the capacity of the refrigerator can be varied on a fixed number
of steps. To follow in a better way the load a thermal storage is adopted.
Once defined all the parameters, as well as system control logic and vapor-
compression’s regulation modality the software starts computing electric and thermal
energy balances, together with economic results.
Case studies analysis
The developed software has been used to test the efficacy of the hybrid CCHP on real
case studies. Two case studies have been analysed.
Case I
The first case study was provided by Impar Impianti s.r.l. [32], a firm that deals with
construction of refrigeration systems for the food industry in northern Italy. The data
provided contain measurement of quantities relative to the refrigeration systems, from
which has been possible to extrapolate the cooling load, condensing and evaporating
temperatures. To make the simulation working faster and to reduce the large amount
of data, it has been decided to use average values on pre-fixed temporal steps. The
annual trend of the cooling load requested by the user is depicted in Figure VI.
XI
Figure VI: Annual trend of cooling load requested by the user, case I
The software receives as an input: annual data, nominal power of the vapor-
compression system, control logics and type of refrigerant. To have a general overview
on the best solution, the approach 4, so the manual selection of CHP’s power, has been
used to perform a size analysis. The results of this analysis show that, only for very
small CHP, and so low CAPEX, the investment result convenient, but not very
interesting since, in the maximum NPV configuration, the results are the one reported
in Table I.
Results
Simple PBP [Years] 9.4
Discounted PBP [Years] 11.0
NPV € 13'000
IRR 7.1%
Thermal energy dissipated [kWh] 20'500
Equivalent hours vapor-compression machine 2'005
Equivalent hours vapor-absorption machine 5'003
Equivalent hours cogenerator 5'517
Table I: Case I maximum NPV configuration results
XII
Case II
The second case study has a cooling demand of higher magnitude and is more
continuous throughout the year, the trend is depicted in Figure VII.
Figure VII: Annual trend of cooling load requested by the user, case II
In case II, the load simulated is more interesting from the CCHP point of view; in fact,
although the investment is much higher, the costs are quickly recovered, generating
interesting NPV while good IRR are achieved. The study has been performed in the
same way of the first case. The vapor-compression nominal power was 1.8MW and
the fluid was still R507A. Starting from an analysis based on CHP’s size, which results
are shown in Figure VIII, it was possible to identify the CHP’s size that maximizes
NPV. This was chosen in order to be consistent with the first case study. Another
solution could be analysing the case with maximum IRR.
XIII
Figure VIII: NPV and IRR of hybrid CCHP with case II data (continuous lines and dots
respectively calculated with approaches 4 and 3)
The dots in the diagram represent NPV and IRR calculated using approach 3, based
on the average annual power requested of 600kWref. Approach 4, represented by points
on continuous lines, gives better results for the same CHP’s size, showing higher NPV
and IRR. The case of maximum NPV employs a cogenerator of electrical power output
170kWel and an absorption chiller of refrigerating nominal power of 165kWref, whose
investment costs are reported in Table II.
Table II: Investment cost breakdown, maximum NPV configuration case II
The software produces balance sheets as outputs. An extract of those results is
presented in Table III, from which it is possible to evaluate the annual saving, that
reaches about 100’000€, 38% of the base case’s annual costs.
0%
5%
10%
15%
20%
25%
30%
€ '0
€ 50'000
€ 100'000
€ 150'000
€ 200'000
€ 250'000
€ 300'000
€ 350'000
€ 400'000
€ 450'000
0 50 100 150 200 250 300
IRR
[%]
NP
V[€
]
Pel,CHP[kW]
NPV IRR
Investment Size [kW] Cost Specific cost [€/kW]
Cogenerator 170 € 190'000 1138
Absorption machine 165 € 64'000 388
BOP (30%)
€ 76'000
CAPEX
€ 330'000
XIV
Cash flows
Energy saving € 101'642
TEE income € 17'252
O&M costs € 25'000
Total CF € 93'894
Table III: Cash flow breakdown, maximum NPV configuration case II
As can be seen in Table IV, where economic and technical results are presented, the
configuration with a 170kWel cogenerator is not the case that optimize the IRR and the
PBP. Nevertheless, the values of these indexes are interesting, since, having the most
profitable size, the IRR shows a robust project and the discounted PBP is a third of the
period of economic analysis (15 years).
Results
Simple PBP [Years] 4.3
Discounted PBP [Years] 4.6
NPV € 390'000
IRR 21.5%
Thermal energy dissipated [kWh] 313'469
Equivalent hours vapor-compression machine 2'560
Equivalent hours vapor-absorption machine 5'066
Equivalent hours cogenerator 5'859
Table IV: Case II maximum NPV configuration results
Conclusions
The aim of the present work was to investigate a new configuration of hybrid CCHP,
in order to estimate investment profitability. The study has been performed through
the development of a model that simulates the vapor-compression system and the
whole hybrid CCHP plant. In order to perform techno-economic feasibility analysis,
two case studies were considered.
XV
Introducing the hybrid CCHP, in the case I, it is possible to save each year about
12’000€. The investment generates an NPV of 13’000€, IRR of 7,1% and 11 years of
PBP. The cogenerator has annual PES of 21%
In case II instead, the hybrid configuration bring consistent benefits. The energy saving
leads to a reduction of operational costs that accounts for 101’642€. The CHP has an
annual PES of 19%, while the project generates an NPV of 390’000€, IRR of 21,5%
and less than 5 years of PBP.
Future development of the present work could comprehend, first of all, tests on
additional case studies, with different load trends and refrigerant fluids, in order to
verify the effective economic validity of the system presented. In particular, additional
simulations could regard real cases in which heat, cooling and electricity are needed
for other processes.
XVI
ABSTRACT
Nowadays, more and more challenging actions of energy saving and emission
reduction are necessary to follow the agreements’ directive imposed by legislative
authorities. CHP and CCHP technologies could represent effective solutions to face
up with the long-term global challenges, as energy security and climate change.
CHP technology allows to produce at the same time electric energy and heat. If a
comparison of generation efficiency with respect to separated production is made, the
result is a lower fuel consumption, meaning economic saving, but also emission
reduction.
The goal of the present work was to design an instrument to model a hybrid CCHP
refrigeration system based on the sub-cooling of the condenser’s exit of the vapor-
compression section, operated by a lithium bromide absorption chiller driven by heat
rejected from the ICE’s operation. The aforementioned configuration was proposed in
order to enhance refrigeration’s efficiency through cogeneration, finding a useful
employment of heat rejected by CHP, achieving high PES and first principle yield.
System’s component modelling started from vapor-compression refrigerator’s model
validation, through the comparison with an already existent web-application owned by
Bitzer. Then the software has been developed introducing different design modalities
of the machines installed, ability to find the optimal management of the system, energy
balances and profitability estimation. Finally, the developed software was applied to
real case studies, to evaluate the real potentialities of this new configuration.
XVII
SINTESI
Oggigiorno, azioni sempre più impegnative di risparmio energetico e riduzione delle
emissioni sono necessarie per stare al passo con le direttive imposte dalle autorità
legislative. Cogenerazione e trigenerazione potrebbero rappresentare soluzioni efficaci
per far fronte alle sfide globali a lungo termine, come la sicurezza energetica e il
cambiamento climatico.
La cogenerazione permette di produrre allo stesso tempo energia elettrica e calore.
Confrontando l’efficienza di generazione rispetto alla produzione separata, ne risulta
una riduzione del combustibile utilizzato, quindi riduzione dei costi ed emissioni.
L’obbiettivo del presente lavoro di tesi era sviluppare uno strumento in grado di
simulare una refrigerazione ibrida tramite CCHP basata sul sotto-raffreddamento
dell’uscita del condensatore della sezione a compressione di vapore operato da un
assorbitore a bromuro di litio azionato dal calore di scarto del CHP. La suddetta
configurazione è stata proposta al fine di efficientare il ciclo frigorifero attraverso la
cogenerazione, trovando un impiego utile al calore di scarto del CHP, ottenendo elevati
PES ed efficienza di primo principio.
In primis, è stato creato un modello del frigorifero a compressione di vapore che è stato
validato tramite il confronto con un’applicazione web. Dopodiché il software è stato
sviluppato introducendo differenti modalità di dimensionamento delle macchine
installate, la possibilità di gestire in maniera ottimizzata il sistema, bilanci energetici e
una stima di fattibilità dell’investimento. Infine, il software sviluppato è stato utilizzato
per analizzare le potenzialità della nuova configurazione applicata a casi reali.
XVIII
CONTENTS
Extended abstract .......................................................................................................... III
Abstract ...................................................................................................................... XVI
Sintesi ....................................................................................................................... XVII
Contents ................................................................................................................... XVIII
Introduction ................................................................................................................... 1
1. Refrigeration cycles ................................................................................................. 3
1.1. Introduction on refrigeration ................................................................................... 3
1.1.1. Vapor-compression refrigeration ................................................................. 4
1.1.2. Vapor-absorption refrigeration ................................................................... 4
1.1.3. Gas cycle refrigeration ................................................................................. 5
1.1.4. Stirling cycle refrigeration ............................................................................ 6
1.2. Vapor-compression .................................................................................................. 8
1.2.1. The cycle ....................................................................................................... 8
1.3. Refrigerator’s compressors ...................................................................................... 9
1.3.1. Reciprocating compressor .......................................................................... 10
1.3.2. Rotary compressor ..................................................................................... 13
1.3.3. Screw compressor ...................................................................................... 16
1.3.4. Centrifugal compressor .............................................................................. 17
XIX
1.4. Regulation criteria of refrigeration units ............................................................... 18
1.4.1. Regulation of condensation and evaporation temperature ...................... 18
1.4.2. Modulation of refrigerant mass flow rate ................................................. 21
1.4.3. Regulation of condenser’s fans .................................................................. 23
1.4.4. Optimised management of a refrigeration cycle ....................................... 25
1.5. Vapor-compression model validation .................................................................... 28
2. CHP & CCHP........................................................................................................... 34
2.1. Introduction on cogeneration & trigeneration ...................................................... 34
2.1.1. Cogeneration.............................................................................................. 34
2.1.2. Trigeneration ............................................................................................. 36
2.2. CHP’s prime mover ................................................................................................. 37
2.2.1. Internal Combustion Engines (ICE) ............................................................. 37
2.2.2. Micro-Gas Turbines (MGT) ......................................................................... 38
2.2.3. Other emerging technologies .................................................................... 40
2.2.4. Prime mover performances ........................................................................ 42
2.3. Absorption Chiller .................................................................................................. 47
2.3.1. Single-effect cycle ...................................................................................... 48
2.3.2. Working fluids pairs ................................................................................... 51
2.4. Typical applications ................................................................................................ 54
2.5. Incentives ............................................................................................................... 55
2.5.1. “Cogenerazione ad Alto Rendimento’’ CAR ............................................... 56
2.5.2. GSE’s role ................................................................................................... 59
2.5.3. “Certificati Bianchi’’ CB type II-CAR ........................................................... 60
2.5.4. “Certificati Bianchi’’ CB or “Titoli di Efficienza Energetica’’ TEE ................ 61
3. Hybrid CCHP .......................................................................................................... 63
3.1. Literature review .................................................................................................... 63
XX
3.1.1. Flexible chiller with hybrid absorption/compression cycle ........................ 63
3.1.2. Hybrid refrigeration system recovering condensation heat ...................... 65
3.1.3. Solar absorption-subcooled compression hybrid cooling system .............. 69
3.2. New concept for hybrid CCHP ................................................................................ 72
3.2.1. Arrangement .............................................................................................. 72
3.2.2. Performances of working fluids varying condensation temperature ........ 74
4. Case studies simulation ......................................................................................... 84
4.1. Introduction to real data simulation ...................................................................... 84
4.1.1. Data management ..................................................................................... 84
4.1.2. Yearly simulation model............................................................................. 88
4.2. Results of Impar’s data simulation ......................................................................... 91
4.2.1. Size analysis ............................................................................................... 91
4.2.2. Maximum NPV case ................................................................................... 95
4.3. Results of case II data simulation ......................................................................... 100
4.3.1. Size analysis ............................................................................................. 100
4.3.2. Maximum NPV case ................................................................................. 103
5. Conclusions ......................................................................................................... 109
List of figures .............................................................................................................. 112
List of tables ............................................................................................................... 116
Abbreviations ............................................................................................................. 118
Bibliography ............................................................................................................... 119
1
INTRODUCTION
The 2030, climate and energy, framework sets three main objectives to be achieved
within the indicated year: a reduction of at least 40% of greenhouse gas emission
(compared to 1990 levels), an improvement of at least 27% in energy efficiency and a
share of at least 27% of renewable energy [2]. Energy efficiency is one of the central
objectives for 2030 as well as a key factor in achieving our long-term energy and
climate goals. In this context, Combined Heat and Power (CHP) and Combined
Cooling Heat and Power (CCHP) systems could represent an effective solution to
reduce costs and emissions. Producing at the same time electrical and thermal energy
is possible to save fuel with respect to separate production, providing economic as well
as environmental benefits.
Although CHP and CCHP represent an interesting technology those systems are not
suitable for all applications: users that require quite constant and contemporaneous
electrical and thermal load are preferred. For this reason, feasibility studies are
necessary to guarantee technology’s effectiveness and economic profitability. The aim
of the present work was to provide a study about a new cooling system: hybrid CCHP
refrigeration; evaluating its potentiality and providing a dedicated model to design the
CCHP’s components and to match them to vapor-compression refrigerator, assessing
technical feasibility and economical profitability of the investment.
The first chapter provides an overview of most important refrigeration techniques
with a particular focus on vapor-compression refrigeration. After a brief analysis about
the compressor solutions that can be applied to commercial refrigerators, a focus on
the regulation criteria of those cycle is provided. In addition, to close the chapter the
2
development of a vapor-compression model, validated through comparison with a pre-
existent software is presented.
The second chapter regards CHP/CCHP technologies; it starts with a general
overview on cogeneration and trigeneration, followed by an analysis on the prime
mover technologies that can currently be applied to those systems, a particular focus
is given to Internal Combustion Engines and Micro-Gas Turbine. In addition, ICE
performances, especially at partial load, are presented. After the brief introduction in
the first chapter, the absorption refrigeration is resumed and deepened. Once the
system’s components have been identified, typical applications are presented, in
which, cogeneration and trigeneration technologies could be particularly interesting.
The chapter closes with CHP/CCHP’s support policies guaranteed by the current
Italian legislative framework.
In the third chapter will be introduced the innovative hybrid CCHP refrigeration
system. After a short literature review about similar applications’ study, the chapter
describes the arrangement of the new refrigeration system and the effects on the vapor-
compression section. Then a description on how the first version of the model was
built is provided. The model is developed to give as outputs energy balances,
performances and simplified saving data. An analysis performed with this software
closes the chapter; the results are presented to demonstrate benefits and technical
feasibility of the configuration proposed.
The fourth chapter contains a description of how the model used to simulate real loads
was built starting from the previous version and how the real data were managed to
perform the simulation. The final software provides a profitability analysis having as
input the annual cooling demand. There’s the possibility to regulate the vapor-
compression chiller by discrete steps, but also in a continuous way. In addition to the
three machines’ design modalities of model’s first version a forth manual logic to
perform size analysis was introduced. The output of the software are annual balance
sheets, charts and detailed economic analysis. In the last part of this work two annual
load are tested and through size analysis for each case is presented in a more deepened
way the cogenerator size that maximize the NPV of the relative investment.
3
1. REFRIGERATION CYCLES
In this chapter will be first illustrated a general overview of the existent types of
refrigerator, then will be deepened the vapor-compression cycle, because this thesis
proposes a new implementation with this refrigeration system. After the revision about
vapor-compression refrigeration, will be presented a sub-chapter on the types of
compressor that it is possible to use in this cycle, followed by a sub-chapter about cycle
regulation. Finally, will be explained the way in which the model validation through
comparison with Bitzer’s data was carried.
1.1. Introduction on refrigeration
A refrigeration cycle is a thermodynamic cycle that can transfer heat from an
environment at low temperature to another at higher temperature. The device that does
a refrigeration cycle can be interpreted and used in two ways:
- refrigeration machine;
- heat pump whose purpose is to provide heat to a warm environment, taking it
from a colder one.
According to the second law of thermodynamics heat cannot spontaneously flow from
a colder location to a hotter area; work is required to achieve this transformation. In
both the usages, work is needed to make the cycle work and so to deliver heat to the
hottest point at temperature T2 (useful effect of the machine seen as heat pump) and to
absorb heat from the coldest point at temperature T1 (useful effect of the machine
working as refrigerator). Formally the easiest method to design a refrigerator is to use
a reverse Carnot cycle whose efficiency is function only of the temperatures of the two
sources (1.1).
4
(1.1) 𝐶𝑂𝑃𝑖𝑑 =𝑇1
𝑇2−𝑇1
Being an ideal cycle, is not possible to replicate it. Real heat pump and refrigeration
cycles can be classified as vapor-compression, vapor-absorption, gas cycle and Stirling
cycle types. To determine the efficiency of those real refrigerators, usually a parameter
called COP (Coefficient of Performance) is used, that it’s exploited also for heat
pumps; so, in order to avoid confusion, sometimes it is written as EER (Energy
Efficiency Ratio) with the same meaning. This parameter is based on the classic
definition of efficiency: useful effect compared to the work done to obtain it. Speaking
about refrigeration, the useful effect is the heat that is extracted through evaporation,
while the work done it’s the mechanical power to make the compressor work. The
equation (1.2) shows this relationship.
(1.2) 𝐶𝑂𝑃 =�̇�𝑒𝑥𝑡𝑟𝑎𝑐𝑡𝑒𝑑
�̇�𝑐𝑜𝑚𝑝𝑟𝑒𝑠𝑠𝑜𝑟
1.1.1. Vapor-compression refrigeration
The vapor-compression refrigeration is a thermodynamic cycle in which compression
work is done on working fluid to achieve a condensation pressure, and so temperature,
that allows the heat exchange with the external environment; so, after compression,
heat is extracted from the cycle to promote the change of state, and, after that, the
refrigerant is expanded and finally evaporated subtracting heat in the opposite change
of state. The working fluid during evaporation gives the desired refrigeration effect.
Further details on this cycle are provided in paragraph 1.2.
1.1.2. Vapor-absorption refrigeration
In the early years of the twentieth century, the ammonia-water vapor-absorption cycle
was popular and widely used but, after the development of the vapor-compression
cycle, it lost lot of importance because of its low coefficient of performance (about
one fifth of that of the vapor-compression cycle). Nowadays, the vapor-absorption
5
cycle is used only where heat is more readily available than electricity, such as waste
heat provided by solar collectors, off-the-grid refrigeration in recreational vehicles or
application like cogeneration and waste heat recovery from industrial processes.
Absorption and compression cycles have different operating principles; the latter is
based on compression and expansion, the other one, instead, on variation of partial
pressure without significant variation of the absolute pressure. In the absorption
system, the compressor is replaced by an absorber, which dissolves the refrigerant in
a suitable liquid, a liquid pump, which raises the pressure and a generator which, on
heat addition, drives off the refrigerant vapor from the high-pressure liquid. Some
work is required by the liquid pump but, for a given quantity of refrigerant, it is much
smaller than the work needed by the compressor in the vapor-compression cycle. In an
absorption refrigerator, a suitable combination of refrigerant and absorbent is used.
The most common combinations are ammonia-water (NH3-H2O) and water-lithium
bromide (H2O-LiBr). After this introduction, this cycle will be widely discussed in
paragraph 2.3.
1.1.3. Gas cycle refrigeration
When the working fluid compressed and expanded is a gas, but does not change phase,
the refrigeration cycle is called a gas cycle. As there is no condensation and
evaporation in a gas cycle, components corresponding to the condenser and evaporator
in a vapor-compression cycle are the hot and cold gas-to-gas heat exchangers. For
given extreme temperatures, a gas cycle may be less efficient than a vapor-
compression cycle because the it works on the reverse Brayton cycle instead of the
reverse Rankine cycle; as such, the working fluid never receives or rejects heat at
constant temperature as is possible to see from the T-s diagram in Figure. 1.1.
