Anno Accademico 2017 - 2018
UNIVERSITA’ DI PISA
DIPARTIMENTO DI INGEGNERIA CIVILE E INDUSTRIALE
Corso di Laurea Magistrale in Ingegneria Chimica
TESI DI LAUREA MAGISTRALE
Refrigeration loops in LNG applications:
Dynamic simulation and optimization study
Relatori: Candidato:
Prof. Ing. Gabriele Pannocchia Giulia Pauli
Dr. Ing. Marco Pelella
Dr. Ing. Lorenzo Gallinelli
Controrelatore:
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Sommario
1. Introduction ... 9
1.1. Aim of the thesis ... 9
1.2. Outline ... 9
2. Introduction to refrigeration cycles ... 11
2.1. Thermodynamics of refrigeration cycles ... 11
2.1.1. Reversed Carnot cycle ... 13
2.1.2. Coefficient Of Performance (COP) ... 14
2.1.3. Vapour refrigeration cycles ... 16
2.1.4. Gas refrigeration cycles ... 24
2.2. Most common types of refrigerants ... 25
2.2.1. Hydrocarbon refrigerants ... 29
2.2.2. Inorganic refrigerants ... 31
2.3. Summary ... 33
3. Refrigeration cycles applied in the liquefaction of natural gas ... 34
3.1. Cascade cycles ... 35
3.1.1. Phillips Optimized cascade process ... 35
3.1.2. Statoil/Linde Mixed Fluid Cascade (MFC) process ... 37
3.2. Mixed Refrigerant cycles ... 39
3.2.1. Single Mixed Refrigerant (SMR) cycles ... 40
3.2.2. Dual Mixed Refrigerant (DMR) cycles ... 47
3.3. Gas-expander cycles ... 51
3.4. Process selection criteria ... 53
3.5. Summary ... 56
4. The case study and reference steady state simulations... 58
4.1. Case study ... 58
4.2. Steady state simulations ... 62
4.2.1. Case 1 - Saturated vapour after intercooler ... 63
4.2.2. Case 2 - Partially liquefied vapour after intercooler ... 65
4.2.3. Case 3 - Separated pressurized hot refrigerant streams ... 67
4.2.4. Case 4 - Medium pressure liquid refrigerant used to precool natural gas ... 69
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5. Steady state optimization ... 75
5.1. Sensitivity analysis ... 75
5.1.1. Mixed refrigerant flow-rate ... 75
5.1.2. Pressure ... 75
5.1.3. Temperature after intercooler and condenser ... 82
5.1.4. Sub-cooled refrigerant temperature after chilling ... 82
5.1.5. Mixed refrigerant composition ... 83
5.2. Numerical optimization ... 92
5.2.1. Introduction to numerical optimization ... 92
5.2.2. Optimization with Honeywell SQP optimizer ... 95
5.2.3. Optimization summary ... 102
6. Dynamic simulations... 105
6.1. Dynamic model ... 106
6.1.1. Loop without cold box ... 106
6.1.1.1. Loop Control ... 110
6.1.2. Cold Box ... 115
6.1.2.1. Cold Box control scheme ... 122
6.1.3. Complete loop ... 123
6.2. Simulated events ... 125
6.2.1. Lower pressure natural gas feed ... 125
6.2.2. Higher pressure natural gas feed ... 128
6.2.3. Lower cooling efficiency of the condenser ... 130
6.2.4. Summary ... 132
7. Conclusions ... 133
7.1. Overview of the work ... 133
7.2. Future work ... 134
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List of Figures
Figure 2. 1 : T-S diagram for an ideal Reversed Carnot cycle. ... 14
Figure 2. 2 : The COP of a reversible refrigerator as a function of TL (TH is taken as 298 K). ... 16
Figure 2. 3 : The COP of a reversible refrigerator as a function of TH (TL is taken as 298 K). ... 16
Figure 2. 4 : Schematic for ideal vapour-compression refrigeration cycle. ... 18
Figure 2. 5 : T-S and P-h diagrams for the ideal vapour-compression refrigeration cycle. ... 18
Figure 2. 6 : Single-stage propane refrigeration system. ... 19
Figure 2. 7 : Simple Joule-Thomson liquefaction cycle. ... 20
Figure 2. 8 : T-s diagram for an actual vapour-compression refrigeration cycle. ... 21
Figure 2. 9 : Refrigeration cycle with multistage compression and expansion. ... 21
Figure 2. 10 : Cascade Refrigeration Cycle. ... 22
Figure 2. 11 : Ammonia absorption refrigeration cycle. ... 23
Figure 2. 12 : Schematic and T-s diagram of an ideal closed reverse Brayton cycle. ... 24
Figure 2. 13 : Operating ranges of some common refrigerants. ... 29
Figure 3. 1 : Phillips Optimized Cascade Liquefaction Process. ... 36
Figure 3. 2 : T-Q profile of the Phillips optimized cascade LNG process. ... 37
Figure 3. 3 : Statoil/Linde Mixed Fluid Cascade (MFC) process. ... 38
Figure 3. 4 :Temperature-Heat diagram using pure or mixed refrigerants. ... 40
Figure 3. 5 : AP-SMRTM refrigerant process for natural gas liquefaction. ... 41
Figure 3. 6 : Propane-precooling mixed refrigerant cycle for natural gas liquefaction. ... 43
Figure 3. 7 : AP-XTM cycle for natural gas liquefaction. ... 44
Figure 3. 8 : Black & Veatch Pritchard PRICO process. ... 46
Figure 3. 9 : AP-DMRTM LNG process for natural gas liquefaction. ... 48
Figure 3. 10 : Shell dual mixed refrigerant process for natural gas liquefaction. ... 50
Figure 3. 11 : IFP/Axens LiquefinTM process. ... 51
Figure 3. 12 : Turbo-expander refrigeration cycle for gas liquefaction. ... 52
Figure 3. 13 : Double Turbo-expander refrigeration cycle for gas liquefaction. ... 53
Figure 4. 1 : Basic process scheme ... 58
Figure 4. 2 : Reference process configuration – Separated case 1... 60
Figure 4. 3 : Heat flow – temperature diagram related to the Separated case 1 ... 64
Figure 4. 4 : Heat flow – temperature diagram related to the Mixed case 1 ... 65
Figure 4. 5 : Process configuration – Mixed case 1 ... 66
Figure 4. 6 : Process configuration – Separated case 3 ... 68
Figure 4. 7 : Process configuration – Mixed case 3 ... 70
Figure 4. 8 : Process configuration – Separated case 4 ... 71
Figure 4. 9 : Process configuration – Mixed case 4 ... 73
Figure 5. 1: Trend of COP and of the total process power requirements, with refrigerant pressure value after expansion in Joule-Thomson valve. ... 76
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Figure 5. 2 : Trend of the two compressors power requirements with refrigerant pressure value after
expansion in Joule-Thomson valve. ... 76
Figure 5. 3 : Trend of the power requirement of the pump with refrigerant pressure value after expansion in Joule-Thomson valve. ... 77
Figure 5. 4 : Trend of the COP and of the total process power requirements, with refrigerant pressure value after the first stage of compression. ... 78
Figure 5. 5 : Trend of the two compressors power requirements with refrigerant pressure value after the first stage of compression. ... 78
Figure 5. 6 : Trend of the power requirement of the pump with refrigerant pressure value after the first stage of compression. ... 79