6
Figure. 1.1: Technical scheme and T-s diagram of gas cycle refrigeration
In the gas cycle, the refrigeration effect is equal to the product of gas’ specific heat
and the temperature rise in the cold heat exchanger. Therefore, for the same cooling
load, gas cycle refrigeration machines require a larger mass flow rate, which in turn
increases their size. Because of their lower efficiency and larger bulk, gas cycle coolers
are not often applied in terrestrial refrigeration. The air cycle machine is very common,
however, on gas turbine-powered jet airliners since compressed air is readily available
from the engines' compression sections. These jet aircraft's cooling and ventilation
units also serve the purpose of heating and pressurizing the aircraft cabin.
1.1.4. Stirling cycle refrigeration
The ideal Stirling cycle is a thermodynamic cycle composed by two iso-thermal and
two constant volume transformations. The first two take place in the heat exchange
sections, in which the working fluid receives or gives heat to the external environment,
together with a mechanical action of expansion or compression, that allows to maintain
the temperature constant. The constant volume transformations are made through a
constant volume transfer of the fluid contained in the motor, from the hot side to the
cold side and the other way around, with a consequent reduction or increase of the
7
average pressure of the gas. The ideal transformations are reported in Figure. 1.2,
respectively on P-V and T-S diagrams.
Figure. 1.2: Thermodynamic diagrams of ideal Stirling cycle: P-V diagram and T-s diagram [3]
The efficiency of ideal Stirling cycle coincides with the Carnot cycle one (1.3).
(1.3) 𝜂𝑖𝑑 = 1 −𝑇𝑚𝑖𝑛
𝑇𝑚𝑎𝑥
The Stirling refrigeration machine is a closed cycle actuated mechanically, that
establish a cyclic pulsation of a fluid confined inside the machine. The cycle fluid
thickens and expand inside the machine, delivering heat constantly in a point of the
device, and subtracting it in another point. The heat delivered and subtracted are
transferred out of the machine and used for the desired purposes. The Stirling cycle
refrigeration, although is less known for the current use, has an important
characteristic: is not linked to any change of state of cycle fluid. Instead other cycles
are linked, so limited, to precise ranges of operating temperatures given by the fluid
used. In Stirling cycles the type of fluid is not particularly binding (if it is gaseous at
operating temperatures). In the practice, the Stirling cycle is characterised by several
irreversibilities, that determine a negative deviation from Carnot’s efficiency.
Analysing the nominal characteristics of the Stirling motors present in literature, as
8
well as data from commercial models it is found a second principle efficiency between
0.3-0.7 [3].
1.2. Vapor-compression
1.2.1. The cycle
The principle is to use a refrigerant that undergoes phase changes. It is one of the many
refrigeration cycles and it is the most widely used method for air-conditioning of
buildings and automobiles. It is also used in domestic and commercial refrigerators,
large-scale warehouses for chilled or frozen storage of foods and meats, refrigerated
trucks, and a host of other commercial and industrial services. Oil refineries,
petrochemical and chemical processing plants and natural gas processing plants are
among the many types of industrial plants that often utilize large vapor-compression
refrigeration systems.
The vapor-compression uses a circulating refrigerant as the medium which absorbs
and removes heat from the space to be cooled and, subsequently, rejects it elsewhere.
Figure. 1.3 depicts a typical, single-stage vapor-compression system with its T-s
diagram.
Figure. 1.3: Single-stage vapor-compression system scheme and its T-s diagram
1
23
4
Tem
pe
ratu
re
Entropy
9
All such systems have four components: a compressor, a condenser, a throttle valve
and an evaporator. Circulating refrigerant enters the compressor as a saturated vapor
and is compressed to a higher-pressure level, with a consequent increase in
temperature. The superheated vapor is at a temperature and pressure such that it can
be condensed with either cooling water or cooling air flowing across the coil or tubes.
This is where the circulating refrigerant rejects heat, which is carried away by either
the water or the air.
The condensed liquid refrigerant is next routed through an expansion valve, where it
undergoes an abrupt reduction in pressure. This pressure reduction results in the
adiabatic flash evaporation of a part of the liquid refrigerant. The auto-refrigeration
effect of the adiabatic flash evaporation lowers the temperature of the liquid and vapor
refrigerant mixture, so as to make it colder than the temperature of the enclosed space
to be refrigerated. The cold mixture, then, goes through the coil or tubes in the
evaporator, while a fan circulates the warm air in the enclosed space across the coil or
tubes carrying the cold refrigerant liquid and vapor mixture, so that the warm air
evaporates the liquid part of the cold refrigerant mixture. At the same time, the
circulating air is cooled and thus lowers the temperature of the enclosed space to the
desired temperature. The evaporator is the place where the circulating refrigerant
absorbs and removes heat, which is subsequently rejected in the condenser and
transferred elsewhere by the water or air. To complete the refrigeration cycle, the
refrigerant vapor from the evaporator is routed back into the compressor.
The thermodynamic of the vapor compression cycle can be analysed on a temperature
versus entropy diagram, as depicted in Figure. 1.3, where point 1 represent the outlet
of the evaporator, point 2 the outlet of compressor, point 3 the outlet of the condenser
and, finally, the point 4 is after the throttle valve.
1.3. Refrigerator’s compressors
The classification of compressors split mainly the machines in two categories:
volumetric machines and centrifugal ones. In the volumetric devices the refrigerant
vapor is compressed in a closed environment through a volume reduction; in the
10
centrifugal machines, instead, the refrigerant vapor is transferred through a strong
acceleration, which results in an increase of pressure.
The volumetric compressor offers the following characteristic:
- high compression ratio;
- independence from certain intervals of inlet section conditions;
- continuity not required.
The centrifugal machines have the following characteristics:
- limited compression ratio;
- high volumetric flow rate.
In practice the most frequently employed type of machines are:
- reciprocating pistons compressor
- rotary piston compressor
- scroll compressor
- screw compressor
- centrifugal compressor
1.3.1. Reciprocating compressor
In this kind of machine, the overheated refrigerant vapor coming from the evaporator,
is sucked at low temperature and pressure by the compressor, where it is compressed
at higher pressure and temperature. The distinction among open, hermetic and semi-
hermetic compressor is not related with the type of compressor, but only with the
position of the engine; in the case of open compressor (Figure. 1.4), the shaft leans out
from the body through a stuffing box and is connected to an external command, which
can be direct or with transmission belts.
11
Figure. 1.4: Reciprocating open compressor
The open compressor is universal, robust and can be used with almost all types of
refrigerant available. Since the circuit is separated any motor’s fault has not influence
on refrigerant behaviour. Within certain ranges, the rotational speed can be varied
acting on transmission belts, which in turn it varies the cooling capacity of system.
The compression phases of a reciprocating compressor are reported schematically in
Figure. 1.5.
Figure. 1.5: Compression phases of a reciprocating compressor
12
The piston goes down creating a depression in the cylinder that opens the suction
valve, allowing the refrigerant vapor passage. When the piston reaches the bottom dead
center, it starts going back and the pressure in the cylinder raises. At this point, the
suction valve closes. The discharge valve is closed, because the pressure in the
environment above is higher than the one in the cylinder. The next process is
compression: the refrigerant vapor is compressed in the cylinder from suction pressure
to the compression one. When the pressure in the cylinder overcome the condensation
one, the discharge valve opens and the compressed vapor exits the cylinder, going
towards the condenser.
The reciprocating hermetic compressors can be divided in two categories: hermetic
and semi-hermetic. The first type is a device constituted by a sealed casing (or
capsule), made of steel sheets. The command motor and the compressor are located
inside the capsule and the compressor is driven directly by the motor. The principal
difference between hermetic and semi-hermetic compressor is that, in the latter, is
possible to access both to the motor and the compressor disassembling the bolted parts.
An example of those compressors can be seen in Figure. 1.6.
Figure. 1.6: Bitzer’s semi-hermetic compressor [4]
13
1.3.2. Rotary compressor
The distinction between rotary and reciprocating compressor lies in the fact that, in the
first one, the rotation of the piston produces a gradual increment of the suction area
and a gradual decrease of the compression chamber. In this way, it is possible to have
the suction and the compression of the refrigerating fluid. There are three types of
rotary compressors:
- Rotating piston: in a cylindric casing there’s the eccentric rotation of the piston,
during which, it remains in contact with the walls of the cylinder. It follows a
division of the compression chamber between suction and compression sides.
Figure. 1.7: Rotating piston compressor
During a revolution of the piston, there’s a gradual extension of the suction
chamber, that is needed to attract refrigerating fluid. In the same time, the
chamber of compression undergoes a gradual reduction; in this way, the
refrigerant present in the chamber is compressed until it reaches the pressure
that allows the opening of the discharge valve.
14
- Sliding-vane: this compressor, depicted in Figure. 1.8, has a cylindric casing,
inside which the piston rotates in an eccentric way with respect to the cylinder’s
axis. For the separation of the work chamber, on the piston are present mobile
paddles that allow the sealing on the cylinder’s walls thanks to the centrifugal
force. It is necessary to have at least two paddles to divide the work area in
single vanes.
Figure. 1.8: Sliding-vane compressor’s section
- Scroll: this machine, of which an example is shown in Figure. 1.9, is composed
essentially by two identic spirals, inserted one in the other, a command shaft, a
casing and a lateral closing cover.
Figure. 1.9: Bitzer’s ECH209Y semi-hermetic scroll compressor
15
One of the two spirals is fixed, instead the other, named orbiting scroll, is
mobile. The driving force comes from an eccentric shaft that has a transitional
motion on a circular guide with progressive movement. The working principle
is depicted in Figure. 1.10.
Figure. 1.10: Working principle of a scroll compressor
Both the spirals get in contact one with the other. The upper spiral is fixed and
has an opening where the compressed vapor is discharged. The mobile spiral
is connected to the command shaft. The intake vapor coming from the
evaporator enters in the perimeter of the spiral; the compressed one exits from
the opening in the centre of the fixed spiral. It is important to ensure the sealing
contact between the surfaces of the spirals, that must have a high passage
precision, because they have the same role of pistons’ ring in a reciprocating
16
compressor. The circulating movement of a guide allow to the spiral to create
variable spaces for the vapor. The relative movement among the two spirals
move the compression space reducing the volume till the discharge opening, in
the last phase the two spirals detach and the compressed vapor can flow outside
the opening.
1.3.3. Screw compressor
The screw compressors employed in refrigerating systems are in most of the cases of
the double rotor type [1]: the working principle is based on the gas displacement. The
two rotors are complementary among them being male and female, with different
asymmetrical profile and contact with small clearance. The principal rotor (male),
driven by the motor, can have 4 or 5 lobes instead the female rotor has 6; an example
can be seen in Figure. 1.11.
Figure. 1.11: Screw compressor
With the rotation of the rotor the lobe’s thrust moves axially from the inlet side to the
discharge side, in this way the existent cavity among the lobes of the rotors increases
reaching its maximum value. At this point the suction occurs. During rotation the
volume of the cavity reduces. When the compression is ended, it starts the expulsion
17
of the compressed vapor. This work cycle is repeated for every single cavity generating
practically a constant flow. In those compressors is present a special valve named slide
valve that can modulate the capacity. The slide valve changes the point where the
compression process begins along the axis of the screw. Most screw compressors can
continuously modulate capacity between 10% and 100% [12].
1.3.4. Centrifugal compressor
The centrifugal compressor for refrigeration machines (named also turbocompressor
or radial compressor) as opposed to volumetric machines, that generate a pression on
the sucked vapor through the reduction of the space, they work as rotodynamic
devices. The centrifugal machines operate through the transformation of a part of the
kinetic energy in pressure. These devices are used when is needed a high refrigeration
capacity maintaining a small footprint. The centrifugal compressors for air
conditioning usually are built with one or two compression stages, according to type
of refrigerant and operating conditions, but they can have up to 6 stages.
Figure. 1.12: Centrifugal compressor’s impeller
The electric motor’s rotor and both the impellers of the first and second stage, Figure.
1.12, are mounted on the same shaft. The compensation of the axial thrust is given by
the contraposition of the two impellers.
18
1.4. Regulation criteria of refrigeration units
The refrigeration system is a device for the creation and conservation of a determined
thermal state through a temperature lower than the environment one. One of its tasks,
for example, consists in maintaining a regulation quantity, in a refrigeration cell, that
is determined through a desired target. The desired thermal state usually is below
environment temperature, therefore is necessary to preserve it also in case of variation
of the external condition. The refrigeration systems, that work in an optimal way, adapt
themselves to the respective working conditions employing a regulation of the
refrigeration power. If a regulation of the capacity is not possible, inevitably the
working conditions will not be respected, and the electrical energy consumed can
increase because the system works in anti-economic way.
The regulation of refrigeration units concerns mostly three aspects [11]:
- regulation of condensation and evaporation temperature;
- modulation of refrigerant mass flow rate;
- modulation of flow rate of fluids that exchange heat respectively with
condenser and evaporator.
The reference system is a refrigeration cycle air cooled through a finned condenser,
which operates in forced convection conditions, thanks to the presence of one or more
fans. This plant type is diffusely installed on Italian territory, especially in the
industrial, commercial and domestic sectors [11].
1.4.1. Regulation of condensation and evaporation temperature
To make a refrigeration system working in an optimal way, it is necessary to make
sure that, the condensation pressure, and so temperature, be as low as possible. If it is
maintained without any necessity to a higher value, the pressure always induces to an
increase of the power absorbed by the compressor; in some case is possible to have an
overload on the command motor [1]. For this reason, devices of safety and limitations
are exploited. On the other hand, if mechanical valves are employed, a condensation
19
pressure too low could create a situation in which, in the expansion valve the necessary
flow rate of refrigerant is no more able to flow, with consequent underfeeding of the
evaporator, that in the end, induces a reduction of the cooling power given by the
system. The refrigeration power can increase only if the condensation pressure is
lowered to a limit value that still allows the flow of the refrigerant through the
throttling valve.
The condensation and evaporation pressures are directly linked to heat exchangers
temperatures, inside of which is present refrigerant under vapor-liquid equilibria.
Referring to the management of condensation pressure it is possible to divide the
refrigeration machines in two categories [11]:
- Head Pressure Control (HPC) case in which the condensation pressure is
maintained constant at its project value. This regulation technique is actuated
through the activation of the compressor when the condensing pressure
decreases under a predetermined threshold value. Due to start-up of the
compressor, the condensation temperature will exceed the upper trigger level
and, consequently, the condenser’s fans will start working to cool down. This
control has a low complexity and, allow the use of a thermostatic valve, whose
role, is to maintain a constant fluid level in the evaporator or alternatively to
maintain a constant superheating ∆T.
- Condensing Temperature Control (CTC) case in which the condensation
pressure is regulated following the environment temperature, through the
condenser’s fan unit. The fans will work to maintain a temperature equal to the
ambient temperature, plus a ∆T margin. This method is better from an energetic
point of view, but more complex. In this case, the regulation of the compressor
is based on the user’s load, in particular, it is activated when the delivery
temperature of cooling fluid overcomes a prefixed threshold. This method
employs an electronic valve, that regulates its opening as function of
superheating at inlet of the compressor.
20
Practically, not all types of compressors are able to manage, in an efficient way, a
variation of the compression ratio between condensation and evaporation. Following,
are reported the ways, in which, the different compressors react, to a variation of the
compression ratio [11].
- Reciprocating compressor: as the compression ratio decreases, the opening of
the discharge valves in compression phase is anticipated; instead, anticipated
closure of intake valve takes place in the aspiration phase. In the first case, the
intake volumetric flow rate remains the same, in the second case instead, is
reduced.
- Screw compressor: decreasing the compression ratio it follows a progressive
opening of the slide valve, to which, is associated a reduction of the volumetric
flow rate elaborated (linked to the reduction of the intake volume), that can be
compensated in case an inverter is present increasing the rotational speed of
the compressor.
- Centrifugal compressor: as the compression ratio decreases, it is possible to
reduce the rotational speed, through an inverter, if present, or to regulate
distributor blades, if Variable Guide Vane (VGV) technology is present. In
both cases, there is a variation of the intake mass flow rate, that as to be limited
to a range near the operating point, to avoid the onset of stall and surge
phenomenon.
- Scroll and sliding-vane compressors: although these technologies are the most
diffused for small size refrigerators, it is not possible to operate an efficient
regulation of condensing pressure. A reduction of this quantity, with respect to
the project value, gives an expansion at the outlet of the compressor, with a
consequent loss of work; instead, an increase with respect to the project value,
would lead to an over-compression, to which, is associated a reduction of the
compression’s efficiency, because of recirculation phenomenon at the
discharge.
21
1.4.2. Modulation of refrigerant mass flow rate
This regulation is essential, because necessary to guarantee, that, the cooling power
requested by the user is matched with the one available from the refrigeration cycle.
Following, are listed the most important modes of regulation, used by compressors in
cooling cycles [11].
- Inverter: using this device is possible to regulate the rotational speed of the
electric motor, regulating the mass flow rate for different types of compressors.
In general, it is possible to observe that, this regulation is more efficient when
applied on volumetric compressors (pistons, screw, sliding-vane or scroll), and
less efficient with centrifugal compressors, for which, the surge limit gives a
maximum variation of 35% of the mass flow rate treated.
- Compressors staging: through the activation/shutdown of compressor is
possible to obtain an alternated operation, that, elaborates a time-average flow
rate of refrigerant equal to the one requested. The cons are that: it gives
fluctuations on the cycle, in particular on condensation and evaporation
temperature, and increase the mechanical stresses of the components. If are
present groups of compressors in parallel, it is possible a modulation in
sequence, and an on/off manage of the single group.
- Slide valve: the present type of regulation is exclusively for screw compressors.
As descripted in the paragraph above, a reduction of the mass flow rate, given
through slide valve, it is associated to a reduction of the compression ratio, for
this reason is preferable to vary the rotational speed with an inverter.
- Digital regulation: this regulation modality regards scroll compressors, for
which, it is possible to space the spirals, establishing communication among
volumes inside them, avoiding the compression. This solution is preferred,
with respect to an on/off managing of the motor, because it reduces the
mechanical stresses, and produces a more continuous regulation.
For reciprocating compressors, besides using an inverter or staging, also the following
types of capacity regulation could be employed [1]:
22
- Intake throttling regulation doesn’t depend from the operation of the
compressor, but from the one of the overall system, as there’s a regulation of
the evaporation pressure, located in the aspiration line. The difference of
pressure is formed across an inlet “butterfly’’ valve, that increase, when the
system is operating at full-load. As the pressure drop across the valve increases,
the density of the entering fluid decreases. This results in a lower mass flow,
reducing compressor's power requirement. This way to regulate the system
works more in antieconomic way when the compression ratio is low. When the
regulator throttles the flow there’s also an increase of the temperature of
compression; so, also this method has some limits.
- Hot gas bypass is by definition the bypass of hot gas among the discharge and
intake sides of the compressor. If in the evaporator, there is a reduction of the
heat flux, also the evaporation pressure decreases, due to this, the regulator
opens, and lets the vapor passing from the discharge side to the intake. Once
the hot vapor is entered in the aspiration line, it mixes with the intake one; so,
the intake gas became hotter. This transient regulation continues until it is
reached an equilibrium state, in which, a lower supply of refrigerant to the
evaporator, it is compensated by a lower supply of refrigerant to the
compressor. Lower the cooling capacity to be maintained, higher is the quantity
of superheated gas that is deviated. In the same time, the temperature of
aspiration duct increases, with a consequent increment of the compressor’s
outlet temperature. The maximum capacity of the regulator is limited by the
oil’s temperature in the compressor; if the discharge temperature becomes too
high, even the lubricating’s one becomes it, with the possibility of
carbonisation of oil, and in the worst cases damage of discharge valves. In
addition, the type of refrigerant employed is important: in the case of R717
(ammonia) the temperature increases rapidly, instead, with refrigerants R134A
and R404A this doesn’t occur [1]. It has to be highlighted that, this regulation
implies consistent losses, because the compressor absorbs the nominal power
also at partial load.