Figure 5. 7 : Trend of COP and of the total process power requirements, with refrigerant pressure value after the second stage of compression. ... 80
Figure 5. 8 : Trend of the two compressors power requirements with refrigerant pressure value after the second stage of compression. ... 80
Figure 5. 9 : Trend of the power requirement of the pump with refrigerant pressure value after the second stage of compression. ... 81
Figure 5. 10 : Trend of COP vs. Methane molar fraction ... 84
Figure 5. 11 : Trend of COP vs. Methane molar fraction ... 85
Figure 5. 12 : Trend of COP vs. Ethane molar fraction ... 86
Figure 5. 13 : Heat flow – temperature diagram for the Reference case without propane ... 87
Figure 5. 14 : Heat flow – temperature diagram related to Case 4 ( ) ... 87
Figure 5. 15 : Trend of COP vs. Propane molar fraction ... 88
Figure 5. 16 : Trend of COP vs. Isobutane molar fraction ... 89
Figure 5. 17 : Trend of COP vs. Normal butane molar fraction ... 90
Figure 5. 18 : Trend of COP vs. Isopentane molar fraction ... 91
Figure 5. 19 : Heat flow –Temperature diagram in the cold box after optimization ... 104
Figure 6. 1 : Compressor map with Surge Control Limit Line. ... 114
Figure 6. 2: Anti-surge system scheme ... 114
Figure 6. 3 : Dynamic Cold Box divided into fifteen zones ... 119
Figure 6. 4 : Dynamic Cold Box with eleven sets in a single zone ... 120
Figure 6. 5 : Temperature profiles in the LNG Heat Exchanger divided in fifteen zones ... 120
Figure 6. 6 : Temperature profiles in the LNG Heat Exchanger with eleven repeating sets ... 121
Figure 6. 7 : Cold Box control scheme ... 123
Figure 6. 8 : Complete liquefaction loop dynamic model ... 124
Figure 6. 9 : Natural gas feed pressure ramp (1) ... 126
Figure 6. 10 : LNG and mixed refrigerant molar flow rate (1) ... 127
Figure 6. 11 : Surge margin 1 and actuator position related to anti-surge valve 1 and LNG-valve (1) ... 127
Figure 6. 12 : LNG temperature trend (1) ... 128
Figure 6. 13 : Natural gas feed pressure ramp (2) ... 129
Figure 6. 14 : LNG and mixed refrigerant molar flow rate (2) ... 129
Figure 6. 15 : Surge margin 1 and actuator position related to surge-valve 1 and LNG-valve (2) ... 130
Figure 6. 16 : LNG temperature trend (2) ... 130
Figure 6. 17 : Condenser product temperature trend ... 131
Figure 6. 18 : LNG and mixed refrigerant molar flow rate (3) ... 131
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List of Tables
Table 4. 1 : Mixed refrigerant and natural gas compositions ... 59
Table 4. 2 : Natural gas inlet and outlet conditions ... 61
Table 4. 3 : Compressors and pump power consumptions – Separated case 1 ... 63
Table 4. 4 : Compressors and pump power consumptions – Mixed case 1 ... 64
Table 4. 5 : Compressors and pump power consumptions – Separated case 2 ... 65
Table 4. 6 : Compressors and pump power consumptions – Mixed case 2 ... 67
Table 4. 7 : Compressors and pump power consumptions – Separated case 3 ... 67
Table 4. 8 : Compressors and pump power consumptions – Mixed case 3 ... 69
Table 4. 9 : Compressors and pump power consumptions – Separated case 4 ... 72
Table 4. 10 : Compressors and pump power consumptions – Mixed case 4 ... 72
Table 4. 11 : Summary of the results associated with the alternative configurations considered ... 74
Table 5. 1 : Reference process features ... 83
Table 5. 2 : Case studies for the couple of components Methane-Nitrogen (C1-N2). ... 84
Table 5. 3 : Case studies for the couple of components Methane-Ethane (C1-C2). ... 85
Table 5. 4 : Case studies for the couple of components Ethane-Propane (C2-C3). ... 86
Table 5. 5 : Case studies for the couple of components Propane-i-Butane (C3-iC4). ... 88
Table 5. 6 : Case studies for the couple of components i-Butane -n-Butane (iC4-nC4). ... 89
Table 5. 7 : Case studies for the couple of components n-Butane-i-Pentane (nC4-iC5). ... 90
Table 5. 8 : Case studies for the couple of components i-Pentane -n-Pentane (iC5-nC5). ... 91
Table 5. 9 : Process consumptions and gains definig the objective function value. ... 94
Table 5. 10 : Process variables which can be manipulated. ... 95
Table 5. 11 : Process constraints. ... 95
Table 5. 12 : Variations in variables subject to constraints ... 96
Table 5. 13: Variations in process variables ... 96
Table 5. 14 : New process features. ... 97
Table 5. 15 : Process constraints. ... 97
Table 5. 16 : Process Variables which can be manipulated. ... 98
Table 5. 17 : Variations in variables subject to constraints ... 98
Table 5. 18 : Variations in process variables ... 99
Table 5. 19 : New process features ... 99
Table 5. 20 : Process variables which can be manipulated. ... 100
Table 5. 21 : Variations in variables subject to constraints ... 100
Table 5. 22 : Variations in process variables ... 101
Table 5. 23 : New process features ... 101
Table 5. 24 : Objective function values reached after the different stages of the optimization. ... 102
Table 5. 25 : Process variables values reached after the different stages of the optimization ... 103
Table 5. 26 : Values reached by the variables subject to constraints after the different stages of the optimization... 104
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Table 6. 1 : Rating valves data ... 107
Table 6. 2 : Nominal and design pump operating data ... 108
Table 6. 3 : Vessels dimensions and verification ... 109
Table 6. 4 : Tuning of Level Controllers ... 112
Table 6. 5 : Tuning of Temperature Controllers ... 112
Table 6. 6 : Tuning of Pressure Controller ... 113
Table 6. 7 : Default zone geometry features ... 116
Table 6. 8 : Default metal properties ... 116
Table 6. 9 : Default layers configuration ... 116
Table 6. 10 : Default features of a single layer ... 116
Table 6. 11 : Steady state UA and first attempt coefficients U’ ... 118
Table 6. 12 : Joule-Thomson sizing data ... 119
Table 6. 13 : Reference coefficients U’ for U flow scaled calculations ... 122
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List of Acronyms
APCI Air Products & Chemicals Inc. C1 Methane
C2 Ethane C3 Propane
C3MR Propane precooled Mixed Refrigerant C4 Butane
C5 Pentane
COP Coefficient of Performance DMR Dual Mixed Refrigerant
FLNG technologies Floating Liquefaction of Natural Gas technologies GWP Global Warming Potential
J-T valve Joule-Thomson valve
LiMuM process Linde Multistage Mixed refrigerant process LNG Liquefied Natural Gas
MCHE Main Cryogenic Heat Exchanger MFC process Mixed Fluid Cascade process MR Mixed Refrigerant
MTPA Million Tons per Annum NGL Natural Gas Liquids ODP Ozone Depletion Potential
OSMR process Optimized Single Mixed Refrigerant process PFHE Plate-Fin Heat Exchanger
POCLP Phillips Optimized Cascade Liquefaction Process PRICO process Poly Refrigerated Integrated Cycle Operation process SMR Single Mixed Refrigerant
SQP algorithm Sequential Quadratic Programming algorithm SWHE Spiral-Wound Heat Exchanger
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To myself,
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1. Introduction
1.1. Aim of the thesis
The interest regarding refrigeration systems comes from the fact that refrigeration finds wide employment in several industrial applications and sectors, such as the petro-chemical, pharmaceutical and chemical ones. Typical applications include gas processing, separation, process fluid storage, chilling, liquefaction, chemical reaction control, and so on, making the refrigeration system an essential element in many industrial plants. Detailed knowledge of these systems then allows to obtain a more reliable and safe process plant, as well as to locate, within the refrigeration cycle, opportunities to manipulate process parameters in order to maximize process efficiency, thus reducing global costs.
The main objective of this thesis was to develop simulation models for a single mixed refrigerant (SMR) refrigeration cycle used for natural gas liquefaction, in order to find suitable operating conditions and control strategies, which allow to build a reliable process model. To this purpose, it has been first developed a steady state model of the process, using UniSim Design as the simulation software, and observed the influence of the main parameters values on the liquefaction process efficiency. This analysis has led to an optimized steady state process model, which has been then converted into a dynamic one. Finally, it has been useful to analyze the dynamic response of the model in case of some simulated accidental events. This work has been developed in cooperation with GE Oil & Gas Nuovo Pignone s.r.l. which has kindly offered its know-how regarding process simulations.
1.2. Outline
This thesis is organized as follows.
2. Introduction to refrigeration cycles.
In Chapter 2, an introduction to the basics of refrigeration cycles, including descriptions of the reference ideal cycles considered, common classifications utilized, and an overview on the principal refrigerant fluids which have been selected to date, is given.
3. Refrigeration cycles applied in the liquefaction of natural gas.
In Chapter 3, an overview on the main types of industrial refrigeration cycles is given, focusing on the most employed for the liquefaction of natural gas in the Oil & Gas industry. These cycles appear to be basically classified into three groups: cascade cycles, mixed refrigerant cycles, and gas-expander cycles.
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4. Case study and reference steady state simulations.
As reported in Chapter 4, starting from hypothetical reference operating conditions, several alternative configurations for a single mixed refrigerant process have been built up in order to identify the one which provides the highest efficiency under the fixed conditions.
5. Steady state optimization
.As described in Chapter 5, the most efficient configuration of the SMR liquefaction process, identified in the previous chapter, has been then optimized. First, sensitivity analysis has been carried out to observe the influence of the adjustable process parameters on the refrigeration loop efficiency. Then, the results have been verified and possibly confirmed through the use of the Honeywell SQP Optimizer, and optimization algorithm implemented in UniSim Design.
6. Dynamic simulations
.As reported in Chapter 6, the optimized steady state process model has been converted into a dynamic one. The compression system and cold box have been separately stabilized in dynamic mode, in order to find the dynamic specifications and control strategies which allow to match the desired boundary conditions and to close the refrigeration loop. Some accidental events have been also simulated.
7. Conclusions.
In Chapter 7, the summary of the obtained results of the work, and suggestions for improvements and future works, are given.
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2. Introduction to refrigeration cycles
Refrigeration is a process in which heat is transferred from one region to another in controlled conditions, in order to maintain or to reduce its temperature. Cyclic refrigeration consists of a refrigeration cycle which uses a refrigerant fluid to provide a required bulk of cooling. This refrigerant absorbs heat from the reservoir which is desired to keep at low temperature, and then it is recovered through compression (gas cycles) and condensation, or being absorbed by another liquid medium, for vapour cycles. The two major concepts in providing the refrigeration are direct and indirect. In this latter option the product itself is used as the refrigerant fluid.
Refrigeration finds wide employment in several industrial applications and sectors, such as the Oil & Gas, refinery, petro-chemical, chemical and pharmaceutical industries. In fact, refrigeration is an integral part of many facilities, in different plant sections, since it permits to liquefy and separate mixture components, control physical or chemical processes, and much more applications, ensuring the required low temperature operating conditions. Typical applications includes dew point control, humidity control for hygroscopic chemicals, heavy components separation, mixtures fractionation, vapour recovery from loading, storage (e.g. boil-off gas) and unloading processes, recovery of solvents, cooling for exothermic chemical reactions, condensation, solidification, refrigeration facilities, process chillers, maintenance of a stored liquid at low temperature to control pressure in the containing vessel, and liquefaction of gaseous streams. This latter employment will be examined in depth in this thesis work, considering particularly the refrigeration plants for natural gas liquefaction process.