23
- Intake flow regulation applies the same idea of the bypass, but without
recompressing the vapor, so it allows to reach a higher efficiency. However, is
possible only to have a step adaption of the cooling capacity, because it acts
only on couples of cylinders. In fact, this type of regulation works only on the
cylinders to which is connected. Essentially, it is based on the combined effect
of the working valves (intake and discharge); when the compressor works at
full-load, the solenoid valve closes the command duct on the intake side, and
contemporaneously, it opens the duct on the discharge side. So, the principal
valve remains closed, and the superheated vapor flows through the check valve
towards the condenser. If a couple of cylinders has to be partialized, the
solenoid valve receives an impulse, and it opens the principal valve closing the
check valve. The bypass duct’s pressure decreases until it almost reaches the
aspiration pressure, and the driving piston of the couple of cylinders works only
against the resistance to the flux of the working valves. Even if, the couples of
cylinders that still work after the partialisation dissipate the heat flux coming
from the couple of cylinders that is not giving work, the outlet temperature of
the compressor increases, with the progressive reduction of the capacity,
because the quantity of refrigerant exiting the evaporator is reduced.
- Inverted flux regulation or cylinders’ partialisation exploits a system, in
which, is possible to exclude one or more pistons closing all the valves during
the entire cycle, or, maintaining open only the intake valves, so the refrigerant
breathed by the couple of cylinders is sent back to the aspiration chamber
through controlled bypass ducts. The piston has not to do volume variation
work. The work remained is only due to friction; this regulation therefore
works in a more efficient way.
1.4.3. Regulation of condenser’s fans
Using as a reference air-cooled condensers, the reduction of air volumetric flow rate
towards finned batteries, in conditions of partialisation, can be operated using the
modalities depicted below [11].
24
- On/Off regulation: is the simplest regulation system and consists in the
activation/shutdown of the fan or fans, as function of the temperature reached
by the condenser. When temperature reaches a prefixed trigger value, the fan
is activated until the threshold less a ∆T of hysteresis is reached. In the case in
which, are present more fans in parallel, they are activated/shutdown in
sequence, and then eventually, only one fan is modulated through on/off cycles,
having a more continuous regulation, with consequent enhancement in terms
of stability and efficiency. When the fan is not working, the condenser still
rejects 10% of its nominal capacity due to natural convection effects [12].
- Step-velocity regulation: consists in the regulation of fan’s velocity, or fans if
present, through electronic devices, that allow, to intervene in a discrete way
on the rotational speed of the electric motor associated to the fan.
- Continuous regulation of velocity (inverter): using an inverter is possible to
modulate the feeding frequency of the motor, varying continuously the
rotational speed, in an analogue way, as depicted for the compressors’
regulation. This is the preferable choice of regulation, because it allows a
reduction of the energy consumed, and enhances the stability of the
refrigeration cycle, for which, is possible to maintain a condensation
temperature constant and without oscillation.
The advantage of using the step-velocity regulation or variable frequency drive
(inverter) can be explained by the fact, that, the power to drive a fan, is proportional
to the cube of fan’s speed [12]. If the fan speed can be cut in half, half the air mass
flow is achieved at only one-eighth of the design fan power. Recapitulating, as
highlighted in the sub-section 1.4.1, depending on the management criteria of
condensation temperature, the fan units will work in two different manners. In case of
HPC logic, fans are activated when a pre-set threshold is exceeded, instead with CPC
regulation, fans are activated to maintain a predetermined margin with respect to
environmental temperature.
25
1.4.4. Optimised management of a refrigeration cycle
According to what illustrated in the paragraph above, during the operation of a
refrigeration cycle is possible to highlight two principal parameters linked to
regulation needs [11].
- Cooling load requested by the user;
- external ambient temperature.
Is possible to refer to the first parameter, for the regulation of the mass flow rate of
refrigerant fluid (paragraph 1.4.2), and to the second parameter, for the regulation of
condensing pressure in system where it is allowed (paragraph 1.4.1).
The condenser must exchange a fixed predetermined thermal power, function of the
load and the efficiency of the cycle, increasing the temperature difference among
condenser and ambient, it is possible to slowdown the fan’s speed and the other way
around.
At the same ambient temperature, an increment of the condensation pressure will lead
to an increase of the power absorbed by the compressor, following the higher
compression ratio needed, and on the other side, a reduction of the electrical power
absorbed by the fans. So, it obviously exists an optimum, that corresponds to the trad-
off of the two effects reported above. This can be visualised through the Figure. 1.13,
that shows the trend of the total power absorbed by the refrigeration machine function
of the imposed condensing pressure.
26
Figure. 1.13: Overall power absorbed by the refrigerator group (expressed in Horse Power) function
of condenser’s pressure for different regulation mode of the fans [12]
The different curves represent the different regulation modalities of the fans. From the
graph is possible to see that, the system’s optimum point (lowest electrical power
absorbed) matches to similar condensation pressures for different regulation systems
analysed. The optimum point associated to a system, that employs the Variable-
Frequency Drive (VDF), shows an 8% lower power absorbed, with respect to the
optimum point, of a system that uses an on/off regulation, thanks to the lower
consumption associated to the fans [12]. Moving left, with respect to the optimum
point, there would be a decrease of the ∆T condenser-environment, so the need to
increase the fan’s load, that leads to the approach of different systems’ curves. This
until the minimum pressure is reached, at which correspond the maximum speed of
the fans, and so the exclusion of regulation systems.
It is important to highlight that are present limits for the right operation of the
compressor. These limitations are provided by the constructor as a map, where is
depicted the area, in which, the compressor works in the right conditions. The map is
built as function of condensation and evaporation temperatures, that, are directly
27
linked to heat exchangers’ pressures, so, the intake and discharge conditions of the
compressor. An example of these maps can be seen from Figure. 1.14.
Figure. 1.14: Application limit of a Screw compressor from Bitzer’s software [5]
The map’s borders define 5 different types of limit. The vertical border on the left side
define a limit regarding the lubricant. In fact, going towards lower evaporation
temperature and so pressure, the mass flow rate treated by the compressor decrease;
being the lubricant and refrigerant mixed, there will be no more enough quantity of
lubricant, for the proper operation of the compressor. The borders, respectively on the
right side and at the top of the map, are limits regarding the lubricant too. In these
cases, being the outlet temperature high, the lubricant properties will change, being no
more optimal to lubricate in a correct way the system. The sloping border on the left
side concerns the compression ratio that, in this region, is pushed to the maximum
affordable by the compressor. Being the evaporation temperature low, and condensing
temperature high, the result is a high compression ratio required. Instead, the last
border, the bottom-one, around 20-25°C, is more an historical limit than a technical
one. Each year, the producers are trying to lower it, despite refrigeration application
customers be reluctant, because they fear reliability problems; instead, for air
conditioning this lowering is already been done. Lowering this limit will bring an
enhancement of the cycle. During winter season for example, when the outside
28
temperature is low, it could be convenient condensate at lower pressure, reducing the
power absorbed by the cycle. Through this map, for each compressor, is possible to
investigate if a couple of condensation and evaporation temperature (or pressures),
match the correct operating conditions of the compressor.
1.5. Vapor-compression model validation
The first step of the present work was to realize a performance simulation model and,
then, validating it through the comparison with data given by Bitzer [4] coming from
an internal simulation software [5], of which, is possible to see the starting page in
Figure. 1.15.
Figure. 1.15: Homepage Bitzer’s software
The data provided were about two configurations of vapor-compression chillers. The
first configuration was a single screw compressor (CSH8561-125Y), with a
condensation temperature of 45°C, providing a cooling power of 184.5kWref; the
second one was with 4 reciprocating semi-hermetic compressors, simulated under two
condensation temperatures, the first as the previous case 45°C and, the second, at 15°C,
having a cooling power respectively of 136kWref and 209kWref. An example of Bitzer
simulation program’s output can be seen in Figure. 1.16.
29
Figure. 1.16: Output of Bitzer’s software in the case of single screw compressor
After the first choice about the type of compressor, the software requires as an input
the following information: the refrigerant, that can be chosen from a list of fluids
compatible with the compressor chosen; condensation and evaporation temperatures;
and other parameters like, subcooling of the liquid in the condenser and superheating
of intake gas. An example of those data is showed in Table. 1.1.
Table. 1.1: Data input of Bitzer’s software
30
Running the Bitzer’s software, the output results are: energy balance quantities like
evaporator capacity, electrical power consumed, thermal power to the condenser;
Coefficient of Performance (COP); and other data like temperature of the liquid
outside the condenser. An example of this output can be seen in Table. 1.2.
Table. 1.2: Data output of Bitzer’s software
The performance simulation model has been realized on Microsoft Excel and Visual
Basic. To perform the model validation, the same input data were used, to simulate the
cycle point by point and calibrate, on the base of the outputs, the unknow parameters
(e.g. machines’ efficiencies). Through an add-in of Excel, RefProp [6], it was possible
to calculate properties of different fluids. The starting points were evaporator and
condenser, where, outlet temperatures are known, because are input values. With these
two quantities is possible to have all the information about these two points. Using the
Refprop’s functions, named ‘’Tliq’’ and ‘’Tvap’’, it was possible to calculate the
properties of the saturated points 1 & 5 knowing the respective temperatures. Then,
knowing the vapor’s inlet temperature in the compressor and the pressure, that is
assumed to be the same of the previous point, being not too long the connection
31
between evaporator and compressor, using the function ‘’TP’’ all useful properties of
point 2 can be computed. The properties at the outlet of the compressor were founded
using a fictitious iso-entropic point and, introducing an iso-entropic efficiency that is
changed to find the best match with the outlet temperature provided by Bitzer’s data.
Then passing at the outlet of the condenser, it was possible to calculate, the properties
of the intermediate point, using the subcooling ∆T provided as input. Finally, between
point 6 and 7 was performed an iso-enthalpic lamination. At point 6 through the
function “TP”, all the relevant quantities can be calculated. Having the enthalpy before
the throttling valve, and the inlet pressure of the evaporator, calculated through
pressure loss coefficient, it was possible to have the other properties, calculating the
quality first with ‘’PH’’ function, and after employing the “PQ” option all the other
unknowns.
Below, in Table. 1.3, it is possible to see an example of cycle’s properties calculation.
To associate the number of the point to the position in the cycle is possible to refer to
Figure. 1.17.
Cycle properties
Position Point T[°C] P[MPa] h[kJ/kg] s[kJ/kg/K] ρ[kg/m3] Quality 1 -8 0.39 410 1.80 16 1
Out-eva 1' -8 0.39 410 1.80 16
In-comp 2 2 0.39 419 1.84 15
Out-comp 3 88 2.21 476 1.86 74
4 49 2.17 427 1.72 97 1 5 45 2.17 270 1.23 1018
Out-cond 5' 43 2.17 267 1.22 1030
In-valve 6 39 2.17 259 1.20 1055
Out-valve 7 -12 0.40 259 1.23 47 0.35 3s 79 2.21 466 1.84 78
Table. 1.3: Simulation of cycle properties example
32
Figure. 1.17: Scheme of the vapor-compression cycle highlighting the most important points
In the Excel simulation software, the pressure losses were guessed through the use of
reference values founded in literature; all the other parameters of subcooling and
superheating taken as equal to the input of Bitzer’s software. By means of outputs
comparison, it was possible to calibrate the machines’ efficiencies, trying to be as close
as possible to Bitzer’s result. For each configuration have been performed two
calculation, one maintaining fixed the mass flow rate of the cycle, and the other one,
fixing the refrigeration power. In both cases, it was possible to find good consistency
with the data provided by Bitzer’s software. An example of the control spreadsheet is
shown in Table. 1.4.
Quantity Flow rate
fixed Err. [%]
Refrigeration
Power fixed Err. [%] Bitzer Results
Flow rate 3601 kg/h 0.00 3621 kg/h 0.56 3601 kg/h Flow rate 1.000 kg/s 0.00 1.006 kg/s 0.56 1.000 kg/s Qrefrig 208 kW -0.56 209 kW 0.00 209 kW Qcond 237 kW -2.02 238 kW -1.47 242 kW Pcomp 33 kW -0.48 33 kW 0.07 33 kW COP 6.39 - -0.07 6.39 - -0.07 6.39 -
Table. 1.4: Model results’ check
33
The calibrated efficiencies founded in the three configurations, are reported in Table.
1.5, characterized with respect to type of compressor, condensing temperature and
refrigeration power.
Compressor Tcond [°C] Pcooling [kW] ηcomp ηiso-s
CSH8561-125Y 45 185 0.93 0.75
4 x 4JE-22Y 45 136 0.93 0.82
4 x 4JE-22Y 15 209 0.90 0.86
Table. 1.5: Calibrated efficiencies
The model’s validation means that, the way in which, all the energetic quantities were
computed, was right, and so, it was possible to go ahead using these vapor-
compression chillers’ models in the further steps.
34
2. CHP & CCHP
In this chapter, cogeneration and trigeneration concepts will be illustrated. After this
introduction, it is presented a deepening on the main machines that compose the CHP
and CCHP systems. The subsequent paragraph will be dedicated to prime mover, with
a particular focus on the most important technologies for those applications; the
paragraph closes with a sub-section on performances calculation. Then absorption
refrigeration will be illustrated, with a section dedicated to the selection of the working
fluids. The chapter closes with the typical applications, in which, CHP and CCHP
technologies are employed, and with the theme of incentives: what types and to which
systems are assigned.
2.1. Introduction on cogeneration & trigeneration
2.1.1. Cogeneration
The cogeneration, or Combined Heat and Power (CHP), is the process of simultaneous
production of mechanical energy (usually converted in electrical energy) and heat,
using only one fuel Figure. 2.1. The combined production has to occur at determined
conditions, in Italy defined by the authority for electric energy and gas. To be
recognized as cogeneration, the process must guarantee a significant energy saving
with respect to separated production, so says the Legislative Decree 16 March 1999,
No. 79, Article 2, paragraph 8 [7]. The conversion of primary energy (typically given
by a fuel) in mechanical and/or electrical energy, leads to, independently from the
technology, the production of heat, that typically, is dissipated in the external
environment. Employing cogeneration, it is possible to recover most of this
transformation’s second product, having relevant economic and energy savings. The
useful thermal energy is used to warm up buildings and/or for industrial processes.
35
The combined production allows a significant reduction of fossil fuel usage and
CO2equivalent emissions.
Figure. 2.1: Cogeneration conceptual scheme
Since this useful technology allows to be more efficient, some incentives can be
provided if the system respect determined constraints. The system must be recognized
as “Cognerazione ad Alto Rendimento” (CAR), the constraints to belong to this
category are defined in the Legislative Decree 8 February 2007, n. 20 and in the
supplementary decree D.M. 4 August 2011 [7], in the paragraph 2.5, the theme of
incentives will be treated in a more detailed way.
The directive 2004/8/CE of the European parliament defines two concepts: the units
of “micro-cogeneration”, as the cogeneration systems with maximum electrical power
lower than 50kWel, and units of “small cogeneration”, as those systems with an
electrical production lower than 1MWel [7].
Usually a cogeneration system is composed by:
- prime mover
- electrical generator
- heat recuperators
The prime mover is fed with a fuel and produces mechanical power. There are different
types of these devices, but the technologies that dominate this field are: Internal
Combustion Engines (ICE) and Micro-Gas Turbines (MGT). Through the connection
among driving motor and electrical generator, it is possible to convert mechanical in
36
electrical energy. The cogeneration, as described above, implies that, the waste heat,
coming from the first transformation, is recovered by using appropriate devices,
known as heat recuperators.
2.1.2. Trigeneration
Trigeneration, or Combined Cooling, Heat and Power (CCHP), is a particular field of
cogeneration systems, in which, besides producing electrical energy, it allows to use
part of the thermal energy recovered, by means of an absorption chiller, to produce
refrigeration power, that can be used for air conditioning or industrial processes. The
conceptual scheme of trigeneration is presented in Figure. 2.2. First fuel and excess air
are burned to drive a prime mover, which in turn drives an electrical generator, that
produces electricity for final use. The energy of high-temperature exhaust, from the
prime mover, is mainly recovered using a heat recovery unit. Using a convenient heat
transfer fluid, the recovered heat can be used in specific heating processes, and in case
of trigeneration, also to drive a cooling unit, for cooling purposes according to the site
demands.
Figure. 2.2: Trigeneration conceptual scheme
The trigeneration extends the possibilities of cogeneration, that are limited by the
necessity to have contemporaneity among electric and thermal use to have an efficient
system. This factor represents a constraint, where the need of thermal power is limited
37
to the winter period, while during summer, the cogeneration system must be stopped,
or left work dissipating the heat produced. Instead, matching a cogenerator with
dedicated absorption chiller, that can generate cooling power using heat as a source,
rather than of electricity, it is possible to exploit the cogeneration implant also in the
hotter months; when the request of air conditioning is very high, or in the industrial
sectors, in which cooling systems are needed.
2.2. CHP’s prime mover
Different heat and power generating technologies are present in literature, to serve as
prime movers for CHP or CCHP applications. These technologies could be divided
into two categories, combustion-based (reciprocating engines, gas turbines, Stirling
engines, Rankine cycles) and electrochemical-based (fuel cells) [13]. Reciprocating
engines and gas turbines are commercially mature technologies, with a wide
availability in the market, while others like organic Rankine cycles, Stirling engines
and fuel cells units are under research and development phase, with limited ways to
market.
In the context of trigeneration, it is important to consider the type, and the quality, of
the waste heat available from each type of prime mover. As said before, the
technologies that dominates the CHP and CCHP business are: Internal Combustion
Engines and Micro-Gas Turbines. In the following sub-sections, these two
technologies will be deepened.
2.2.1. Internal Combustion Engines (ICE)
Internal combustion engine is a mature technology, and can be classified in two main
types, compression ignition and spark ignition engines. It is about engines adapted for
the stationary applications and equipped with a heat recovery system. They are called
also reciprocating or piston engines, and they convert pressure into rotation motion,
using pistons contained in cylinders, where chemical reaction of the fuel combustion
takes places. They have a relatively high electrical power output and limited
38
investment cost. For these reasons, they are the most commonly used prime mover for
medium-scale (100-5000 kWel) CHP and CCHP applications [14]. Diesel oil, natural
gas or gasoline can be used to fuel these engines. Regarding the quality, and the type
of waste heat available, as depicted in Figure. 2.3, it can be recovered at different
levels, from exhaust gases (at 200-400°C) and from jacket water cooling and oil
cooling (at 90-125°C) [15].
Figure. 2.3: Simplified heat recovery process from an ICE
The cons of this type of technology are: the noisy operation, constantly necessity of
maintenance, high vibrations, high emissions rate, and due to this, the need of an
emissions abatement systems to obtain values like gas turbines ones.
2.2.2. Micro-Gas Turbines (MGT)
Gas turbine systems operate on Brayton’s thermodynamic cycle. These systems are
generally composed by: generator, combustion chamber, recuperator, compressor and
turbine connected through a shaft. Heat rejected is recovered from the hot exhaust
gases, using a heat recovery system. The heat recovered has a high quality, because of
39
its high temperature. Micro-Gas Turbine systems have been developed with size
ranging from 30 to 400kWel [16]. They have a small number of moving parts, and they
can be fuelled by various gaseous and liquid fuels. Regenerative Brayton cycle,
combined with high-speed centrifugal turbo machines are considered nowadays as
potential alternative to conventional ICE, especially for mini-scale systems [15]. Gas
turbines are more compact with respect to ICE-based systems, and require less
maintenance, and having such temperature of exhaust gases (about 250°C) is possible
to feed an absorption chiller. An example of the employment of Micro-Gas turbine in
a trigeneration system is depicted in Figure. 2.4.
Figure. 2.4: Trigeneration using a gas turbine as a prime mover [17]
However, Micro-Gas Turbine applications, in the residential and building sector, are
still very limited, due to their low electrical efficiency, and inflexibility to load profile
changes [13].
40
2.2.3. Other emerging technologies
In the field of CHP and CCHP for medium and micro generation, ICE and MGT are
the most consolidated and diffused technologies, and many solutions are commercially
available. Both the technologies have architectural features, and specific solutions,
that allow a reduction of system’s size, without significantly compromising
performances. However, there are other technologies under research or development
phase, not widely present commercially. A list of them will be presented below.
- Organic Rankine units (ORC): these engines convert heat into useful work
relying on the thermodynamic Rankine cycle. Rankine systems can be divided
mainly in two categories, based on the type of working fluid: steam Rankine
cycles, that use water as a working fluid; and organic Rankine cycles, that
employ organic fluids having either a lower, or higher boiling point [18]. Due
to their durability, cost effectiveness, lower operational temperature and
pressure, high level of safety and simplicity, ORC CHP and CCHP systems
have received increasing attention recently [13]. These organic substances
have a high molecular weight, and their characteristics allow to exploit small
enthalpy drops, at medium-low temperature, where the traditional Rankine
cycle would have limits in the turbine design, and in the investment and
operational costs. The system in its simpler configuration is composed by a
recovery boiler, turbine, condenser and a circulating pump. Most of the systems
then adopt an additional exchanger, an internal recuperator, that absorb the heat
still present in the fluid, at the outlet of the turbine, to pre-heat the feed of the
boiler. The internal recuperator pre-heats the flow entering the boiler, with a
consequent reduction of the heat to be introduced in the cycle, and an increase
of the overall efficiency. The thermal sources compatible with ORC
technology are various, from the waste heat of industrial process, to renewable
sources like solar or geothermal energy, biomass boiler or flue gases from
MTG or ICE.
41
- Stirling engines: are reciprocating motors, with a closed cylinder, where fuel
combustion takes place in separate combustion chamber, and thus it is known
as a piston external combustion engine [19]. It is a closed cycle, in fact, the
cycle encloses fixed quantity of permanently gaseous working fluid, like air or
helium. The motor can be fuelled with any source of heat, because of the
external combustion chamber, for example combustion of wood, carbon,
natural gas, biogas, liquid fuels and solar or nuclear energy. CHP systems that
employs a Stirling engine have higher flexibility, lower emissions and are
quieter in their operation compared to ICE. However, small scale Stirling CHP
systems tend to be more expensive, with low electrical efficiency in the range
of 10-15% [20]. Due to this, very few studies have been considered and
investigated Stirling engines, as prime mover for cogeneration and
trigeneration systems.
- Fuel cells: is an entirely different approach of electricity production compared
to traditional prime mover technologies. They produce direct current through
an electrochemical process, without direct combustion of the fuel. Fuel cells
are more expensive with respect to Stirling engines, but generally they show
higher electrical efficiency under varying load. The system is composed by a
fuel cell stack with cathode, anode and electrolyte, a fuel reformer for fuel
preparation and a power conditioner to transform DC to AC electricity. For
example, using hydrogen and oxygen as reactants, they react in presence of an
electrolyte to produce water, which generates an electrochemical potential,
driving an electric current through an external circuit. Different types of fuel
cells are used for CHP systems depending on the size and nature of application,
including Solid Oxide Fuel Cells (SOFC), Polymer Electrolyte Fuel Cells
(PEFC), Proton Exchange Membrane Fuel Cells (PEMFC), Alkaline Fuel Cells
(AFC), Phosphoric Acid Fuel Cells (PAFC) and Molten Carbonate Fuel Cells
(MCFC). PEFC have small dimensions, with lower costs, compared to other
fuel cells, but also lower electrical efficiency [13]. SOFC micro-CHP systems
can use natural gas as a fuel, because they operate at very high temperature, up
42
to 1000°C, due to their hard ceramic-based electrolyte, and can reach a very
high electrical efficiency up to 55% [13]. PEMFC systems have low operating
temperature, and they can meet shifts in power demands. The pros of these
technologies applied to CHP are that, the systems are very quiet with high level
of reliability, modularity and rapid adaptability to load changes. The cons are
that, the investment cost is very high, and the design is complex compared to
other technologies.
2.2.4. Prime mover performances
In the present work, it has been chosen to use a system, that employs an internal
combustion engine as a prime mover. The choice was among the two most popular
technologies in this application field (CHP and CCHP for medium size), because they
are reliable, diffuse, commercially available and since the scope of this thesis is not to
focus on the optimisation of the prime mover. Among MGT and ICE has been chosen
the latter. The reasons of this choice are the higher specific of cost of MGT as can be
seen from Figure. 2.5; the use of Br-Li absorption chiller, since coupled with ICE, they
allow the recovery of waste heat not only from flue gases, but also form jacket water
and cooling oil (for further deepening on absorption chillers’ refrigerants refer to
paragraph 2.3); and finally, because of the high variability of the case study’s load,
since MGT technology is not flexible to accommodate rapid change in load profile.
43
Figure. 2.5: Specific cost comparison between ICE and MTG technologies for small size CHP [21]
For the simulation of the case studies, it has been necessary to characterize the
performances of the prime mover varying size and load. To achieve this task, it was
used a previous work done in a thesis, about cogeneration and trigeneration systems
sizing and control strategy optimization [21]. In this work there is a machines list, in
which are present various data about: CHP prime movers, natural gas boilers,
absorption chillers and thermal storage. This wide amount of data has been collected
by technical data sheets provided by several companies. The 2nd version of machines
list is an excel spreadsheet, in which, under the CHP’s section, are present parameters
like electrical and thermal power produced, efficiencies and specific costs. A task
performed in the present work was to create a 3rd version, including the parameters at
partial load, fundamental to study a system with high frequency load changes. These
data were taken from technical sheets. In this way, it was possible to study how the
prime mover operate at partial load, so the amount of electric energy generated, and
the thermal power available.
For the CHP’s prime mover working at full load have been studied: the trend of
electrical efficiency as function of electrical power, the trend of thermal efficiency as
function of electrical and thermal power. For thermal power two correlation have been
44
used, depending on the design logic, that will be explained in the following chapters.
Following, it is possible to see the trend of efficiencies in graphs represented in Figure.
2.6, Figure. 2.7 and Figure. 2.8 with the relative interpolation equation in (2.1), (2.2)
and (2.3).
(2.1) 𝜂_𝑒𝑙 = 0.0267 ∗ ln 𝑃𝑒𝑙 + 0.2336
(2.2) 𝜂_𝑡ℎ = −0.032 ∗ 𝑙𝑛 𝑃𝑒𝑙 + 0.6736
(2.3) 𝜂_𝑡ℎ
𝜂_𝑒𝑙= −0.254 ∗ ln 𝑃𝑡ℎ + 2.8344
Figure. 2.6: ICE’s electrical efficiency as function of electrical power
0,00
0,05
0,10
0,15
0,20
0,25
0,30
0,35
0,40
0,45
0,50
1 10 100 1000 10000
η_e
l
Pel [kW]
45
Figure. 2.7: ICE’s thermal efficiency as function of electrical power
Figure. 2.8: ICE’s ratio thermal-electrical efficiency as function of thermal power
Once the size of the engine is chosen, electrical and thermal efficiency at full load are
known, and so, the first principle total efficiency of the motor (2.4). To avoid results’
incongruency of partial load curves interpolation, it has been done an assumption: the
total efficiency remains constant at partial load, the change is in the shares of electrical
0,00
0,10
0,20
0,30
0,40
0,50
0,60
0,70
0,80
1 10 100 1000 10000
η_t
h
Pel [kW]
0,00
0,50
1,00
1,50
2,00
2,50
3,00
3,50
1 10 100 1000 10000
η_t
h/η
_el
Pth [kW]
46
and thermal efficiency. It has been verified, that, this assumption represents a
reasonable approximation, as can be seen in Figure. 2.9.
(2.4) 𝜂_𝐼 = 𝜂_𝑒𝑙 + 𝜂_𝑡ℎ
Figure. 2.9: ICE’s total efficiency at different loads as function of electrical power
After this premise, only the trend of electrical efficiency must be study, to fully
characterise the system’s performances at partial load.
The trend of electrical efficiency at 75%, 50% and 25% of the capacity has been
reported on graphs as function of electrical power output. To avoid incongruency in
the interpolation, with not reasonable results, it has been considered the ratio of
electrical efficiency at the current load with respect to nominal one. Following, it is
reported as an example the work done at half load, Figure. 2.10, with the relative linear
interpolation equation (2.5).
(2.5) 𝜂_𝑒𝑙@50%
𝜂_𝑒𝑙@𝑛𝑜𝑚= 3 ∗ 10−5 ∗ 𝑃𝑒𝑙 + 0,8752
0,50
0,55
0,60
0,65
0,70
0,75
0,80
0,85
0,90
0,95
1,00
0 500 1.000 1.500 2.000 2.500 3.000
η_I
Pel [kW]
η_I
η_I_75%
η_I_50%
47
Figure. 2.10: ICE’s electrical efficiency at partial load with respect to nominal one as function of
electrical power
2.3. Absorption Chiller
An absorption refrigerator is a device that uses a heat source (e.g. solar energy, a fossil-
fuelled flame, waste heat from factories, or district heating systems), to provide the
energy needed to drive the cooling process. There are direct flame machines (or direct
supply), in which the thermic source is composed by a fuel (usually natural gas, LPG,
biomass or other fuels); and machines in which the hot source is a heat transfer fluid
(water, diathermic oil, hot flue gases, steam), supplied at the right thermal level
(indirect supply machines).
Absorption machines are devices that exploit the solubility, and high affinity, of two
substances, of which one serves as refrigerant and the other as absorbent.
An absorption refrigerator changes the gas back into a liquid using a method that needs
only heat and has no moving parts other than the refrigerant itself.
y = 3E-05x + 0,8752
0,60
0,65
0,70
0,75
0,80
0,85
0,90
0,95
1,00
0 500 1.000 1.500 2.000 2.500 3.000
η_e
l@5
0%
/η_e
l@n
om
Pel[kW]
48
The combinations refrigerant/absorbent that have been established commercially are
two:
- water and lithium bromide: H2O-LiBr, where water works as refrigerant;
- ammonia and water: NH3-H2O, in which the refrigerant role is played by
ammonia.
2.3.1. Single-effect cycle
From the thermodynamic point of view, the cooling cycle works with three thermic
source: the cold source, that is at the lowest temperature, and is constituted by the
environment that needs to be cooled; the hot sink, in which heat is discharged, is the
external environment, and it is at intermediate temperature; finally, through the source
at highest temperature is introduced heat in the cycle (flame or hot heat transfer fluid).
Figure. 2.11: Single-pressure absorption refrigerator scheme and its P-T diagram
Figure. 2.11 depicts a scheme of single-effect absorption refrigerator system with its
P-T diagram. All such systems have these components: a generator (named also
desorber or concentrator), an absorber, a pump, a condenser, a throttle valve and an
evaporator.
In comparison, a vapor-compression refrigerator uses a compressor, usually powered
by either an electric, or internal combustion motor, to increase the pressure on the
gaseous refrigerant.
49
The low-pressure part of the machine is constituted by evaporator and absorber,
thermally separated, but in communication among them. In the evaporator is present
the refrigerant, that evaporates removing heat from the cold source (useful effect); the
refrigerant evaporated is absorbed by the rich solution (in absorbent), present in the
absorber. The absorption of the vapor generates in the absorber a low-pressure
environment, leading to a continuous flux of vapor, from the evaporator to the absorber
[28]. The mixture obtained in the absorber must be removed with a pump, to avoid
the progressive dilution by the refrigerant. The role of the pump is not only to
overcome the loss of pressure in the cycle, but also to raise the pressure of the fluid to
its maximum value, sending the mixture in the component named generator. In the
generator, heat is provided at the right thermic level in order to evaporate the
refrigerant present in solution. The refrigerant evaporated is first condensed, giving
heat to the intermediate temperature source, and then it is expanded through a
lamination valve, to re-enter back in the evaporator and to restart the cycle.
The solution rich in absorbent substance, produced in the generator, is laminated too,
to get back in the absorber and restore the right solution’s concentration.
Two different pressure are needed: a higher-level for the group generator-condenser,
in order to have condensation according to external temperature, and a lower-level in
the group absorber-evaporator to remove heat through evaporation.
To enhance the performance of the machine, it is present a regenerative heat
exchanger, between the concentred solution exiting the generator (rich in absorbent),
and the lean solution coming from the absorber. In this way is obtained the pre-heating
of the liquid that is sent to generator, with obvious saving of the heat introduced,
having at the same time cooling of the solution entering the absorber, with consequent
lowering of the heat to be removed. In the absorber, in fact, is present a cooling circuit
(usually the same of the condenser).
For absorption machine is defined an efficiency, named COP (as for vapor-
compression chiller), defined as, ratio of useful effect with respect to energetic
50
expenses sustained for the operation of the machine. The coefficient of performance
of an absorption system is obtained from equation (2.6).
(2.6) 𝐶𝑂𝑃𝑎𝑏𝑠𝑜𝑟𝑝𝑡𝑖𝑜𝑛 =�̇�𝑐𝑜𝑜𝑙𝑖𝑛𝑔
�̇�𝑖𝑛+�̇�𝑝𝑢𝑚𝑝
The work input for the pump is negligible relative to the heat input at the generator,
therefore, the pump’s work is often neglected for the purposes of analysis [22].
In a single-pressure (or single-effect) absorption system, the COP is in the range 0.7-
0.8 if the working fluid pair is H2O-LiBr, instead for NH3-H2O pair, the values are
lower as Figure. 2.12 shows; the COP goes to 1.1-1.35 in double-effect machines as
reported by Xu [23].
Figure. 2.12: Absorption chiller’s COP varying fluid and refrigeration capacity
The double-effect absorption chillers are characterized by the adoption of two cycle,
as the one reported before, used together in series in order that, the heat of condensation
of the upper cycle will be the inlet energy to the generator of the bottom cycle.
0,00
0,10
0,20
0,30
0,40
0,50
0,60
0,70
0,80
0,90
0 200 400 600 800 1000 1200 1400
CO
P
Pcooling [kW]
H2O-LiBr
NH3-H2O
51
The values of the COP are relatively low, but has to be considered that, these devices
use thermal energy at relatively low temperature, so a supply source less valuable,
from the thermodynamic point of view, than electric energy used by traditional vapor-
compression cycles.
The advantages offered by absorption machines, besides the possibility to use waste
heat and the implementation of fluids not harmful for the ozone, are surely the high
reliability, deriving from the presence of very few moving parts, high life-time, low
noisiness and absence of vibration, the reduced electric energy request and the good
performances at partial load.
2.3.2. Working fluids pairs
The performances of an absorption system are critically dependent on the chemical,
and thermodynamic properties of the working fluid [24]. A fundamental requirement
of absorbent/refrigerant combination is that, in the liquid phase, they must have a
margin of miscibility, within the operating temperature range of the cycle. The mixture
should also be chemically stable, non-toxic and non-explosive. In addition to these
requirements, Holmberg [25] reports that, the following ones are desirable.
- There should be a difference, large as possible, in boiling point (at the same
pressure) between the pure refrigerant and the mixture.
- The refrigerant should have high heat of vaporization and high concentration
within the absorbent, in order to maintain low circulation rate between the
generator and the absorber, with respect to the cooling capacity.
- Favourable transport properties, that influence heat and mass transfer.
- The couple of fluid selected should be non-corrosive, environmentally friendly,
and cheap.
Many working fluids are suggested in literature. A survey of absorption fluids
provided by Marcriss [26] suggests that, there are 40 refrigerant and 200 absorbent
52
compounds available. However, the most common working fluids are NH3-H2O and
H2O-LiBr.
The use of H2O-LiBr pair, for absorption refrigeration systems, began around 1930.
The most important advantages, using this couple as a working fluid, are:
- the non-volatility of the absorbent (LiBr) avoid the introduction of the rectifier,
to prevent the presence of the absorbent in the sections after generator, that
would lead to an impairment of the performances;
- the refrigerant (H2O) has an extremely high heat of vaporisation;
- the heat input temperatures to the generator are relatively low, around 80-90°C
[21].
However, in addition to these outstanding features, H2O-LiBr has also negative
characteristics:
- the use of water as refrigerant limits the low temperature application to that
above 0°C;
- as water is the refrigerant, the system must be operated under vacuum
conditions;
- at high concentration, the solution is prone to crystallisation;
- the pair is corrosive to some metals and expensive.
The other important fluids pair, in vapor-absorption refrigerator, is NH3-H2O as
anticipated above. Since the invention of an absorption system, NH3-H2O has been
widely used for both cooling and heating purposes. Both NH3 (refrigerant), and H2O
(absorbent) are highly stable for a wide range of operating temperature and pressure.
Ammonia, as water, has a high latent heat of vaporisation, that is necessary for an
efficient performance of the system. The most important advantage, with respect to
H2O-LiBr pair, is that it can be used for very low temperature applications, as the
freezing point of ammonia is -77°C. This pair is also environmentally friendly and
cheap.
53
However, there are many drawbacks employing this couple of fluids:
- the rectifier is needed, because of the volatility of both NH3 and H2O, to strip
away water that evaporates with ammonia (without a rectifier water would
accumulate in the evaporator and offset the system performance);
- its high pressure, toxicity and corrosive action to copper and copper alloy;
- requires higher temperature of heat input to the generator, excluding the
recovery of low-temperature heat.
In the work proposed by Steiu [27], NaOH (sodium hydroxide) was added to the NH3-
H2O mixture, to improve the separation of NH3 in the generator, and reduce both
chiller’s driving temperature and rectifier losses. Cycle simulation, based on their
experimental data, showed that the COP was about 20% higher than the conventional
NH3-H2O, under same conditions.
In the present work, a single-effect H2O-LiBr absorption chiller has been chosen,
because of these reasons:
- the fluids pair offers higher reliability and in general better performances;
- it allows to exploit also low-temperature heat input (waste heat from jacket
water and cooling oil in ICE could be employed);
- it has lower specific cost, compared to absorption chiller with NH3-H2O pair as
can be seen in Figure. 2.13;
- for the case study presented extra low temperatures are not needed (used to
sub-cool till 10°C).
54
Figure. 2.13: Specific cost with respect to cooling capacity for NH3-H2O and H2O-LiBr pairs [21]
2.4. Typical applications
The combined production of electric energy and heat finds application both in
industrial, mostly in self-production, and civil sectors. The heat is used for industrial
processes, or in civil sector for urban heating, through district heating network, as well
as for cooling through absorption systems. The heat is exploited as vapor or other heat
transfer fluids (e.g. hot water, diathermic oil), or as hot air. The electric energy, that
can rely on an extended distribution network, is self-consumed or introduced in the
grid. The favoured users for cogeneration are the ones characterised by a constant
demand in time of electric and thermal energy. For example, hospitals, pools and sports
centers, shopping centers as well as food industries, paper mills, industries related to
oil refining and chemical industries. In case of civil uses, including environments
heating or urban district heating, the heat is generally produced at relatively low
temperature, and the heat transfer fluid is water. In case of industrial uses, the heat is
generally produced at higher temperature and pressure. There are also hybrid
situations, in which, there is production of heat at different temperature and pressure
levels. In these cases, usually, there is a unique location of usage (e.g. an industrial
€-
€200,00
€400,00
€600,00
€800,00
€1.000,00
€1.200,00
€1.400,00
€1.600,00
- '200 '400 '600 '800 1'000 1'200 1'400 1'600
Spe
cifi
c co
st [
€/k
W]
Pcooling[kW]
LiBr (Single effect)
Ammonia-Water
55
plant), where the more valuable heat is intended to the workings, instead the low-
temperature heat is used to heat the production environments. In some industrial
sectors, combined production of electric energy and heat is a widely consolidated
option, which may become more relevant in terms of contribution to national electric
demand and energy saving.
Therefore, cogeneration is a widespread practice in the industrial sector, but in the last
years is spreading also in the tertiary and residential sectors. It is an example the
diffusion of small-size cogeneration systems, installed in independent houses or
condominiums. Also, district heating networks, that in Italy are mainly fuelled by heat
produced by cogeneration, have registered, in the last years, a constant increase in term
of network extension and volume heated [29]. Lastly, it is necessary to highlight that,
exploitation of useful heat produced by the cogeneration system also for cooling
(CCHP or trigeneration), allows to maximize the exploitation of the thermal energy,
making the use of the system convenient for a higher number of hours per year.