2.1. Thermodynamics of refrigeration cycles
A cycle is a series of thermodynamic processes in which the final conditions and properties of the matter which experiences them, are identical to the initial ones. In a refrigeration cycle, processes produce a cooling effect and are arranged to operate in a cyclic manner, so that the refrigerant can be reused.
The first law of thermodynamics, the law of conservation of energy, states that energy can be neither created nor destroyed, thus the net change in the total energy of a system during a process is equal to the difference between the total energy entering and the total energy leaving the system:
Where and are the total energy entering and exiting respectively, and is the difference between them.
For a closed system undergoing a process between initial (1) and final (2) states involving heat and work interactions with the surroundings, the expression becomes:
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Where , , , are the heat and work entering and exiting the closed system respectively. The difference between inlet and outlet total energy is defined by the sum of contributions associated with internal ( ), kinetic ( ) and potential ( ) energy respectively.
If changes in kinetic and potential energy are negligible, the expression is then: ( ) ( ) ( )
Where the change in internal energy is defined by the product between the closed system mass, and the difference between the specific internal energy at the final and initial states.
Considering a control volume involving a steady-flow process, with mass entering and leaving the system (associated with a certain energy – enthalpy, ) and heat and work interactions with the surroundings, the total energy content of the control volume remains constant, and thus the first law of thermodynamics can be expressed as:
̇ ̇ ̇ ̇ ̇
̇ ̇ ̇ ̇ ̇ ̇
A refrigerator is a cyclic device used to transfer heat from a low- (at TL) to a high-temperature (at TH) medium. Refrigeration cycles typically absorb heat from a low-temperature reservoir (the process stream) and continuously release it as heat to a high-temperature sink, generally an ambient temperature fluid, such as cooling water or air. According to the second law of thermodynamics, which refers to the inefficiencies of practical thermodynamic systems and states that heat cannot flow spontaneously from a cold to a warmer body without any external source of energy, a work input is required to accomplish these heat transfers. In fact, the second law of thermodynamics indicates that it is impossible to have 100% efficiency in heat to work conversion, and Clausius statement states that it is impossible to construct a cyclic device (e.g., refrigerator and heat pump), which spontaneously transfers heat from the low-temperature side (cooler) to the high-temperature side (hotter). Furthermore, the second law of thermodynamics also states that the entropy of a system increases in any heat transfer or conversion of energy within a closed system, and thus all energy transfers or conversions are irreversible. Much effort has been spent in minimizing the entropy generation (irreversibility) in thermodynamic systems and applications, so far, and to quantify the work potential of energy, a property called exergy has been defined. Exergy represents the upper limit on the amount of work a device can deliver, or the lower limit of work absorbed. It is important to note that exergy does not represent an actual amount of work and that the exergy is a property of the system–environment combination and not of the system alone, and thus it depends on the conditions of the environment as well as the properties of the system. The work potential or exergy of the kinetic or potential energy of a system is equal to the kinetic or potential energy itself since it can be converted to work entirely. On the other hand, the internal energy and enthalpy of a system are not entirely available for work, and thus only part of the thermal energy of a system can be converted into work. Exergy analysis is a useful tool for improving the efficiency of energy resource utilization since it reveals how much it is possible to design more efficient energy systems by reducing inefficiencies.
The reversibility is defined as the statement that both the system and its surroundings can be returned to their initial states. On the other hand, the irreversibility shows the destruction of availability and thus both
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the system and its surroundings cannot be returned to their initial states due to the irreversibilities occurring, such as friction, and electrical and mechanical effects.
The reversible work is defined as the maximum amount of useful work output or the minimum work input for a system undergoing a reversible process between two specified initial and final states. The difference between the reversible work ( or ) and the actual work ( or ) is due to the irreversibilities present during the process, and this difference is called irreversibility or exergy destroyed ( ):
or
Irreversibility is a positive quantity for all actual (irreversible) processes since for work-producing devices and for work-consuming devices. Since irreversibility can be viewed as the lost opportunity to do useful work, the smaller the irreversibility associated with a process, the greater the work that is produced (or the smaller the work that is consumed) and the performance of a system can be improved by minimizing the irreversibility associated with it [1].
2.1.1. Reversed Carnot cycle
The ideal Carnot cycle is a theoretical thermodynamic cycle consisting of reversible processes, two isothermal heat transfers and two adiabatic transformations, compression and expansion. A system undergoing this cycle, is a heat engine which converts part of the heat energy absorbed from an high temperature reservoir, to mechanical work. During this process, heat losses are transferred to (surroundings) a low-temperature sink. Carnot’s theorem states that considering heat engines operating between two reservoirs at fixed constant temperatures, an irreversible thermodynamic engine cannot be more efficient than a reversible one. Thus, a heat engine operating in a reversible Carnot cycle, can serve as a model for real heat engines, providing the upper limit on the thermal efficiency achieved in converting absorbed heat into useful work.
The ideal reversed Carnot cycle comprises the same processes of the Carnot cycle, but with reversed directions for any heat and work interactions. It is the most efficient refrigeration cycle operating between two reservoirs at constant temperatures, and so often used as a reference point for comparison in the design and evaluation of industrial refrigeration systems [2]. This cycle consists of two isentropic (adiabatic) and two isothermal reversible processes as illustrated in the T-S diagram (Gibbs Diagram Temperature– entropy Diagram) in Figure 2. 1. In this diagram is shown a two-phase envelope, inside of which the refrigerant is present as a vapour-liquid mixture. To the left of the two-phase envelope the refrigerant is present as liquid, and to the right, as vapour.
The refrigerant fluid absorbs heat at low constant temperature (TL) medium from point 1 to point 2, experiences ideal isentropic (reversible) compression from point 2 to point 3 (work input is required and the temperature of the refrigerant increases), rejects heat isothermally to a high temperature reservoir (at TH) along path 3 to 4, and finally expands isentropically from point 4 back to point 1.
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The first law of thermodynamics states that when energy, work, or heat, passes through a system, its internal energy changes, according to the law of conservation of energy [3]. Since the internal energy variation for a working fluid in a refrigeration loop is zero, the energy balance can be written as:
| | | | | | | |
The refrigeration provided by this ideal reversed Carnot cycle, and the required work input are determined by the following relationships:
( ) ( )
| | | | ( ) ( ) Since in the Carnot Reversed Cycle e :
( ) ( ) Turbine Compressor Evaporator Condenser Cold Medium, TL Hot Medium, TH 3 4 2 1 QH QL s T 1 2 3 4 Q H QL TH TL Work Input
Figure 2. 1 : T-S diagram for an ideal Reversed Carnot cycle.