2.5. Incentives
The combined production, with respect to separate production of the same quantities
of electric energy and heat, if efficient, leads to:
- an economic saving consequent to the lower usage of fuel;
- a reduction of the environmental impact, consequent to the reduction of both
emissions and residual heat released in ambient (lower atmospheric pollution
and thermal pollution);
- lower transmission and distribution losses for the electrical national system,
consequent to the location of production system near to users, or self-
consumption of the produced energy;
- substitution of less efficient and more pollutant heat supply methods (boilers,
both for civil or industrial uses, characterised by lower efficiency levels, high
environmental impact and low flexibility relatively to the use of fuels).
56
For the aforementioned reasons, institutions provide incentives to promote
cogeneration and trigeneration. To benefit from these economic incentives, some
characteristics are required. In this paragraph will be illustrated how, in Italy, the
incentives for cogeneration are assigned.
There are two types of incentives for regarding energy efficiency:
- “Cogenerazione ad Alto Rendemento’’ (CAR): incentives for high efficiency
cogeneration;
- “Titoli di efficienza energetica’’ (TEE) or “Certificati Bianchi’’ (CB): denote
how many tons of equivalent oil (toe) are saved, thanks to determined
intervention, done to make more efficient the energy production process.
The information present in the present paragraph are related to documents [29] & [30].
2.5.1. “Cogenerazione ad Alto Rendimento’’ CAR
In this subchapter will be illustrated the reference regulatory framework about high
efficiency cogeneration.
- Directive of February 11, 2004 of the European Parliament and of the Council
n. 2004/8/CE: prefix the goal to increase the energetic efficiency and enhance
the security of fuel supply, in the internal market, of the “Cogenerazione ad
Alto Rendimento’’. This directive defines the concept of useful heat.
- Legislative Decree 8 February 2007, n.20: implements the directive
2004/8/CE from 1st January 2011. To define CAR, utilises a criterion based on
PES (Primary Energy Saving) (2.7), that represent the saving of primary energy
that cogeneration allows to obtain, with respect to separate production, of the
same quantity of electric and thermal energy.
57
(2.7) 𝑃𝐸𝑆𝐶𝐻𝑃 = 1 −𝐸𝑓𝑢𝑒𝑙
𝐸𝑒𝑙_𝑢𝑠𝑒𝑓𝑢𝑙
𝜂𝑒𝑙_𝑟𝑒𝑓∗𝜂𝑛𝑒𝑡𝑤𝑜𝑟𝑘±∆𝜂𝑇𝑎𝑚𝑏+
𝐸𝑡ℎ_𝑢𝑠𝑒𝑓𝑢𝑙
𝜂𝑡ℎ_𝑟𝑒𝑓
Following, in Table. 2.1, Table. 2.2 and Table. 2.3, are reported the values of
reference efficiencies and corrective factors to be employed in the PES’s
calculation.
Reference efficiencies (Regulation 12th October 2015)
Fuel type ηel_ref ηth_ref water
ηth_ref steam,
gas <250°C
ηth_ref gas
>250°C
Natural gas, LPG, LNG, biomethane 53.0% 92% 87% 84%
Diesel oil, fuel oil 44.2% 85% 80% 77%
Biofuels 44.2% 85% 80% 77%
Biogas 42.0% 80% 75% 72%
Woody biomass 33.0% 86% 81% 78%
Urban waste 25.0% 80% 75% 72%
Table. 2.1: Reference efficiencies for PES’s calculation
Grid efficiency - ηnetwork
Grid voltage Power
introduced Power self-consumed
<0.45 kV (LV) 88.8% 85.1%
0.45-12 kV (MV) 91.8% 89.1%
12-50 kV (MV) 93.5% 91.4%
50-100 kV (MV) 95.2% 93.6%
100-200 kV (HV) 96.3% 95.1%
200-345 kV (VHV) 97.2% 96.3%
>345 kV (VHV) 100.0% 97.6%
Table. 2.2: Grid efficiency for PES’s calculation
∆ηTamb corrective factor (only for gaseous fuels)
Climate zone Average
temperature [°C] Correction factor in percentage points
Area A 11.315 +0.369 Area B 16.043 -0.104
Table. 2.3: Ambient temperature’s corrective factor for PES’s calculation
58
In Table. 2.3, area A refers to the following regions: Valle d'Aosta, Trentino
Alto-Adige, Piemonte, Friuli-Venezia Giulia, Lombardia, Veneto, Abruzzo,
Emilia-Romagna, Liguria, Umbria, Marche, Molise and Toscana. Instead, area
B refers to: Lazio, Campania, Basilicata, Puglia, Calabria, Sardegna and
Sicilia.
- Decree of the Ministry of Economic Development of August 4, 2011: it
completes the directive 2004/8/CE and legislative decree 8 February 2007. The
aforesaid decree explicates the methodologies to rate the operating of a unit as
CAR. It defines when the production of electric energy, from combined
production, can be considered entirely from cogeneration, based on the type of
technology and on the yearly first principle efficiency. If the production unit
shows a first principle efficiency lower to the threshold values, it defines a
method to divide the electrical energy produced in two parts, of which, one
quantifiable as cogeneration, for detailed information refer to document [29].
The decree resumes that, CAR systems are identified through PES index. The
PES must be higher than 10%, instead, for small or micro cogeneration units
(so respectively of power lower than 1MW and 50kW), it is enough to have a
positive PES.
- Decree of the Ministry of Economic Development of September 5, 2011: it
establishes the conditions and procedures to access cogeneration’s support
scheme. As settled by this decree, the new cogeneration units or the
reconstructed ones, are entitled, for each year that they satisfy the CAR
requirements, to the recognition of II type “Titoli di Efficienza Energetica’’ (or
“Certificati Bianchi’’), in proportional number to the achieved energy saving
(RISP). The number of “Certificati Bianchi’’, of which a producer is entitled
year by year, is calculated on the basis of art. 4 of the Ministerial Decree of 5
September 2011.
59
- Guidelines for the implementation of the Decree of the Ministry of
Development Economic September 5, 2011: its goal is to exemplify the
calculation method of the relevant quantities for the purposes of CAR
recognition and the access mechanism of “Certificati Bianchi’’.
- Legislative Decree 4 July 2014, n. 102: implements the Directive 2012/27/EU,
in which, the Directive 2004/8/CE and Directive 2006/30/EU have been
repealed. The present legislative decree has not introduced any change with
respect to the legislation described before.
- Delegated Regulation (EU) 2015/2402 of 12 October 2015: it has revised the
reference efficiencies values, for the separated production of electric energy
and heat, to calculate and verify the PES index. The new efficiencies,
dependent from the type of fuel and the starting date of operations, are applied
from the year 2016.
- Decree of the Ministry of Economic Development March 16, 2017: applies to
high efficiency micro-cogeneration units and to micro-cogeneration units
fuelled by renewables sources. The goal of the present decree is to minimize
the burden on producers, and rationalize the exchange of information between
municipalities, network operators and GSE about realization, connection and
operation of these type of systems.
2.5.2. GSE’s role
GSE in the high efficiency cogeneration:
- withdraws electric energy with dispatching priority, and recognizes to the
operator, the prices defined by the authority for electricity;
- certifies the operation as CAR for the cogeneration units that request it;
- determines the number of CB, of which the units recognized as CAR, are
entitled;
60
- if the producer requests it, proceeds to the withdrawal of CB, to a price defined
on the starting date of operation in Decree of the Ministry of Economic
Development of September 5, 2011;
- it performs activity of verification and control on incentivized facilities,
communicating to Ministry of Economic Development the result of the
inspection.
2.5.3. “Certificati Bianchi’’ CB type II-CAR
The article 4 of the Ministerial Decree of 5 September 2011 imposes that the
cogeneration units have the right, for each solar year in which they satisfy the CAR’s
requirements, to the release of CB, in commensurate number to the primary energy
saving achieved in the year, if positive, calculated as follow (2.8):
(2.8) 𝑅𝐼𝑆𝑃 =𝐸𝐶𝐻𝑃
𝜂𝐸_𝑅𝐼𝐹+
𝐻𝐶𝐻𝑃
𝜂𝑇_𝑅𝐼𝐹− 𝐹𝐶𝐻𝑃
Where:
- RISP: is the primary energy saving, expressed in MWh, realized by the
cogeneration units in the solar year, for which has been requested the access to
the support scheme.
- ECHP: is the electric energy produced by “CHP part’’ of the cogeneration unit
in the same solar year.
- HCHP: is the useful heat produced by the cogeneration unit in the same solar
year.
- FCHP: is the feeding energy consumed by “CHP part’’ of the cogeneration unit
in the same solar year.
- ηTRIF: is the average conventional efficiency of the Italian thermal production
park. Assumed equal to 0.82 in case of exhaust gas direct use and, 0.90 in case
of hot water/steam production.
61
- ηERIF: is the average conventional efficiency of the Italian electric production
park, assumed equal to 0.46 and corrected as reported in sub-section 2.5.1.
On the base of the primary energy saving (RISP), the cogeneration unit has right to a
number of CB (2.9) equal to:
(2.9) 𝑁° 𝐶𝐵 = 𝑅𝐼𝑆𝑃 ∗ 0.086 ∗ 𝐾
Where 0.086 is the coefficient for the conversion of MWh in toe, and K is a coefficient
function of the unit’s power, calculated through the following formula (2.10).
(2.10) �̇�𝐶𝐻𝑃_𝑎𝑣𝑒𝑟𝑎𝑔𝑒 =𝐸𝐶𝐻𝑃
ℎ𝑜𝑝
If an unambiguous accounting of the number of operating hours (hop) is not possible,
the formula is based on the number of equivalent hours.
2.5.4. “Certificati Bianchi’’ CB or “Titoli di Efficienza Energetica’’ TEE
In accordance with the procedures established by the D.M. 11 January 2017, the
project of energy efficiency, predisposed to achieving national quantitative saving
targets, can be performed by obliged subjects (or society controlled by them), through
direct actions or intervention to increase the energy efficiency, realized by subjects
admitted to the mechanism.
The energy efficiency projects are eligible, to access the CB mechanism, if:
- realized by the same subject project holder, in one or more sites;
- realized with an intervention’s starting date subsequent to application’s
transmission to GSE;
- they generate additional energy savings, i.e. primary energy saving calculated
as difference between baseline consumption and the energetic consumption of
the configuration after the intervention;
62
- there is a suitable documentation attesting that, in the intervention have been
used new components, or regenerated ones, for which, CB have not been
recognized yet;
- prepared and transmitted to the GSE, according to the type of project, as
expected in Annex 1 of the D.M. 11 January 2017;
- classifiable among the types of intervention reported in Table 1, Annex 2, of
the D.M. 11 January 2017.
On the base of the gross primary energy saving (RL measured in MWh), the
cogeneration unit has the right, for the period of recognition, to a number of TEE (2.11)
equal to:
(2.11) 𝑁° 𝑇𝐸𝐸 = 𝑅𝐿 ∗ 0.086 ∗ 𝑘
Where k is a parameter function of the useful life; for the first half, it is assumed the
value of 1.2, instead for the remaining half 0.8.
The useful life, between 7-10 years, is given in a table, as function of the type of
intervention. For example, for the case study treated in the present work, it is possible
to refer, under the section industrial sector, to installation of chiller groups and heat
pumps, including freezing and refrigeration systems, where the indicated useful life is
7 years.
In conclusion, both type of CB, TEE or type II derived from CAR, can be applied to
the present situation. Therefore, the two different approaches will be tested on the case
studies, to figure out which one is more convenient in terms of CB number.
63
3. HYBRID CCHP
In the third chapter, the hybrid CCHP configuration tested during the present thesis
will be presented. Before, there is an introduction on the existent technologies and
similar schemes, then is reported the actual scheme and the work performed on it.
3.1. Literature review
In this paragraph, a review of similar application regarding hybrid combined cooling,
heat and power will be presented, configurations present in literature, that study the
integrated use of vapor-compression and absorption refrigeration.
3.1.1. Flexible chiller with hybrid absorption/compression cycle
In this concept, studied by Schweigler [8], an electrically driven high-speed turbo-
compressor, which is directly integrated in a single-stage H2O-LiBr absorption cycle,
provides a pressure lift for the refrigerant vapor transmitted from the evaporator to the
absorber, or from the desorber to the condenser, resulting in increased flexibility of the
cycle, for operation under non-standard boundary conditions.
Two general system concepts for a single stage hybrid sorption & compression cycle
under European climate conditions have been identified Figure. 3.1.
64
Figure. 3.1: Configurations of the two concepts DECO & EVA
The EVA-concept: comprising a high-speed turbo-compressor between evaporator
(E1) and absorber (A1). The DECO-concept: the compressor is arranged between
desorber (D1) and condenser (C1). In Figure. 3.2, are shown the effects of the
mechanical vapor compression, for EVA and DECO configuration, for a cycle with
fixed temperatures of the evaporator and desorber.
Figure. 3.2: Pressure-Temperature diagram of EVA & DECO concept respectively
The study then has been performed only for EVA configuration, because it has the
following advantages in comparison to DECO concept:
65
- it allows for lower salt concentration of sorbent solution, so an increase of
temperature level of cooling water can be accomplished;
- the compressor handles saturated vapor, allowing for same pressure ratio less
compression work.
For the EVA configuration, 3 operating modes have been investigated. In the mode 1,
an increased temperature lift is accomplished, so it is possible to reject heat in an
eventual hotter environment. For mode 2, it is possible to reduce the driving heat
temperature, thus using low-grade heat sources. In the third and last modality, the
capacity booster, is increased the chilled water capacity.
The results given by this study have shown that, respectively for the 3 different
modalities, a compression ratio of 3 either allows: a rise of the cooling water inlet
temperature of 10 K; enables the use of low temperature waste heat, below 70 °C; even
doubles the chilled water capacity under given standard boundary conditions.
3.1.2. Hybrid refrigeration system recovering condensation heat
In order to enhance the performance of conventional absorption refrigeration systems,
Wang [9] has performed a study on an absorption-compression hybrid refrigeration
system, that can recover all condensation heat for the generation of refrigerant, by
improving the grade of condensation heat through a vapor compressor. In Figure. 3.3,
are shown the scheme differences among a conventional absorption refrigerator and
the unconventional scheme proposed by Wang [9].
66
Figure. 3.3: Configuration of a traditional absorption refrigeration system (left) compared with an
absorption-compression hybrid one (right)
The refrigerant vapor, which may be a mixture (NH3-H2O was used as a working fluid
pair), is produced as usual in the generator. After the rectifier, is installed a
compression unit, that compresses the pure refrigerant vapor (NH3), thus increasing its
temperature and pressure, making it become superheated vapor. Then the vapor flows
back to generator condensing into liquid. So, in this solution, condensation heat is not
rejected to the environment, but recovered for the process of generation, with the help
of the compressor. The generation heat, entered from outside, is reduced or even
eliminated in this system, and the performance of absorption refrigeration system
could therefore potentially be improved. Furthermore, the sub-cooler, before the
throttle valve, is useful for improvements of cooling capacity, efficiency and stability
of the system. The pressure-temperature diagram, of this unconventional solution, is
illustrated in Figure. 3.4. There are at least three pressure levels, with respect to the
two pressure levels of conventional single-effect absorption refrigeration.
67
Figure. 3.4: P-T diagram of an absorption-compression hybrid refrigeration system
This innovative concepet has been investigated using a sensitivity analysis on the
variation of generation, evaporation and cooling water temperature. Morover, it has
been studied whether to use a sub-cooler and the choice among single or two-stage
compression.
From this work, has emerged that, absorption-compression hybrid refrigeration system
recovering heat for condensation can improve the grade of condensation heat,
replacing part of the generation heat, entered form outside, at the cost of a little
electricity. The comparison between conventional and hybrid configuration has drawn
some conclusions about this new system:
- it can decrease generation heat introduced in the cycle by 70–80%, compared
with that in conventional configuration, as Figure. 3.5 shows, having a Primary
Energy Efficiency (PEE) that can reach 97.1%;
- it can recover all condensation heat, which will largely decrease the volume of
the condenser/sub-cooler, and benefit the miniaturization of the architecture;
- compared to conventional systems, the hybrid configuration allows to employ
lower generation temperature, maintaining the same cooling capacity, as
Figure. 3.6 depicts;
68
- the two-stage compressor with intermediate cooling is suitable due to its
effective control of the discharge temperature. Moreover, the use of a sub-
cooler can increase both the cooling capacity and Primary Energy Efficiency
(PEE) by 2.2%.
Figure. 3.5: Effects of generation temperature on generation heat, comparison among
conventional and hybrid system
69
Figure. 3.6: Effects of generation temperature on cooling capacity, comparison among
conventional and hybrid systems
3.1.3. Solar absorption-subcooled compression hybrid cooling system
In this work, Zeyu Li [10] aims to contribute to study the solar absorption-subcooled
compression hybrid cooling system, to show the potential of this technology. The
evaluation is by the comparison with solar PV cooling (most economical solution of
solar cooling).
The proposed configuration can be seen in Figure. 3.7. The hybrid system consists of
an absorption sub-system (H2O-LiBr single-effect machine) and a compression one
(conventional vapor-compression chiller, with R410A as a working fluid). The
absorption sub-system is exclusively driven by solar energy. The cooling output of
absorption chiller sub-cools the refrigerant of vapor-compression refrigeration in the
sub-cooler, and finally transfers to the evaporator of compression sub-system.
Accordingly, the electric energy of compression sub-system is saved.
70
Figure. 3.7: Arrangement of a solar absorption-subcooled compression hybrid cooling system
The evaluation has been performed through the dynamic simulation, based on the
monthly typical cooling demand and meteorological data, of the 10-floor-office-
building locating at subtropical Guangzhou.
It has been found that, this new scheme is economically better than the solar PV
cooling, though the saving of CO2 emission in this case is less. Besides, the system
becomes more economical and attractive as the performance of its compression
subsystem is poor. In particular these conclusions can be drawn.
- The payback periods of solar absorption-subcooled compression hybrid
cooling system and solar PV cooling system are 20.3 and 20.4 years,
respectively [10].
71
- The net present value of the SASCHCS is 50% as much as higher than that of
solar PV cooling system.
- The savings of annual CO2 emission, in the system studied and solar PV
cooling system are 8.8 and 17.2 tons, respectively.
- The economic advantage reduces with the decrease of collector price, as
Figure. 3.8 shows. The effect of collector price, on the payback period and net
present value, in the SASCHCS, is weaker, which can be explained by the
cheaper price of Evacuated Tube Collector (ETC).
- The solar absorption-subcooled compression hybrid cooling system becomes
to be the more economical, and attractive solution, as the performance of its
compression subsystem is poor. The payback period goes down to 16.2 years,
as the COP of compression subsystem comes down by 20%.
Figure. 3.8: PBP and NPV as function of price of collector
72
3.2. New concept for hybrid CCHP
3.2.1. Arrangement
In the present work, it has been proposed a new configuration, of hybrid cooling, that
applies to a real case study, regarding refrigeration cells for fruit storage. Usually, the
cooling cells for food storage applications involve the use of a vapor-compression
chiller. The idea proposed, it is to install a cogeneration unit and an absorption chiller,
that will enhance the refrigeration effect, using waste heat coming from the prime
mover. The internal combustion engine will generate electric energy, that will be used
by the old vapor-compression refrigeration system, and it will produce, at the same
time, useful heat, that will be exploited by an absorption refrigerator. The useful effect
of the new absorption chiller installed, it is used to sub-cool the refrigerant at
condenser’s outlet of the vapor-compression section, providing margin to increase the
cooling effect, or maintaining the same refrigeration power, to reduce the electric
energy consumed by the compressor. Following, in Figure. 3.9, is shown a simplified
block scheme of hybrid Combined Cooling, Heat and Power configuration, that it has
been studied.
73
Figure. 3.9: Hybrid CCHP conceptual scheme
A temperature-entropy diagram instead is depicted in Figure. 3.10, and it shows the
effect of the sub-cooling on the vapor-compression cycle. Since it has been decided to
use a lithium bromide absorption chiller, for the reasons reported in paragraph 2.3 sub-
section 2.3.2, the temperature that is possible to reach is limited, due to constraints of
the fluid adopted as refrigerant, which is water. From the machine catalogue created
by Casati & Chinetti [21], under the section absorption machine, it is possible to see
that the minimum temperature reached by H2O-LiBr single-effect absorption chillers
is 7°C. So, it has been chosen to sub-cool the exit of the vapor-compression condenser
to 10°C, considering a pitch-point ∆T of 2-3°C. It is a reasonable assumption, if, a
counter-current plate heat exchanger, is installed.
74
Figure. 3.10: Temperature-entropy diagram comparing a traditional vapor-compression cycle and
one with sub-cooling at condenser’s outlet
The heat that can be extracted through the absorption chiller, so, consequently the heat
recovered from the ICE, is function of the condensation temperature, and the type of
fluid employed in the vapor-compression section. In the following section of this
chapter, on the basis of the performances study done, varying fluid and condensation
temperature, it will be showed the dependency of the principal parameters as function
of these variables.
3.2.2. Performances of working fluids varying condensation temperature
To perform the present analysis, it has been used the vapor-compression model
validated in paragraph 1.5. It has been chosen a reference refrigeration power of
1.084MWref and maintaining it constant (letting the mass flow rate of refrigerant to be
free to vary), it has been considered the savings viable using the CCHP configuration
with respect to the base case. The base case corresponds to a vapor-compression
chiller, that uses electric energy drawn by the national grid, at the price of 0.13 €/kWh,
value that consider the lasts years average prices in Italy for industry. The analysis has
Tem
pe
ratu
re [
°C]
Entropy [kJ/kg/K]
With Subcooling Without Subcooling
75
been conducted on condensation temperature in the range 20-60°C, on the base of what
reported in paragraph 1.4 and technical maps of Bitzer’s compressors [5].
The study has been considered different fluids, chosen considering data provided by
Bitzer, and other fluids commonly used in refrigeration. The fluids present in the
analysis are: R22, R134A, R404A, R407C, R407F, R507A & R717(NH3).
The analysis starts from the spreadsheet containing the vapor-compression model
validated through the data provided by Bitzer. In the software has been introduced a
modification: the external cooling of refrigerant till 10°C. Varying fluid and
condensation temperature, it is possible to have energy balance quantities, that will be
used to match the trigeneration system with the vapor-compression chiller.
For each fluid has been defined a critical condition case, used to design the CCHP’s
components. The critical condition corresponds to the maximum value of condensation
temperature, therefore 60°C.
Relatively to the sizing of the ICE, it has been decided to evaluate different options.
The 3 approaches are following reported:
- 1 (For COMP): ICE sized to meet the rated electric power of the compressor,
corresponding to the critical design conditions.
- 2 (For MA): ICE designed to meet the rated heat demand of the absorption
machine. This latter is in turn designed to provide a full sub-cooling of the
compression chiller’s refrigerant in critical design conditions.
- 3 (For % of loading): ICE size chosen as the highest among the 2 previous
approaches.
To evaluate performances of the prime mover is possible to refer to sub-section 2.2.4.
Previously, it has been anticipated that, different logics were used, and this has an
impact on the efficiencies’ calculation. In particular, the electrical efficiency for both
the modalities is calculated in the same manner. Instead, the difference is in the
calculation of the thermal efficiency. In the first approach, it is possible to evaluate the
thermal efficiency as function of the electrical power output of the ICE. For the second
approach, it is not possible, because the ICE’s electrical power is implicit, since the
76
sizing starts from the thermal power required by the absorption machine. The solution
founded is to use the ratio of thermal and electric efficiency, as function of the CHP’s
thermal power output (see Figure. 2.8), from which is possible then to evaluate the
nominal electrical power with the following equation (3.1):
(3.1) �̇�𝑒𝑙_𝐶𝐻𝑃𝑓𝑜𝑟𝑀𝐴_𝑐𝑟𝑖𝑡 =�̇�𝑖𝑛_𝑀𝐴_𝑐𝑟𝑖𝑡
𝜂_𝑡ℎ𝜂_𝑒𝑙
Regarding the other condensation temperatures (all but 60°C), the machines are
working at partial load. The modality in which the system is working has a role also
here. In fact, the design logic acts also on how the CHP follows the load. There will
be two value of partialisation, depending if, first or second logic is used. The third
logic, as in the machines’ design, will chose the highest one.
To evaluate ICE’s efficiencies at partial load, again is possible to refer to sub-section
2.2.4. After the machine’s design, it is possible, through the function described by
equation (2.5), to evaluate the electrical efficiencies at partial loads. Once the values
of efficiency at 75%, 50% and 25% of load have been calculated, it is possible to use
a linear interpolation on those value using the current load.
After efficiencies’ characterisation, it is possible to use basic energy balances to
calculate all the quantities relevant for the analysis, for example: thermal power input
to CHP, electrical power produced and useful thermal power. To evaluate the avoided
cost respect to the base case, some hypothesises are required. The electrical energy
generated can be exactly the one needed, or over/under-produced, depending on the
logic used. If approach 2 is used, a price of 0.13 €/kWh, it is adopted for the excess of
electric energy produced by the cogenerator, assuming a self-consumption within the
same plant.
Regarding the consume of natural gas, to evaluate its rate, and the relative expenses,
Lower Heating Value (LHV) and specific cost are needed. The values implemented in
the software are: 35000 kJ/Sm3 and 0.3 €/Sm3, on the basis of average values of the
Italian market. For the calculation of the primary energy saving, a reference electric
77
efficiency of the Italian electric production park of 0.46 has been used, in agreement
with the contents of paragraph 2.5.
Among the 3 approaches, it has been decided to develop the analysis on the third one.
This choice has been done because it will be possible to exploit better the capacity of
the cogenerator, and to have higher saving in term of primary energy and costs.
Below are presented the results of the analysis. Charts are used to highlights the trend
of different key indicators varying the condensation temperature.
In the first chart, Figure. 3.11, is studied, for each fluid, the fraction of refrigeration
power represented by the sub-cooling; the plot highlights the magnitude of thermal
power, that can be extracted after condensation. Increasing the condensation
temperature obviously the outlet temperature of the condenser increase, which in turn
increase the thermal power that has to be extracted to sub-cool the fluid. As showed
clearly from the chart, the lines have a linear trend, because fixing the refrigeration
capacity, the sub-cooling power is linearly dependent from condensation temperature.
The fluids for which, it is possible to exploit more heat coming from the cogenerator,
are R404A and R507A. In fact, around 30-40°C of condensation temperature (usual
working condition for vapor-compression refrigerator), it can be reached a sub-cooling
power in the range of 15-25% of the relative refrigeration power of the vapor-
compression cycle. The fluid that lead to the least recovery of heat is ammonia (NH3
or R717). Although it reaches a very high temperature after the compressor, the
circulating mass of refrigerant is low, bringing to less power to be extracted to reach
the pre-fixed temperature of sub-cooling.
78
Figure. 3.11: Heat exchanged for sub-cooling (from condensing temperature to 10°C)
with respect to refrigeration power, as function of Tcond
In this second chart, Figure. 3.12, is presented the trend of relative variation of COP.
Increasing the condensation temperature, maintaining the refrigeration power fixed,
the COP of the vapor-compression system will inevitably decrease, because it will be
required more power from the compressor to bring the fluid at a higher-pressure level
(and so temperature). At the same temperature instead, the COP of the vapor-
compression machine in hybrid configuration assumes always a higher value. In fact,
providing sub-cooling to the refrigerant, it is possible to reduce the mass flow rate,
having a saving on the power required by compressor. The variation of COP is low for
low condensation temperature since, the power, that can be extracted by sub-cooling,
is limited; it is more and more marked as the condensation temperature increase, and
additional sub-cooling can be performed. The most promising fluids remain R404A
and R507A, for which, in the critical conditions case, the COP is almost doubled,
because as seen in the previous chart, there is the opportunity to sub-cool more.
0,00
0,10
0,20
0,30
0,40
0,50
0,60
10 20 30 40 50 60 70
Qsc
/Qre
frig
Tcond[°C]
R404A
NH3
R407F
R134A
R407C
R22
R507A
79
Figure. 3.12: Variation of COP with respect to base case as function of Tcond
In the third plot, Figure. 3.13, is showed the magnitude of electrical power of CHP
compared to the part that feeds the compressor. The continuous lines represent the
trends under approach 3, instead dashed ones under approach 2. Looking to the
continuous line is possible to notice on which approach the CHP has been designed,
and on which logic the system at partial load is working. Observing the last point of
each line, the one in critical conditions, it is clear that, only for R22 and NH3 the CHP
has been sized on the base of approach 1; since, the power produced by CHP, and
consumed by vapor-compression cycle coincide. It is also possible to observe that, up
to a condensation temperature of about 45°C, all the fluids make the system working
on the first logic. At 45°C, the fluids R404A and R507A start working with second
logic, so the cogenerator will follow the thermal load imposed by the absorption
refrigerator. It occurs because, only after this point, the heat needed by the absorber is
high enough to make the CHP working at a higher use of the capacity with respect to
first logic. For other fluids R134A, R407F, R407C the logic switch is delayed at higher
condensation temperatures.
0%
10%
20%
30%
40%
50%
60%
70%
80%
90%
100%
10 20 30 40 50 60 70
CO
P v
aria
tio
n[%
]
Tcond[°C]
R404A
NH3
R407F
R134A
R407C
R22
R507A
80
Figure. 3.13: Ratio electrical power produced by CHP and power required by compressor as function
of Tcond (continuous and dashed lines respectively refer to approach 3 and 2)
Figure. 3.14 depicts the fraction of thermal power exceeding the demand of the
absorption machine for different condensation temperatures.
Figure. 3.14: Thermal power dissipated (or available for other uses) with respect to the one available
from CHP as function of Tcond
0,00
0,20
0,40
0,60
0,80
1,00
1,20
1,40
1,60
1,80
10 20 30 40 50 60 70
Pe
l,CH
P/P
el,c
om
p
Tcond[°C]
R404A
NH3
R407F
R134A
R407C
R22
R507A
0,00
0,10
0,20
0,30
0,40
0,50
0,60
0,70
0,80
0,90
1,00
10 20 30 40 50 60 70
Qd
issi
pat
ed
/Pth
_CH
P
Tcond[°C]
R404A
NH3
R407F
R134A
R407C
R22
R507A
81
Obviously, all these lines start from a high value, because when the vapor-compression
chiller is condensing at 20°C, the absorption machine is inevitably working at very
low load, requiring less thermal power. All fluids show a decreasing trend, in particular
for R404A and R507A the thermal power dissipated goes to zero after 45°C, because
as said before, the system is working with the second logic (For MA). For ammonia,
the thermal power dissipated remain an important share of CHP’s thermal power; in
fact, as highlighted before in Figure. 3.11, it is the fluid that allow to sub-cool less,
always making the absorption chiller working in a low-load condition. Must to be
considered that, usually, in an industrial environment are present processes that
requires heat in different forms, or anyway, there could be the need of heating
buildings or possible offices, that could be a solution to employ the thermal power
excess.
In the last two plots, Figure. 3.15 and Figure. 3.16, are presented the trend of savings,
respectively, in term of costs and primary energy. The configuration generates savings
in both terms. It is a combination of two factors, firstly the introduction of the
absorption machine allows to have higher COP (has to be considered that, the cost to
reduce the compression work, is introducing thermal power in the absorption chiller);
secondly, because it is less expensive produce in a combined way, electric and thermal
energy, buying natural gas, than buying directly electric energy from the national grid.
82
Figure. 3.15: Avoided cost with respect to base case as function of Tcond (0.13 €/kWh, LHVNG of 35000
kJ/Sm3 and a NG price of 0.3 €/Sm3)
It can be seen that in the usual range of operating condensation temperature, 30-40°C,
the best fluids are R407C, R407F, R404A and R507A. For higher condensation
temperature, the saving has a steep increase due to the fact that, to follow the thermal
load, the CHP produces electric energy in excess, that is accounted as avoided
withdrawal for other uses in the plant. Regarding the primary energy saving, as
avoided costs, increases with condensation temperature. In fact, for the highest
condensation temperature, the logic switches, inducing the CHP to produce more and
more electrical energy, that is self-consumed, and considered as avoided from
conventional production.
00
10
20
30
40
50
60
70
80
90
100
10 20 30 40 50 60 70
Avo
ide
d c
ost
w/r
/t B
C [
€/h
/MW
fr]
Tcond[°C]
R404A
NH3
R407F
R134A
R407C
R22
R507A
83
Figure. 3.16: Primary energy saving as function of Tcond (reference efficiency of the Italian electric
production park of 0.46, LHVNG of 35000 kJ/Sm3)
After this detailed analysis, it seems possible to conclude that, this technology has
potential to improve refrigeration in traditional vapor-compression chiller. To prove
the convenience of this new hybrid configuration, it is necessary to simulate the
operation under a real case yearly load, to figure out if, the savings produced under a
variable load, can overcome the investment cost.
0,00
0,20
0,40
0,60
0,80
1,00
1,20
1,40
1,60
10 20 30 40 50 60 70
Pri
mar
y e
ne
rgy
savi
ng
[MW
pe
/MW
fr]
Tcond[°C]
R404A
NH3
R407F
R134A
R407C
R22
R507A
84
4. CASE STUDIES SIMULATION
In this chapter, the simulation of case studies will be presented. The real load data was
provided by Impar Impianti s.r.l. [32], a firm which deals with realization of
refrigeration systems for the food industry in northern Italy. As will be highlighted in
the chapter, case I data load is characterized by high variability and low level of
system’s usage. Besides Impar case study, it has been simulated also a completer and
more continuous load, to evaluate the convenience of hybrid CCHP, and make a wider
study with different system sizes.
4.1. Introduction to real data simulation
4.1.1. Data management
Load data have been provided as CSV files. Each document contains many days of
sampling, approximately every minute, about many measurable variables of the
machine (e.g. electrical consumption, pressures). To make the simulation working
faster and to reduce the large amount of data, it has been decided to use average values
on pre-fixed temporal steps. The variability of sampling has been analysed concluding
that, averages on a temporal step of 15 minutes is a good approximation, that not lead
to an excessive smoothing of data, as Figure. 4.1 shows.
85
Figure. 4.1: Comparison of different time-step data averaging for electrical current absorbed
In Table. 4.1, is showed an extract of those data. The columns extracted, containing
useful quantities for simulation purposes, are: hour and date of sampling, electric
current absorbed, measured in Ampere, and condensing and evaporating pressures,
measured in Bar gauge.
Day Step I[A] Pcond[barg] Peva[barg]
15/08/2017 10:00:00 10.7 14.8 0.9
15/08/2017 10:15:00 86.0 15.0 3.1
15/08/2017 10:30:00 10.7 14.8 1.4
15/08/2017 10:45:00 66.2 15.0 2.5
15/08/2017 11:00:00 18.3 14.8 1.8
15/08/2017 11:15:00 80.3 15.1 2.5
15/08/2017 11:30:00 18.1 14.8 1.9
15/08/2017 11:45:00 41.0 14.8 2.6
15/08/2017 12:00:00 140.5 15.3 3.1
Table. 4.1: 2 hours averaged data extract
From pressures, using the Excel’s add-in RefProp [6], is possible to calculate
condensing and evaporating temperatures.
0
50
100
150
200
250
17:00:00 18:00:00 19:00:00 20:00:00 21:00:00 22:00:00 23:00:00 00:00:00
I[A
]
Measurement 1 hour 1/2 hour 1/4 hour
86
The cooling load is an implicit quantity, that can be calculated through the electrical
current consumed by the cooling implant. As additional data provided, it was known
that, as usual in an industrial environment, the connection is three phases to chained
voltage, so 400 V, with a frequency, being the installation on Italian territory, of 50Hz.
Through this information and equation (4.1), where I represent the electrical current,
is possible to calculate the electrical power absorbed in kWel.
(4.1) �̇�𝑒𝑙 =√3∗400∗𝐼
1000
To have the load requested by the user, it was necessary to introduce an approximation.
Looking to the trend of the condensation and evaporation pressures over the year, it is
possible to see that, considering these quantities constant is a reasonable
approximation. As an example, is reported in the next page, in Figure. 4.2, the trend
of condensation temperature during the year. The negative picks are not related to
changes in this quantity, but to the lack of data of some minutes, that inserted into
RefProp’s functions lead to not significant outputs.
After this approximation, the cooling load requested by the refrigeration cells can be
easily calculated employing the COP of the traditional vapor-compression chiller
installed, that remains constant during the year. It occurs for the way in which the
simulation model has been built, since the variation of the capacity is accommodated
by changes in the mass flow rate of refrigerant. In Figure. 4.3, it is reported the cooling
load trend extracted by those data. As it is possible to see, the capacity has a high
variability and the chiller was not employed in 2 months of summer.
Figure. 4.2: Annual trend of vapor-compression chiller’s condensing temperature, case I
Figure. 4.3: Annual trend of cooling load requested by the user, case I
0
5
10
15
20
25
30
35
01/01/2017 01/04/2017 30/06/2017 28/09/2017 27/12/2017
Tco
nd
[°C
]
0
50
100
150
200
250
300
350
400
450
500
01/01/2017 01/04/2017 30/06/2017 28/09/2017 27/12/2017
Qco
olin
g[kW
]
87
88
4.1.2. Yearly simulation model
The analysis of the yearly load has been carried out using the same principle
highlighted in sub-section 3.2.2. Since the amount of data is very large, running the
vapor-compression chiller model, validated on the base of Bitzer’s data in paragraph
1.5, for each average value, is not a smart choice. Due to this problem it has been
necessary to introduce interpolation maps, for relevant quantities of vapor-
compression chiller, based on the variation of load and on condensation temperature.
These maps are built every time that input quantities are changed and maintained
constant during simulations. Values present in those maps are directly taken from
vapor-compression simulation model. Through a Visual Basic algorithm load and
condensation pressure are changed and introduced in the vapor-compression model;
then, outputs are reported in the tables, relative to the maps. Maps examples are
provided in Figure. 4.4, Figure. 4.5 and Figure. 4.6. In the first chart, it is depicted the
trend of the sub-cooling power provided by the absorption machine. In this example,
a part of the surface is flat, because the sub-cooling required overcomes the nominal
cooling power of the machine, 165kWref.
Figure. 4.4: Sub-cooling power interpolation map
20%
60%
100%
0
50
100
150
200
20 25 30 35 40 45 50 5560
Load
[%]
Sub
-co
olin
g[kW
]
Condensation temperature[°C]
89
If, as in this case, the cooling power of the vapor-absorption chiller in not enough to
sub-cool the refrigerant till 10°C, it is considered; the refrigerant in fact, will be cooled
down till a temperature higher than the threshold pre-fixed; the consequent
improvements will be less important.
Figure. 4.5: Compressor’s electrical power interpolation map
Figure. 4.6: COP interpolation map
0
200
400
600
800
1000
1200
1400
1600
10 20 30 40 50 60 70
Pe
l,co
mp
[kW
]
Tcond[°C]
100%
80%
60%
40%
20%
0,00
1,00
2,00
3,00
4,00
5,00
6,00
10 20 30 40 50 60 70
CO
P[k
W]
Tcond[°C]
100%
80%
60%
40%
20%
90
Then from these maps, through the use of XonGrid’s [31] function “Interp2dTab”, it
was possible to interpolate the real data of cooling load and condensation temperature,
having as a result the relevant quantities for the calculation as electrical power
consumed in both cases new and base one, and sub-cooling possible with the
absorption machine employed.
Another modification introduced was about the approaches. Besides approaches 1 (For
COMP), 2 (For MA), 3 (For % of loading), a fourth approach, for the purposes of the
present study, was introduced:
- 4 (Manual): it allows to select manually the nominal CHP’s electrical power
to find an optimal size for the technical-economic analysis.