2.1.2. Coefficient Of Performance (COP)
The performance (efficiency) of a refrigeration cycle can be evaluated through a coefficient of performance (COP), which is defined as the ratio of the desired refrigerating effect ( ), to the work spent in order to produce it:
| |
15 | |
( )
This equation gives the maximum value of COP for any refrigerator operating between two reservoirs at fixed temperature and . In fact, an ideal refrigerator operating in reversed Carnot cycle requires the minimum work to provide the desired refrigeration effect.
From this relation it is possible to infer that if is fixed, for example to the ambient temperature value, the COP increases as the temperature at which the refrigerant evaporation occurs ( ) increases [4], as shown in Figure 2. 2. At the same time, for a fixed temperature , a decrease in the condenser temperature (see Figure 2. 3) would cause a decrease in the temperature difference across the refrigeration cycle ( ), thus increasing the coefficient of performance and decreasing the power requirements for a given cooling duty [5]. In summary, the lowest possible temperature and the highest temperature should be used in order to increase the cycle COP. Furthermore, changes in the temperature of the low-temperature reservoir ( ) has more influence in the coefficient of performance than those in , so that the following relation is satisfied [3]:
|(
) | |(
) |
While the COP of an ideal refrigeration cycle depends only on the temperatures of the reservoirs considered, the COP of a real refrigeration cycle is also highly influenced by operating conditions, and particularly by temperature approaches in the heat exchangers. In fact, COP increases if the temperature difference between the working fluid and sinks decreases. Moreover, the model of compressor used and the thermodynamic properties of the specific employed working fluid may influence the refrigeration cycle performance.
The ideal reversed Carnot cycle is not actually realized, cause it requires ideal compression and expansion devices which can handle a two-phase flow. Actually, refrigerant is usually superheated before compression, and sub-cooled before expansion. Furthermore, expansion of a liquid in a turbine is impractical and thus it is achieved at constant enthalpy by a throttling valve. Moreover, in a real refrigeration cycles the compression and expansion steps do not occur isentropically. In fact, real processes comprise irreversible transformations, due to dissipative phenomena related to intermolecular frictional or viscous effects inside the system. This irreversibility increases global entropy, and thus the sum of the entropy of the system and its environment.
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Figure 2. 2 : The COP of a reversible refrigerator as a function of TL (TH is taken as 298 K).
Figure 2. 3 : The COP of a reversible refrigerator as a function of TH (TL is taken as 298 K).
2.1.3. Vapour refrigeration cycles
Generally, cyclic refrigeration systems can be classified as vapour and gas cycles. Vapour refrigeration cycles could be then divided into vapour-compression and vapour-absorption refrigeration loops. In both cases, process stream cooling is provided by the refrigerant vaporization in an evaporator.
0 3 6 9 12 15 18 220 230 240 250 260 270 280 COPr e v TL (K)
COP of a reversible refrigerator vs. TL
0 5 10 15 20 25 30 310 320 330 340 350 360 370 COPr e v TH (K)
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Vapour-compression refrigeration cycle
It is the most commonly used, and the cooling effect is reached through a refrigerant fluid which is vaporized and condensed alternately, and compressed in the vapour phase. Four main components are required to accomplish a vapour-compression refrigeration system, as follows:
Evaporator: a mixture of vapour and liquid, or liquid refrigerant absorbs its latent heat of vaporization from the process stream which has to be cooled, and gradually changes from liquid to vapour state. Evaporation is the gaseous escape of molecules from the surface of a liquid and is accomplished by the absorption of a considerable quantity of heat without any change in temperature. The evaporated gases exert a pressure called the vapour pressure. As the temperature of the liquid rises, a greater loss of the liquid from the surface occurs, which increases the vapour pressure. In the evaporator of a refrigeration system, a low-pressure cool refrigerant vapour is brought into indirect contact with the stream to be cooled, in order to absorb heat and boil producing a low-pressure saturated vapour [1].
Compressor : the low pressure vapour refrigerant is recovered and compressed, reaching a high pressure associated to a saturation temperature of the refrigerant slightly higher than the atmospheric temperature. Increasing the refrigerant pressure raises its boiling temperature above the process stream temperature.
Condenser: in the condenser the superheated vapour refrigerant rejects the latent heat previously absorbed in the evaporator and the work energy taken up during compression, usually to an ambient temperature fluid, such as water or air. The refrigerant is so de-superheated and condensed to a saturated liquid. In some practical applications, it is desired that the condenser cools the refrigerant below the condensation temperature (sub-cooling) to reduce flashing when the refrigerant pressure is reduced in the throttling device, and thus to reduce the amount of vapour entering the evaporator, improving system performance [1]. Expansion valve: this device reduce the pressure and temperature of the liquid refrigerant at
constant enthalpy, before it is recycled to the evaporator. The expansion process partially vaporizes the liquid refrigerant, reducing the cooling effect provided in the evaporator. The refrigerant is then returned to the beginning of the next cycle.
Figure 2. 4 illustrates a schematic of an ideal refrigeration system using the Reverse-Rankine vapour-compression refrigeration cycle .
18 Compressor Evaporator Condenser Cold Medium, TL Hot Medium, TH 2 3 1 4 QH QL Expansion Valve h3 = h4 Saturated liquid + vapour Saturated liquid Saturated vapour
Figure 2. 4 : Schematic for ideal vapour-compression refrigeration cycle.
The T-s and P-h diagrams for an ideal vapour-compression refrigeration cycle are shown in Figure 2. 5.
s T 4s 1 2 3 Q H QL TH TL 4 Win h P 1 2 3 QH QL TH TL Win 4
Figure 2. 5 : T-S and P-h diagrams for the ideal vapour-compression refrigeration cycle. 1-2 Reversible adiabatic isentropic compression of the refrigerant, in a compressor
2-3 Reversible heat rejection at constant pressure in a condenser 3-4 Irreversible throttling at constant enthalpy, in an expansion valve 4-1 Reversible heat absorption at constant pressure in an evaporator
In this refrigeration cycle, frictional pressure drops and heat losses to the surroundings are ignored. The irreversibility within the expansion valve could be avoided replacing it with an isentropic turbine, so that refrigerant would enter the evaporator at state 4s. This alternative provides a reduction in power requirement but is not convenient due to its higher cost and complexity.
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Basic propane refrigeration process
A typical single-stage propane refrigeration system, illustrated in Figure 2. 6, is constituted by a vapour-compression refrigeration cycle, as described above.
Condenser
Suction
Drum
Receiver
50 °C, 16 bar
Evaporator
-40 °C, 1.1 bar
Expansion Valve
Compressor
17 bar
Figure 2. 6 : Single-stage propane refrigeration system.
As shown in Figure 2. 6, propane vapour enters the compressor at 1 bar and approximately -40°C and is then compressed to 17 bar (and so warmed to about 87°C). Power required by the compressor is defined as follows:
̇ ̇ ( )
Where is the work of compression, is the adiabatic efficiency of the compressor, and and are the specific enthalpies at the end and at the beginning of compression, respectively.
The warm compressed propane vapour is then cooled to about 50°C and condensed against an ambient fluid, such as water or air. Liquid propane is collected in a receiver at 16 bar, then it flashes through an isenthalpic Joule-Thomson valve, where it is subjected to a pressure drop to 1.1 bar and also a temperature decrease to the saturation temperature of liquid propane.
Cold saturated propane liquid then enters a heat exchanger and evaporates at −40°C while absorbing latent heat of vaporization from the process stream to be cooled. Once chillers were kettle type, in which propane refrigerant boils and evaporates on the shell side while the process streams flow in an immersed tube bundle. Today high-performance chillers are used, such as plate-fin heat exchangers. Finally, propane vapour leaves the chiller, enters the suction drum and returns the compressor to undergo the refrigeration cycle again [7].
Joule-Thomson cycle
A schematic of a simple Joule-Thomson liquefaction cycle, exploiting the Joule-Thomson effect in order to liquefy a natural gas feed stream, is illustrated in Figure 2. 7.
20 Air-cooled Heat exchanger 1 3 J-T Valve Compressor 4 Heat exchanger 5 6 2 LNG product Liquid receiver Methane make-up gas
qL Thermodynamic
boundary
Figure 2. 7 : Simple Joule-Thomson liquefaction cycle.
First the natural gas feed stream is compressed, then it is cooled and expanded from 101 to 1 bar through a J-T valve. Expansion cools the natural gas to about -47°C without condensing, and thus it is completely recycled to the heat exchanger, where it provides further cooling to the high pressure natural gas, before recompression. Proceeding this way, the temperature of the expanded natural gas stream is progressively reduced, until liquid is formed and then separated in the liquid receiver. Liquefied natural gas is finally withdrawn as product from the liquid receiver, while the amount of low-pressure natural gas recycled gradually decreases, so that an increasing addition of makeup gas to the compressor is needed to compensate for the lower flow rate. When steady state is reached no further cooling can be achieved. Fixed the external heat leak , the temperature of the compressed gas and the enthalpies of the streams (LNG product, 2, and recycled gas, 3) exiting the thermodynamic boundary shown in Figure 2. 7, the only way to increase liquefaction is to decrease the inlet gas enthalpy , which is obtained by increasing the inlet pressure. An improvement can be achieved in this simple Joule-Thomson cycle through equipment addition and more complex operations. For example, double expansion provides higher efficiency and is widely used in LNG facilities [7].
Actual vapour-compression refrigeration cycle
In an actual vapour-compression refrigeration cycle, as shown in Figure 2. 8, the irreversibilities of the several components are taken into account. The refrigerant leaves the evaporator as superheated vapour (1), and the condenser as sub-cooled liquid (3), in order to prevent the compressor and expansion valve from handling a two-phase stream. Pressure drops are not neglected, thus the heat transfers do not occur at constant pressure and temperature. Also the pressure drop through the compressor suction line can be significant. Finally, the compressor irreversibility would cause an increase in entropy and thus in temperature, but it is possible to reach a lower entropy (state 2’) by using a multi-stage compressor with intercooler.
21 S T 1 2 3 4 2'
Figure 2. 8 : T-s diagram for an actual vapour-compression refrigeration cycle.
If the liquid refrigerant is sub-cooled when it enters the expansion valve, the fraction vaporized during the expansion process is less, and thus a greater proportion of liquid refrigerant provides cooling through vaporization in the evaporator (vapour provides no cooling, as it enters and leaves the evaporator at the same temperature and enthalpy). This means that for a given refrigeration duty, a lower refrigerant flow rate can be used, resulting in a decreased compression requirement. Furthermore, in some configurations it is possible to use a separator after the expansion device, in order to send only the liquid fraction to evaporator, avoiding distribution problems in the heat exchanger, while the vapour is directly recycled to the compressor [5].
To reduce the overall power requirement of the refrigeration system, it is useful to introduce multistage compression with intercooling, in order to split overall compression work, and multistage expansion, with a separator between the stages, called economizer, in order to reduce the vapour flow rates and thus the power requirement in the low-pressure compression stage. A refrigeration cycle with multistage compression and expansion is shown in Figure 2. 9 [5].
h P 1 2 4 Inter-cooler Economizer (Separator) 2 Expansion Valve Expansion Valve Condenser W 1 3 4 5 6 7 8 9 3 8 7 6 5 9
Figure 2. 9 : Refrigeration cycle with multistage compression and expansion.
Efficiency achieved by a vapour-compression refrigeration cycle, can be also increased by using a cascade refrigeration cycle, as shown in Figure 2. 10, which operates with lower temperature differences and requires lower compression duties compared to a single refrigeration cycle, using different refrigerant
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fluids in each cycle. A cascade system allows to reach higher refrigerating effects (increased refrigeration capacity) and COPs, cause the combination of more than one cycle provides process cooling at different temperature levels. Furthermore, the heat released from the high temperature refrigeration cycle during refrigerant condensation, is absorbed by the vaporizing refrigerant in the low temperature one [5]. On the other hand this configuration requires additional equipment and thus higher capital cost.
h P 1 2 4 Evaporator 2 Condenser W 1 3 4 5 6 7 8 3 8 7 6 5 W
Figure 2. 10 : Cascade Refrigeration Cycle.
( )
( ) ( )
When the combined refrigeration systems employ the same refrigerant fluid, the heat exchanger between the stages is replaced by a mixing chamber (or flash chamber) since it has better heat transfer characteristics [8].
Vapour-absorption refrigeration cycle
It is an attractive refrigeration method only when a source of inexpensive thermal energy, such as geothermal or solar energy, can be directly used, reducing the electrical energy requirement. These cycles involve the absorption of a refrigerant by a transport medium, and the most common system utilizes ammonia (NH3) as the refrigerant and water (H2O) as the relative transport medium. A schematic of an ammonia–water system is shown in Figure 2. 11.