In the annual simulation, it has been decided to develop the energy balance and
economic analysis on the base of two different configurations:
- continuous regulation: in which the vapor-compression chiller is able to follow
continuously the load thanks to the presence of an inverter;
- step regulation: the capacity of the refrigerator can be varied on a fixed number
of steps. To follow in a better way the load a thermal storage is adopted.
In the real case, not being present an inverter, the continuous modulation is not
possible. It has been decided to present also the continuous regulation alternative, to
make a comparison with the discrete regulation.
Regarding the discrete regulation, it has been introduced the possibility to vary the
number of discrete step available on the nominal power. In the simulations that will be
presented a 5-step regulation of the capacity has been used. Concerning the thermal
storage, a maximum temperature of -5°C has been fixed; it represents the threshold,
that if exceeded, will bring to the activation of the vapor-compression chiller, and
consequently, to the other devices being part of the hybrid CCHP. Regarding the type
of fluid employed and its characteristic, information referring to literature have been
91
used, showed in Table. 4.2. However, these variables can be modified by the user
before the simulation.
Thermal storage parameters (glycol)
m[kg] cp[kJ/kg/K] Tmax[°C]
20000 2.51 -5
Table. 4.2: Thermal storage parameters defined by user
After this premise, on how the data were managed and the simulation carried out, it is
possible to introduce the case studies simulation.
4.2. Results of Impar’s data simulation
The working fluid employed is R507A, the nominal power of the vapor-compression
chiller is 520kWref, able to cover all the picks of refrigeration power requested by the
user, and finally, the cooling power on which, the CCHP’s machines are dimensioned,
according to the different approaches is 343kWref, an annual average value of the
fluctuating load.
4.2.1. Size analysis
On the yearly data provided by Impar Impianti s.r.l. [32] a simulation employing
approach 3 has been performed, then to provide a complete overview on the
implementation of technology, using approach 4, so the manual design, has been
performed a size analysis, to see how economic indicators vary changing the power of
the cogenerator installed.
A description of the economic analysis will introduce the discussion of the results
obtained. The analysis has been performed over 15 years, starting the year after
installation. Revenues are constituted by savings that, the installation of the CCHP will
generate, with respect to the base case, constituted by the vapor-compression
92
refrigeration system alone. The investment costs are relative to the cogenerator and
absorption chiller and are calculated using functions of the machine’s size, provided
by Casati and Chinetti [21]. To complete the investments cost is considered the
Balance of Plant (BOP), as the 30% of cogenerator and absorption chiller price, to
consider auxiliaries, installation and connection costs. Among the auxiliaries is present
a cooling tower, that is needed since a lithium bromide absorption chiller has been
employed. The overall investment is depreciated over 10 years. Then Operating and
Maintenance costs (O&M) have been considered, including:
- insurance calculated as 1.5% of the total CAPEX;
- CHP management and maintenance computed as function of the size and the
relative operating hour;
- CHP extraordinary maintenance reserve as function of same input quantities
as previous point;
- absorption chiller maintenance calculated with respect to the size of the
machine.
For the first year, during the installation, the expenses are constituted by the overall
investment cost. The subsequent years, in which the hybrid CCHP is working, variable
O&M costs constitute the expenses to be sustained. These costs are accounted with an
inflation rate of 1%. Regarding the taxation, the following formula has been used (4.2):
(4.2) 𝑇𝑎𝑥𝑎𝑡𝑖𝑜𝑛 = (𝑅𝑒𝑣𝑒𝑛𝑢𝑒𝑠 − 𝐸𝑥𝑝𝑒𝑛𝑠𝑒𝑠 − 𝐴𝑚𝑜𝑟𝑡𝑖𝑧𝑎𝑡𝑖𝑜𝑛) ∗ 𝑇𝑎𝑥 𝑅𝑎𝑡𝑒
Taxation is considered only, if the content of the parenthesis is positive. For
amortization, constant share criterion has been applied, therefore the investment cost
is equally divided in the 10 years and the tax rate is 27%.
Besides revenues due to savings, another income is present; it is related to incentives,
in particular to earnings from CB (“Certificati Bianchi’’) for which calculation is
possible to refer to paragraph 2.5. Both modalities of access to the incentive
mechanism can be applied for the present situation. Therefore, the two approaches
were analysed to find out which one, for this configuration, allow to receive the
93
greatest financial support. In Table. 4.3, it is reported an example of how was carried
out the comparison among the CB and CB type II relative to CAR.
Quantity CB type II - CAR CB
RISP/RL [kWh] 647'937 326'307
N° TEE [toe] 55.72 28.06
Incentive [€] 17'252 8'688
Table. 4.3: Comparison of CB calculation approaches
As it is possible to see, the CB type II, in which the control volume is placed on the
cogenerator, is the most convenient approach and for this it was the approach
introduced in the software.
The cash flows can easily be calculated employing the sum of earnings (from savings
and CB), costs and taxation (obviously the last two are negative). Finally, the project
balance has been built summing for each year, the discounted cash flow present at the
year before, to the cash flow of the current year. A discount rate of 4.5% has been
used.
The indicators that have been chosen to describe the economic analysis are: NPV (Net
Present Value), IRR (Internal Rate of Return) and PBP (Pay Back Period). Regarding
the PBP, it will be presented as discounted, and in a simpler form, calculated as in
formula (4.3):
(4.3) 𝑃𝐵𝑃𝑠𝑖𝑚𝑝𝑙𝑒 =𝐶𝐴𝑃𝐸𝑋
𝑆𝑎𝑣𝑖𝑛𝑔𝑠−𝐴𝑛𝑛𝑢𝑎𝑙 𝑒𝑥𝑝𝑒𝑛𝑠𝑒𝑠
All those results can be presented together because as it is possible to see in the
following graphs, with this particular load, the configuration appears instantaneously
not an interesting investment to be performed.
The Net Present Value, Figure. 4.7, define the value in the present of expected cash
flows. For a CHP power of 96kWel, founded with approach 3, the value assumed by
the NPV is about -100’000€, very far from a profitable investment (NPV positive).
94
This occurs because, although the hybrid CCHP configuration generates a saving of
about 20’000€ each year, it is not enough to repay the huge investment cost. If the
hybrid CCHP system could work more during the year, the saving generated will be
more consistent, with the possibility to overcome the CAPEX. However, the load data
is not a variable in this analysis, being the real load requested by the user. To make the
investment economically convenient, it has been performed a size analysis, to find out
if reducing the dimension of cogenerator and absorption chiller, and so, their
investment costs, is possible to recover the expenses. For this purpose, approach 4 has
been introduced, as highlighted in sub-section 4.1.2, that allows to select manually the
CHP’s electrical power output. As the size of the machines decreases, the NPV raises,
Figure. 4.7, and, for very small machines, is possible to have it positive, meaning a
profitable investment. Although is possible to have a Net Present Value not negative,
its maximum value is about 15’000€; not a huge return on an investment not negligible.
Figure. 4.7: NPV and IRR of hybrid CCHP, case I
The Internal Rate of Return, whose trend is still depicted in Figure. 4.7, is the discount
rate that makes the NPV of a series of cash flows equal to zero, and it gives a measure
of the resilience of the problem. Reducing the dimension of the machines installed, it
-10%
-08%
-06%
-04%
-02%
00%
02%
04%
06%
08%
-€ 100'000
-€ 80'000
-€ 60'000
-€ 40'000
-€ 20'000
€ '0
€ 20'000
€ 40'000
€ 60'000
€ 80'000
0 20 40 60 80 100 120IR
R[%
]
NP
V[€
]
Pel,CHP[kW]
NPV IRR
95
is possible to have an IRR positive, although the investment could generate profit, the
values assumed by IRR are not so high, indicating a low robustness of the project.
The PBP for most of the sizes, gives values that exceed the period of analysis of 15
years, for this reason the discounted PBP trend is interrupted, Figure. 4.8. When the
size is small is possible to have PBP lower than 15 years, but still too long, making the
project not very reliable.
Figure. 4.8: Pay Back Period of hybrid CCHP, case I
4.2.2. Maximum NPV case
In this sub-section, results of the case that maximize the NPV will be presented in a
more detailed way. To have a return on the investment, as seen in previous section, it
is necessary to reduce a lot the dimensions of the machines, which in turn reduce the
investment costs.
Using approach 4, manually setting CHP’s electrical power production to 20kWel, it
is possible to have the maximum of NPV, 13’000€. In this case, the absorption chiller
assumes a nominal refrigeration power of 26kWref. In Table. 4.4 are presented
investment costs related to this configuration.
00
02
04
06
08
10
12
14
16
18
20
0 20 40 60 80 100 120
Ye
ars
Pel,CHP[kW]
Simple PBP Discounted PBP
96
Investment Size [kW] Cost Specific cost [€/kW]
Cogenerator 20 € 40'000 2112
Absorption machine 26 € 22'000 842
BOP (30%)
€ 19'000
CAPEX
€ 81'000
Table. 4.4: Investment cost breakdown, maximum NPV configuration case I
Being the nominal power of the machines very low, the specific costs will increase,
leading to an important value of the final CAPEX. In Table. 4.5, it is possible to see
the values of economic indicators and other parameters obtained as a result. As
highlighted in the general overview, also in the best-case, results do not indicate an
interesting investment, showing low NPV, IRR near to discount rate and very high
period to recover the costs.
Results
Simple PBP [Years] 9.4
Discounted PBP [Years] 11.0
NPV € 13'000
IRR 7.1%
Thermal energy dissipated [kWh] 20'500
Equivalent hours vapor-compression machine 2'005
Equivalent hours vapor-absorption machine 5'003
Equivalent hours cogenerator 5'517
Table. 4.5: Case I maximum NPV configuration results
In Table. 4.6, the cash flow breakdown is reported; instead, in Figure. 4.9, it is possible
to visualize the yearly cash flows and the recovering process of the investment made
in the first year.
Cash flows
Energy saving € 11'755
TEE income € 2'425
O&M costs € 3'100
Total CF € 11'080
Table. 4.6: Cash flow breakdown, maximum NPV configuration case I
97
Figure. 4.9: Project balance diagram, maximum NPV configuration case I
In the next pages, will be presented the balance sheets relative to the two types of
regulation studied: continuous Table. 4.7 and Step Table. 4.8. The results presented up
to now are relative to discrete regulation, since, as said before, is the solution present
in the real case. In the balance sheets is present, for each month, the energy relative to
each machine and, below the annual values, the number of equivalent hour is reported.
As it is possible to see the operation of the vapor-compression chiller is very
discontinuous; instead, having reduced to very small size both the cogenerator and the
absorption chiller, makes them work for a reasonable number of hours throughout the
year. The values present in the tables shows that, there is not so much difference in the
regulation mode, in fact, both are configuration save about 12’000€, that is 19% of the
operational costs sustained in the base cases.
-€ 100'000
-€ 80'000
-€ 60'000
-€ 40'000
-€ 20'000
€ '0
€ 20'000
€ 40'000
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16
Pro
ject
Bal
ance
, CF
Years
Table. 4.7: Continuous regulation’s balance sheet, maximum NPV configuration case I
Continuous regulation
Month Qfr [kWh] Qsc
[kWh] Pel_CHP
[kWh] Pth_CHP
[kWh] Qdissipated
[kWh]
NG cost [€]
EE acquired
BC [kWh]
EE acquired
[kWh]
EE excess [kWh]
Primary energy saving [kWh]
Costs BC [€]
Costs [€]
Savings [€]
January 92'121 15'512 13'672 25'538 3'750 1'357 30'519 11'844 150 9'941 5'454 3'771 1'683
February 95'938 14'673 12'596 23'520 2'847 1'250 31'974 14'675 188 9'834 5'714 4'241 1'473
March 94'720 15'316 12'495 23'766 2'228 1'255 31'810 14'997 818 10'411 5'685 4'280 1'405
April 52'172 13'154 10'287 20'321 2'008 1'060 17'476 4'673 1'919 8'371 3'123 1'919 1'204
May 11'459 5'462 3'580 7'769 272 393 3'762 117 1'757 3'380 672 329 343 June 1'156 784 490 1'085 8 55 381 2 376 496 68 36 32
July - - - - - - - - - - - - -
August 63'422 8'787 6'972 13'199 791 698 22'343 12'972 685 6'314 3'993 3'316 677
September 244'593 17'494 13'895 25'704 762 1'371 82'012 62'043 58 13'492 14'656 14'050 606
October 169'951 18'405 14'807 27'285 1'035 1'457 58'369 37'254 2 14'131 10'431 9'072 1'359 November 125'208 16'387 13'670 25'197 1'926 1'346 42'403 23'183 0 11'968 7'578 6'084 1'493
December 98'874 16'046 14'059 26'231 3'663 1'395 32'758 13'499 129 10'436 5'854 4'148 1'706
Annual 1'049'614 142'021 116'525 219'616 19'290 11'637 353'807 195'259 6'083 98'774 63'228 51'245 11'983
h eq 2'018 5'517 5'826
98
Step regulation
Month Qfr [kWh] Qsc
[kWh] Pel_CHP
[kWh] Pth_CHP
[kWh] Qdissipated
[kWh]
NG cost [€]
EE acquired
BC [kWh]
EE acquired
[kWh]
EE excess [kWh]
Primary energy saving [kWh]
Costs BC [€]
Costs [€]
Savings [€]
January 92'118 14'364 12'915 23'794 3'452 1'271 30'247 12'597 40 9'494 5'406 3'827 1'579
February 95'992 13'770 12'070 22'237 2'724 1'188 31'792 15'155 20 9'430 5'682 4'264 1'418
March 94'848 13'373 11'585 21'343 2'412 1'140 31'630 15'572 - 9'276 5'653 4'302 1'351 April 52'390 9'117 8'390 15'457 2'639 826 17'383 5'944 - 5'727 3'107 2'032 1'074
May 13'494 2'608 2'580 4'753 1'115 254 4'390 953 - 1'424 785 447 337
June 4'264 767 820 1'511 447 81 1'326 264 - 350 237 134 103
July 780 141 150 276 80 15 243 49 - 66 43 25 19
August 63'362 7'470 6'260 11'533 927 616 22'307 13'436 - 5'431 3'987 3'344 643 September 231'088 17'239 13'770 25'369 740 1'355 81'702 61'927 40 13'389 14'601 13'926 675
October 169'962 18'435 14'770 27'212 865 1'453 58'049 37'009 48 14'282 10'374 8'965 1'409
November 125'190 16'368 13'570 25'001 1'685 1'335 42'123 23'057 49 12'133 7'528 6'014 1'514
December 98'878 15'140 13'455 24'789 3'414 1'324 32'548 14'081 - 10'093 5'817 4'183 1'634
Annual 1'042'366 128'793 110'335 203'276 20'500 10'857 353'742 200'043 196 91'094 63'218 51'463 11'755
h eq 2'005 5'003 5'517
Table. 4.8: Step regulation’s balance sheet, maximum NPV configuration case I
99
100
4.3. Results of case II data simulation
In second case study, it has been analysed a more complete load; for a user that requires
more refrigeration power and in a more continuous way during the year. The working
fluid employed is the same of the previous analysis R507A, the nominal power of the
vapor-compression machine is 1.8MWref and finally, the refrigeration power on which
logics are applied to calculate the size of the cogenerator and absorption chiller is
600kWref.
4.3.1. Size analysis
The analysis has been repeated in the same way in which it was performed for the
previous case study. It has been found that, the continuous operation of the system
leads to higher saving with respect to the base case. The significant savings allow a
faster recovery of the investment, although it is much more consistent. As Figure. 4.10
shows, hybrid CCHP on this load could generate interesting NPV, up to 390’000€. In
the rage of electrical power among 100-200kWel, the Net Present Value overcomes
350’000€. Obviously, all the results are concord, since the NPV in widely far from
zero also the IRR shows an interesting trend, still depicted in Figure. 4.10, as is known
for the Internal Rate of Return, the higher the better, indicating the robustness of the
project. The IRR reaches its maximum, 26%, employing a CHP of about 100kWel,
anyway, for all the sizes analysed, the IRR is above 15%, showing good reliability of
the project. Despite the NPV and IRR are consistent, highlighting the convenience of
the hybrid configuration, they show optimum points for different cogenerator’s size.
101
Figure. 4.10: NPV and IRR of hybrid CCHP with case II data (continuous lines and dots respectively
calculated with approaches 4 and 3)
The dots in the diagram represent NPV and IRR calculated using approach 3, based
on the average annual cooling power requested, 600kWref. As it is possible to see,
approach 4, represented by points on continuous lines, gives better results for the same
CHP’s size, showing higher NPV and IRR. The reason of this fact has to be searched
in the method on which the absorption chiller is designed. Since the size of the
absorption chiller, with approach 4, is chosen to exploit the cogenerator’s thermal
power output, the nominal cooling power will be higher, than if it would be designed
with approach 3, as Figure. 4.11 shows; allowing to sub-cool more, and so to generate
higher savings, during periods in which the CCHP system is working to provide a
cooling power higher than 600kWref. The higher savings achieved, are not balanced by
the increase in the absorption chiller’s investment cost, in fact, as depicted by Figure.
4.11, the total CAPEX are similar, resulting in a substantial difference on the NPV.
Concerning the Pay Back Period, both forms are plotted in Figure. 4.12. The years
required, to repay the investment made, are always lower than the period on which the
economic analysis have been conducted. Under PBP point of view, the best
configuration achieves a discounted PBP lower than 4 years.
0%
5%
10%
15%
20%
25%
30%
€ '0
€ 50'000
€ 100'000
€ 150'000
€ 200'000
€ 250'000
€ 300'000
€ 350'000
€ 400'000
€ 450'000
0 50 100 150 200 250 300
IRR
[%]
NP
V[€
]
Pel,CHP[kW]
NPV IRR
102
Figure. 4.11: CAPEX of hybrid CCHP and nominal absorption chiller’s cooling power with case II
data (continuous lines and dots respectively calculated with approach 4 and 3)
Figure. 4.12: Pay Back Period of hybrid CCHP with case II data (continuous lines and dots
respectively calculated with approach 4 and 3)
0
50
100
150
200
250
€ '0
€ 50'000
€ 100'000
€ 150'000
€ 200'000
€ 250'000
€ 300'000
€ 350'000
€ 400'000
€ 450'000
0 50 100 150 200 250 300
Qfr
_N,M
A[k
W]
CA
PEX
[€]
Pel,CHP[kW]
CAPEX Qfr_N MA [kW]
0
1
2
3
4
5
6
7
0 50 100 150 200 250 300
Ye
ars
Pel,CHP[kW]
Simple PBP Discounted PBP
103
4.3.2. Maximum NPV case
Since in the size investigation is emerged that NPV and IRR show their maximum
values for different CHP’s electrical power, another choice could be to consider the
size that maximize the Internal Rate of Return; in this section instead, it has been
chosen to highlight the case of highest NPV to be consistent with what done in first
case study. The present configuration employs a cogenerator of electrical power output
170kWel and an absorption chiller of refrigerating nominal power of 165kWref, whose
investment costs are reported in Table. 4.9.
Table. 4.9: Investment cost breakdown, maximum NPV configuration case II
The specific costs are lower than the previous case, because the sizes of the machines
are bigger; the total CAPEX is obviously higher, more than 4 times, the one of first
case study. Although the investment is considerably larger, in this case the project
generates interest at first sight. Following, in Table. 4.10, are presented the economic
indexes chosen to conduct the analysis.
Results
Simple PBP [Years] 4.3
Discounted PBP [Years] 4.6
NPV € 390'000
IRR 21.5%
Thermal energy dissipated [kWh] 313'469
Equivalent hours vapor-compression machine 2'560
Equivalent hours vapor-absorption machine 5'066
Equivalent hours cogenerator 5'859
Table. 4.10: Case II maximum NPV configuration results
Investment Size [kW] Cost Specific cost [€/kW]
Cogenerator 170 € 190'000 1138
Absorption machine 165 € 64'000 388
BOP (30%)
€ 76'000
CAPEX
€ 330'000
104
Although the case of optimum NPV, as highlighted before, does not maximize the IRR
and PBP, they show interesting values since, having the most profitable size, the IRR
still indicates a robust project and the discounted PBP is a third of the period of
economic analysis (15 years).
In Table. 4.11, the cash flow breakdown is reported. In Figure. 4.13, it is presented the
recovery process of the investment made to set up the hybrid CCHP system, that after
the installation, lead to cash flows due to savings compared with the base case.
Cash flows
Energy saving € 101'642
TEE income € 17'252
O&M costs € 25'000
Total CF € 93'894
Table. 4.11: Cash flow breakdown, maximum NPV configuration case II
Figure. 4.13: Project balance diagram, maximum NPV configuration case II
As in the previous case study, in the next pages, will be presented the balance sheets
of the continuous and step regulation modalities, respectively in Table. 4.12 and
Table. 4.13. The over-sizing of the vapor-compression chiller, with respect to the
-€ 400'000
-€ 200'000
€ '0
€ 200'000
€ 400'000
€ 600'000
€ 800'000
€ 1000'000
1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16
Pro
ject
Bal
ance
, CF
Years
105
average annual load, is necessary to follow the picks of September and October. Due
to this fact, also in this case, the vapor-compression system’s equivalent hours are not
so high. In the last column, it is possible to evaluate the saving produced each year,
about 100’000€, that represent the 38% of the base case’s annual costs. In addition, in
Figure. 4.14 and Figure. 4.15, a typical operating week of respectively cogenerator and
thermal storage are presented. On these charts, it is possible to visualize the trends of
electrical and thermal power produced, together with the share dissipated; regarding
the thermal storage, it is depicted the trend of glycol’s temperature. It is possible to
visualize that the cogenerator works only on two steps, because its nominal electrical
power is lower than the one requested by the compressor in nominal conditions.
Continuous regulation
Month Qfr [kWh] Qsc
[kWh] Pel_CHP [kWh]
Pth_CHP [kWh]
Qdissipated [kWh]
NG cost [€]
EE acquired BC [kWh]
EE acquired
[kWh]
EE excess [kWh]
Primary energy savings [kWh]
Costs BC [€]
Costs [€]
Savings [€]
January 324'614 63'687 80'971 122'535 36'565 7'136 107'547 5'550 0 40'285 18'485 8'752 9'733
February 336'743 64'311 79'567 117'473 30'230 6'909 112'245 11'395 0 45'994 19'292 10'227 9'066
March 330'035 63'064 73'982 108'952 25'391 6'415 110'823 16'352 0 45'792 19'048 11'175 7'873
April 183'299 36'969 41'071 64'997 21'524 3'719 61'388 9'454 0 17'053 10'551 6'472 4'080
May 324'701 63'704 80'994 122'573 36'580 7'138 107'576 5'550 0 40'291 18'490 8'754 9'736
June 314'554 61'701 78'482 118'724 35'410 6'915 104'204 5'349 0 39'081 17'910 8'472 9'438
July 324'748 63'714 81'007 122'586 36'578 7'139 107'592 5'550 0 40'305 18'493 8'755 9'738
August 322'986 63'363 80'547 121'886 36'357 7'098 107'009 5'545 0 40'091 18'392 8'713 9'680
September 766'460 99'740 108'253 151'514 11'847 9'109 271'227 128'025 1 95'322 46'618 46'379 239
October 594'790 104'353 118'457 165'173 18'837 9'946 204'278 49'854 6 97'797 35'111 24'458 10'652
November 438'227 84'172 101'314 144'123 27'051 8'606 148'409 18'386 0 71'415 25'508 13'959 11'549
December 346'068 67'435 85'133 127'687 36'341 7'463 114'655 7'245 0 44'721 19'707 9'572 10'135
Annual 4'607'225 836'213 1'009'777 1'488'222 352'711 87'594 1'556'956 268'256 6 618'147 267'605 165'687 101'918
h eq 2'560 5'068 5'940
Table. 4.12: Continuous regulation’s balance sheet, maximum NPV configuration case II
106
Step regulation
Month Qfr [kWh] Qsc
[kWh] Pel_CHP
[kWh] Pth_CHP
[kWh] Qdissipated
[kWh]
NG cost [€]
EE acquired BC [kWh]
EE acquired
[kWh]
EE excess [kWh]
Primary energy savings [kWh]
Costs BC [€]
Costs [€]
Savings [€]
January 324'540 63'692 79'142 119'654 32'141 6'971 107'536 7'089 0 43'375 18'483 9'000 9'483
February 336'690 64'276 78'009 115'698 26'893 6'792 112'210 12'616 0 48'130 19'286 10'404 8'882
March 330'120 62'968 75'145 111'032 23'867 6'528 110'865 14'416 0 48'639 19'055 10'656 8'400
April 183'240 36'964 46'405 72'290 21'874 4'162 61'376 2'465 0 22'015 10'549 4'868 5'681
May 324'720 63'735 79'214 119'757 32'187 6'977 107'583 7'051 0 43'396 18'491 8'996 9'495
June 314'550 61'703 76'737 116'115 31'354 6'763 104'207 6'833 0 41'844 17'911 8'719 9'192
July 324'720 63'823 79'377 120'040 32'359 6'993 107'560 6'838 0 43'367 18'487 8'950 9'537
August 323'010 63'278 78'712 119'252 32'350 6'942 107'023 7'147 0 42'707 18'395 8'988 9'407
September 766'440 99'733 107'501 150'686 10'372 9'054 271'220 128'644 0 96'201 46'617 45'883 734
October 594'810 104'372 115'542 162'150 15'644 9'737 204'303 52'786 0 98'397 35'115 24'850 10'266
November 438'210 84'055 97'410 139'853 22'692 8'320 148'405 22'326 0 72'318 25'508 14'711 10'796
December 346'050 67'191 82'845 124'261 31'736 7'262 114'653 9'341 0 47'550 19'706 9'937 9'770
Annual 4'607'100 835'790 996'040 1'470'787 313'469 86'501 1'556'942 277'552 0 647'937 267'603 165'961 101'642
h eq 2'560 5'066 5'859
Table. 4.13: Step regulation’s balance sheet, maximum NPV configuration case II
107
Figure. 4.14: CHP’s typical operational week, step regulation
Figure. 4.15: Thermal storage’s typical operational week, step regulation
108
109
5. CONCLUSIONS
The aim of this work was to investigate a new configuration of hybrid CCHP, in order
to estimate investment profitability. The present study was performed through the
development of a model that simulates the vapor-compression system. In order to
perform techno-economic feasibility analysis, two case studies were simulated.
In the first phase (paragraph 1.5), the vapor-compression model was validated through
the comparison with data provided by Bitzer [4], about their compressors. Once the
model was validated, it was possible to proper design the hybrid CCHP system. On
the bases of a market study and Casati and Chinetti’s [21] study the components of the
system were chosen. ICE was chosen as prime mover, since the market of CHP and
CCHP is mainly dominated by it, for many reasons: it is the most cheap, mature and
reliable technology. Regarding the absorption chiller, H2O-LiBr has been chosen as
fluid pair, since, with respect to the other fluid pair commercially available, has many
advantages, including: lower costs, good reliability, better performances and, last but
not the least, the possibility to exploit also a low temperature heat source as the ICE’s
cooling fluids.
In the second phase (sections 2.2.4 and 3.2.2), a model for the hybrid CCHP system
was built. For this reason, ICE and absorption chiller technical and economic data
collected by Casati and Chinetti [21] were used to evaluate the performance and the
costs of the machines, adding information about the operation at partial load
conditions. Using this model, the nominal performance of the new system proposed
was investigated. The analysis was carried out varying condensation temperature and
type of refrigerant. The most important reason of this analysis was to figure out the
magnitude of the possible sub-cooling. The best results were obtained with R507A and
R404A, that for ordinary condensation temperatures, showed that the heat rejection for
sub-cooling account for 20-30% of the nominal refrigeration power. An additional
110
output of this research was the benefits of the system on the COP of the vapor-
compression section and the saving achieved in terms of costs and primary energy. At
condensation temperature of 35°C, R507A and R404A showed the highest PES,
respectively 23% and 22%; instead, the worst refrigerant is ammonia that lead to a
slightly negative PES.
In the third phase (chapter 4), on the base of the previous model, a software to test the
performance of the system under a variable real load provided by Impar Impianti s.r.l.
[32] was built. The calculation time was optimized through the use of maps to allow
the simulation of a wide amount of data; in fact, each 15 minutes a time step was
simulated. The software is addressed to users that want to perform a simplified
feasibility analysis, starting from few input data. Initially 3 CHP’s design approaches
were created, following the electrical load, the thermal power requested by the
absorption chiller, and finally, the logic that selects, among the previous two, the one
that maximizes the CHP’s size. Primarily simulation of data, provided by Impar s.r.l.
[32], was performed. Due to the characteristic of this load, the hybrid CCHP system
designed with approach 3 resulted as not economically feasible.
It was decided to introduce a fourth approach to insert manually the machine’s size, in
order to perform an investigation and figure out if, by reducing the CHP and absorption
machine sizes, a recover of the investment was possible. The NPV resulted positive
only for very small CHP, around 20kWel. With this configuration is possible to save
each year about 12’000€. The investment generates an NPV of 13’000€, IRR of 7,1%
and 11 years of PBP. The cogenerator has annual PES of 21%
The second case simulated has a higher average cooling power required and works in
a more continuous way during the year. Repeating the same analysis done for the first
case study, the outputs highlighted the feasibility of the investment. Moreover, also in
this case, a size analysis was performed, showing that with approach 4 compared to
approach 3 the hybrid CCHP system shows better results, with higher NPV and IRR
at same CHP’s electrical power output. The reason of this fact has to be searched in
the method on which the absorption chiller is designed. Since the size of the absorption
chiller, with approach 4, is chosen to exploit the cogenerator’s thermal power output,
the nominal cooling power will be higher.
111
The last task performed was to study the configuration that generates the highest NPV.
With a nominal power of the vapor-compression system of 1.8MWref, the best case
installs a CHP and an absorption chiller of respectively 170 kWel and 165 kWref, with
an overall investment cost of 330’000€. The investment generates an NPV of
390’000€, IRR of 21.5% and a PBP of about 5 years. The CHP has an annual PES of
19% and considering a step regulation of the vapor-compression chiller, the yearly
saving achievable is more than 100’000€. Looking to those outputs, they show that
this hybrid system results to be effective if the load allows an important integration of
the sub-cooling, with consequent savings.
To perform these detailed analysis, software with increasing levels of features and
complexity have been developed, starting from a model to represent the vapor-
compression section, up to a complete model on the overall hybrid CCHP system. The
final software provides a profitability analysis having as input the annual cooling
demand. There’s the possibility to regulate the vapor-compression chiller by discrete
steps, but also in a continuous way. In addition to the three machines’ design
modalities of model’s first version a forth manual logic to perform size analysis was
introduced. The output of the software are annual balance sheets, charts and detailed
economic analysis.
Future development of the present work could comprehend, first of all, tests on
additional case studies, with different load trends and refrigerant fluids, in order to
verify the effective economic validity of the system presented. In particular, additional
simulations could regard real cases in which heat, cooling and electricity are needed
for other processes.
112
LIST OF FIGURES
Figure. 1.1: Technical scheme and T-s diagram of gas cycle refrigeration ............... 6
Figure. 1.2: Thermodynamic diagrams of ideal Stirling cycle: P-V diagram and T-s
diagram [3] .................................................................................................................. 7
Figure. 1.3: Single-stage vapor-compression system scheme and its T-s diagram ..... 8
Figure. 1.4: Reciprocating open compressor ............................................................ 11
Figure. 1.5: Compression phases of a reciprocating compressor ............................. 11
Figure. 1.6: Bitzer’s semi-hermetic compressor [4] ................................................. 12
Figure. 1.7: Rotating piston compressor ................................................................... 13
Figure. 1.8: Sliding-vane compressor’s section ........................................................ 14
Figure. 1.9: Bitzer’s ECH209Y semi-hermetic scroll compressor ............................ 14
Figure. 1.10: Working principle of a scroll compressor ........................................... 15
Figure. 1.11: Screw compressor ................................................................................ 16
Figure. 1.12: Centrifugal compressor’s impeller ...................................................... 17
Figure. 1.13: Overall power absorbed by the refrigerator group (expressed in Horse
Power) function of condenser’s pressure for different regulation mode of the fans [12]
................................................................................................................................... 26
Figure. 1.14: Application limit of a Screw compressor from Bitzer’s software [5] .. 27
Figure. 1.15: Homepage Bitzer’s software ................................................................ 28
Figure. 1.16: Output of Bitzer’s software in the case of single screw compressor ... 29
Figure. 1.17: Scheme of the vapor-compression cycle highlighting the most important
points .......................................................................................................................... 32
Figure. 2.1: Cogeneration conceptual scheme .......................................................... 35
113
Figure. 2.2: Trigeneration conceptual scheme .......................................................... 36
Figure. 2.3: Simplified heat recovery process from an ICE ...................................... 38
Figure. 2.4: Trigeneration using a gas turbine as a prime mover [17] .................... 39
Figure. 2.5: Specific cost comparison between ICE and MTG technologies for small
size CHP [21] ............................................................................................................ 43
Figure. 2.6: ICE’s electrical efficiency as function of electrical power .................... 44
Figure. 2.7: ICE’s thermal efficiency as function of electrical power ...................... 45
Figure. 2.8: ICE’s ratio thermal-electrical efficiency as function of thermal power 45
Figure. 2.9: ICE’s total efficiency at different loads as function of electrical power 46
Figure. 2.10: ICE’s electrical efficiency at partial load with respect to nominal one as
function of electrical power ....................................................................................... 47
Figure. 2.11: Single-pressure absorption refrigerator scheme and its P-T diagram 48
Figure. 2.12: Absorption chiller’s COP varying fluid and refrigeration capacity ... 50
Figure. 2.13: Specific cost with respect to cooling capacity for NH3-H2O and H2O-
LiBr pairs [21] ........................................................................................................... 54
Figure. 3.1: Configurations of the two concepts DECO & EVA ............................... 64
Figure. 3.2: Pressure-Temperature diagram of EVA & DECO concept respectively64
Figure. 3.3: Configuration of a traditional absorption refrigeration system (left)
compared with an absorption-compression hybrid one (right) ................................. 66
Figure. 3.4: P-T diagram of an absorption-compression hybrid refrigeration system
................................................................................................................................... 67
Figure. 3.5: Effects of generation temperature on generation heat, comparison among
conventional and hybrid system ................................................................................. 68
Figure. 3.6: Effects of generation temperature on cooling capacity, comparison among
conventional and hybrid systems ............................................................................... 69
Figure. 3.7: Arrangement of a solar absorption-subcooled compression hybrid cooling
system ......................................................................................................................... 70
Figure. 3.8: PBP and NPV as function of price of collector ..................................... 71
Figure. 3.9: Hybrid CCHP conceptual scheme ......................................................... 73
Figure. 3.10: Temperature-entropy diagram comparing a traditional vapor-
compression cycle and one with sub-cooling at condenser’s outlet .......................... 74
114
Figure. 3.11: Heat exchanged for sub-cooling (from condensing temperature to 10°C)
with respect to refrigeration power, as function of Tcond ........................................... 78
Figure. 3.12: Variation of COP with respect to base case as function of Tcond ......... 79
Figure. 3.13: Ratio electrical power produced by CHP and power required by
compressor as function of Tcond (continuous and dashed lines respectively refer to
approach 3 and 2) ...................................................................................................... 80
Figure. 3.14: Thermal power dissipated (or available for other uses) with respect to
the one available from CHP as function of Tcond ....................................................... 80
Figure. 3.15: Avoided cost with respect to base case as function of Tcond (0.13 €/kWh,
LHVNG of 35000 kJ/Sm3 and a NG price of 0.3 €/Sm3) .............................................. 82
Figure. 3.16: Primary energy saving as function of Tcond (reference efficiency of the
Italian electric production park of 0.46, LHVNG of 35000 kJ/Sm3) ........................... 83
Figure. 4.1: Comparison of different time-step data averaging for electrical current
absorbed ..................................................................................................................... 85
Figure. 4.2: Annual trend of vapor-compression chiller’s condensing temperature,
case I .......................................................................................................................... 87
Figure. 4.3: Annual trend of cooling load requested by the user, case I .................. 87
Figure. 4.4: Sub-cooling power interpolation map ................................................... 88
Figure. 4.5: Compressor’s electrical power interpolation map ................................ 89
Figure. 4.6: COP interpolation map ......................................................................... 89
Figure. 4.7: NPV and IRR of hybrid CCHP, case I ................................................... 94
Figure. 4.8: Pay Back Period of hybrid CCHP, case I ............................................. 95
Figure. 4.9: Project balance diagram, maximum NPV configuration case I ............ 97
Figure. 4.10: NPV and IRR of hybrid CCHP with case II data (continuous lines and
dots respectively calculated with approaches 4 and 3) ........................................... 101
Figure. 4.11: CAPEX of hybrid CCHP and nominal absorption chiller’s cooling power
with case II data (continuous lines and dots respectively calculated with approach 4
and 3) ....................................................................................................................... 102
Figure. 4.12: Pay Back Period of hybrid CCHP with case II data (continuous lines
and dots respectively calculated with approach 4 and 3) ....................................... 102
Figure. 4.13: Project balance diagram, maximum NPV configuration case II ....... 104
115
Figure. 4.14: CHP’s typical operational week, step regulation .............................. 108
Figure. 4.15: Thermal storage’s typical operational week, step regulation ........... 108
116
LIST OF TABLES
Table. 1.1: Data input of Bitzer’s software ............................................................... 29
Table. 1.2: Data output of Bitzer’s software ............................................................. 30
Table. 1.3: Simulation of cycle properties example ................................................... 31
Table. 1.4: Model results’ check ................................................................................ 32
Table. 1.5: Calibrated efficiencies ............................................................................. 33
Table. 2.1: Reference efficiencies for PES’s calculation........................................... 57
Table. 2.2: Grid efficiency for PES’s calculation ...................................................... 57
Table. 2.3: Ambient temperature’s corrective factor for PES’s calculation ............. 57
Table. 4.1: 2 hours averaged data extract ................................................................. 85
Table. 4.2: Thermal storage parameters defined by user .......................................... 91
Table. 4.3: Comparison of CB calculation approaches ............................................ 93
Table. 4.4: Investment cost breakdown, maximum NPV configuration case I .......... 96
Table. 4.5: Case I maximum NPV configuration results ........................................... 96
Table. 4.6: Cash flow breakdown, maximum NPV configuration case I ................... 96
Table. 4.7: Continuous regulation’s balance sheet, maximum NPV configuration case
I .................................................................................................................................. 98
Table. 4.8: Step regulation’s balance sheet, maximum NPV configuration case I ... 99
Table. 4.9: Investment cost breakdown, maximum NPV configuration case II ....... 103
Table. 4.10: Case II maximum NPV configuration results ...................................... 103
Table. 4.11: Cash flow breakdown, maximum NPV configuration case II ............. 104
Table. 4.12: Continuous regulation’s balance sheet, maximum NPV configuration
case II ....................................................................................................................... 106
117
Table. 4.13: Step regulation’s balance sheet, maximum NPV configuration case II
................................................................................................................................. 107
118
ABBREVIATIONS
CHP Combined Heat and Power
CCHP Combined Cooling, Heat and Power
ICE Internal Combustion Engine
MGT Micro Gas Turbine
FC Fuel Cell
O&M Operations and Maintenance
CAR Cogenerazione ad Alto Rendimento
TEE Titoli di Efficienza Energetica
CB Certificati Bianchi
NPV Net Present Value
IRR Internal Rate of Return
PBT Pay Back Time
PES Primary Energy Saving
VGV Variable Guide Vane
VFD Variable Frequency Drive
HPC Head Pressure Control
CTC Condensing Temperature Control
SOFC Solid Oxide Fuel Cells
PEFC Polymer Electrolyte Fuel Cells
PEMFC Proton Exchange Membrane Fuel Cells
AFC Alkaline Fuel Cells
PAFC Phosphoric Acid Fuel Cells
MCFC Molten Carbonate Fuel Cells
119
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