23 Evaporator Condenser Cold Medium, TL Hot Medium, TH QH QL Pure NH3 Rectifier Generator H2O NH3+H2O NH3+H2O Absorber Expansion Valve Pump Wpump Pure NH3 Q Qgen Thermal Energy Input Expansion Valve Qcond Regenerator Cooling Water
Figure 2. 11 : Ammonia absorption refrigeration cycle.
This system differs from the vapour-compression one in that the compressor is replaced by an absorption system which comprises an absorber, a pump, a generator, a regenerator, an expansion valve, and a rectifier. Low pressure vapour ammonia leaves the evaporator and enters the absorber where it dissolves in water and reacts with it forming NH3 • H2O. This is an exothermic reaction so it is necessary to use cooling water in order to promote ammonia dissolution. The liquid NH3 - H2O solution is pumped to the regenerator where it is warmed, and then it enters the generator. Here a thermal energy source transfers heat to the solution. Thus a vapour rich in NH3 is formed and passes through a rectifier, which separates the water and returns it to the generator. Meanwhile the hot highly diluted NH3 - H2O solution passes through a regenerator, where it transfers heat to the solution leaving the pump, and is throttled to the absorber pressure before entering it. Compared with vapour-compression systems, the main advantage of this system is that the steady–flow work input required is lower, since it is proportional to the specific volume and here liquid is compressed instead of a vapour.
The work input to the pump is usually very small and negligible compared to the energy provided by the thermal source, thus the coefficient of performance for an absorption refrigeration cycle can be defined as follows:
On the other hand, absorption refrigeration cycles are much more expensive and complex, and less efficient, thus this option is competitive only for large scale industrial installations and when the unit cost of thermal energy available is low compared to electricity.
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2.1.4. Gas refrigeration cycles
Gas Refrigeration cycles use gas as working fluid, generally air, nitrogen or methane, so that it is compressed and expanded without any change phase. Compared to vapour refrigeration cycles, gas cycles are less efficient, because the refrigerant fluid receives and rejects sensible heat, instead of latent heat at constant temperature. Thus, to reach the same refrigerating effect, they require higher mass flow rates of working fluid, and also higher work inputs to compressors. The use of a turbine to obtain work output from gas expansion, allows to reduce the work input requirements to the compressor. These cycles works on the reverse Brayton cycle instead of the reverse Rankine cycle, as shown in Figure 2. 12.
3 4 2 1 s T T C W 1 3 2 4 P = cost P = cost
Figure 2. 12 : Schematic and T-s diagram of an ideal closed reverse Brayton cycle.
Ideal Reverse Brayton cycle is composed by two adiabatic processes (isentropic compression and expansion) and two isobaric heat transfers (it’s like a Carnot cycle with isothermal processes replaced by two isobaric heat transfer processes). Low pressure gas is compressed isentropically (1-2) and the work input required by compressor is defined as follows:
̇( ) ̇ ( )
Compressed gaseous refrigerant enters a heat exchanger and rejects heat at constant pressure to a high-temperature sink (2-3). Then, the gas flows through a turbine, where it undergoes isentropic expansion (3-4) and provides a net work output defined as:
| | ̇( ) ̇ ( )
Expansion causes a decrease in working fluid temperature, and finally cold low pressure gas enters a second heat exchanger, where it absorbs heat at constant pressure from the process stream to be cooled (4-1), providing the desired refrigerating effect. Applying steady flow energy equation, heat released and absorbed by the refrigerant fluid, are respectively expressed as:
| | ̇( ) ̇ ( ) | | ̇( ) ̇ ( )
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Known that the coefficient of performance (COP) is defined as the ratio of the desired refrigerating effect ( ), to the net work spent for compression, considering the work output of the turbine, the COP of an ideal gas refrigeration cycle is defined as follows:
| | | | | | | | | | | | ( ) ( ) ( )
Reversed Brayton Cycle can be seen as a modification of a reversed Carnot Cycle operating between two reservoirs at constant temperatures and , and thus for fixed and the COP of Reverse Brayton Cycle is always lower than the COP of reverse Carnot Cycle:
( )
( ) ( ) ( )
2.2. Most common types of refrigerants
A refrigerant is a fluid that in a refrigeration system absorbs heat during evaporation at low temperature and low pressure, and rejects heat during condensation at almost ambient temperature and higher pressure. In natural gas liquefaction processes, the refrigeration capacity requirement is an important parameter in the refrigeration technology or refrigerant selection. Refrigerants can be classified in many different ways, based upon different classification criteria.
For example, according to their refrigeration capacity they can be classified into two categories. The first group includes refrigerants working at near ambient temperatures in refrigerators operating as a closed loop. Typical refrigerants belonging to this category are inorganic compounds, halocarbon compounds, and hydrocarbons. They have relatively high critical temperatures so that they can be liquefied near ambient or medium cold temperatures. The second group includes cryogenic fluids and gases such as methane, air, oxygen and nitrogen. They have very low boiling temperatures, usually below , so they enable process stream cooling to cryogenic temperatures in open loop liquefiers.
The selection of a suitable refrigerant for a specific refrigeration system is affected by many factors such as desired final process fluid temperature and throughput rate, the coefficient of performance (COP), safety requirements, cost, and the size and location of the refrigeration plant. Furthermore, other parameters that must be taken into account include the following basic thermodynamic and physical properties required for common refrigerants used in these applications [1]:
High critical temperature, so that refrigerants can be liquefied at ambient or medium low temperatures.
Evaporating pressure as low as possible [4] but still higher than atmosphere pressure, to avoid air infiltrations, and condensing pressure not too high, to reduce the compressor power requirement. Low boiling temperature, in order to avoid to operate at pressure below the atmospheric one
(vacuum operation) in the evaporator.
High volumetric refrigerating capacity in order to reduce the size of the compressor and refrigerant flow rates.
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High latent heat of evaporation, to reduce the required amount of refrigerant and consequently also the overall power requirement.
Low viscosity and density, to provide lower pressure drops.
High thermal conductivity, which reduce the heat exchange area requirement.
Non-toxicity, non-corrosiveness to the material of construction, chemical stability and compatibility with the lubricating oil employed for compressors.
Low specific volume, which reduces the size of the equipment and also provides higher compressor efficiency.
Furthermore, refrigerants should be: Ozone- and environment friendly Nonflammable and nonexplosive Easily detectable in case of leakage Low cost
Another common classification criterion for refrigerants used for refrigeration or liquefaction cycles, is based on their chemical compositions, as the following:
Hydrocarbons: refrigerants belonging to the hydrocarbon group, such as ethane, propane, butane and isobutane. They are recovered from petroleum and thus usually applied in refrigeration systems in the petrochemical industry, due to their low cost and ready availability. On the other hand, they are flammable compounds, so appropriate safety requirements must be considered to prevent fire hazards. Although they are highly flammable, they may offer advantages as alternative refrigerants because they are inexpensive to produce and have zero ozone depletion potential (ODP), very low global warming potential (GWP), and low toxicity.
Halocarbons: light hydrocarbons with one or more halogen atoms, such as chlorine, fluorine, and bromine, substituted for hydrogen atoms. This group of refrigerants is called the halogenated hydrocarbons or halocarbons, and also referred to freons. They are colourless, non-inflammable, non-corrosive to most metals and generally non-toxic. In this group, the most commonly used refrigerants are the so-called chlorofluorocarbons, CFCs, once commonly used as refrigerants, solvents, and foam-blowing agents, but their use rapidly decreased, because of their environmental impact and cause they have been found to be carcinogenic . CFCs are odorless, nontoxic, and heavier than air, as well as dangerous if not handled properly.
Inorganic Refrigerants: inorganic compounds such as ammonia (NH3), carbon dioxide (CO2), water (H2O) and air. Among these compounds, ammonia has received the greatest attention for practical applications.
Cryogenic gases: gases and gaseous mixtures used as cryogenic fluids. Refrigerants belonging to this group are nitrogen, oxygen, air, methane, helium, etc.
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Refrigerant Mixtures: mixtures of two or more pure refrigerant fluids. Mixtures allows to manipulate and tailor the composition in order to suit the temperature requirements and to control properties such as toxicity and flammability. They can be broadly divided into two categories.
Azeotropic Refrigerants: mixtures of two or more refrigerants which cannot be separated into their components by distillation. They behave as a single substance and their properties are different from those typical for their constituents. Their vapour and liquid phases retain the same compositions over a wide range of temperatures, and thus under constant pressure they evaporate and condense as a single substance at constant temperature.
Zeotropic Refrigerants: mixtures of two or more refrigerants that show different compositions for the liquid and the vapour phase at the vapour-liquid equilibrium state. In fact, they do not evaporate and condensate at constant temperatures (temperature increases at the evaporation and decreases at the condensation). Compared to azeotropic refrigerants, on equal operating conditions they show lower evaporating temperature providing higher refrigerating effects [3].
History of Refrigerants
Early mechanical refrigeration systems employed sulfur dioxide, methyl chloride and ammonia. In the late 1920s, these toxic refrigerants have been replaced by CFCs (Chlorofluorocarbons), thanks to the improved process of synthesis utilized to provide them. Since the late 1970s, increased environmental concern brought to new strict legislative regulations, which phased out the use of substances such as chlorofluorocarbons (CFCs) and hydrochlorofluorocarbons (HCFCs), because of their destructive effects on the ozone layer.
Montreal Protocol, established in 1987 and later revised, provides guidelines for individual country legislation aiming to reduce and then phase-out chlorine-containing refrigerants (because the major responsible for ozone depletion is chlorine). Thus, air conditioning and refrigeration industry introduced chlorine free refrigerants, such as Hydrofluorocarbons (HFCs), which are non-ozone-depleting, nonflammable, recyclable and energy-efficient refrigerants of low toxicity [10].
Furthermore, Kyoto Protocol, established in 1997, focused attention on the impact of human activity on climate change, introducing the need for replacement of several chemicals due to their effect on global warming (greenhouse gases, such as include water vapour, carbon dioxide, methane, nitrous oxide, and some refrigerants, build-up in the atmosphere and trap heat). Thus, CFCs and HCFCs had to be substituted, cause they usually have a high global warming potential, and the potential to cause damage to the ozone layer if released to the atmosphere [11].
Two factors that measure the environmental impact of refrigerants, are the Ozone Depletion Potential (ODP) and the Global Warming Potential (GWP). Refrigerants with an Ozone Depletion Potential equal to zero and a low Global Warming Potential (GWP), less than 150 (over 100 years) [12] (or than 4,000 [13]), are considered acceptable.
Nowadays, environmental concerns have increased interest in using natural refrigerants, such as hydrocarbons, which are regaining popularity as alternatives to synthetic fluorocarbon refrigerants. In fact,
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the best refrigerants from an environmental point of view, chosen as substitutes of CFCs and HCFCs, are hydrocarbons, ammonia and carbon dioxide.
Refrigerants designation
Refrigerants designation is generally composed by a letter R as prefix and an identifying number. For hydrocarbons, the first digit on the right (units) represents the number of fluorine atoms (F), the second digit (tens) is one more than the number of hydrogen atoms (H) and the third digit on the right (hundreds) is one less than the number of carbon atoms (C). In case of isomers, as they become more and more unsymmetrical, lowercase letters (i.e. a, b or c) are present. Inorganic compounds belong to the 700 series, and thus their identification numbers are obtained adding the relative molecular mass to 700 [14].
Refrigerants selection
When choosing the appropriate refrigerant fluid to employ in a refrigeration system, besides the factors listed above, it is necessary to focus on the temperature range considered. In particular, the freezing point of the refrigerant used must be well below the evaporator temperature. Furthermore, the latent heat of vaporization of a substance, and thus the amount of heat absorbed by a certain refrigerant flow rate, diminishes with increasing temperature and becomes zero when the critical temperature is reached (at the critical point liquid and vapour phases are indistinguishable). This can be inferred also from the shape of the two-phase envelope in a T-s diagram. Therefore, it is desirable to choose a refrigerant fluid which has a critical point well above the evaporator and condenser temperatures, in order to reduce the needed refrigerant flow-rate and consequently the power requirement. Considering the lower limit of the operating temperature range fixed to the normal boiling point (to avoid vacuum operations), and the upper limit set to ambient temperature if the critical temperature is well above it, or to a temperature corresponding with a heat of vaporization of 50% of that at atmospheric pressure, it is possible to obtain approximate operating ranges for refrigerants, as shown in Figure 2. 13. This kind of diagram allows to easily select the suitable refrigerant to be used for a particular refrigeration system, considering that the required refrigerating effect and thus the evaporator temperature, must be not too close to the upper limit of the refrigerant operating range [5].
29 60 80 100 120 140 160 180 200 220 240 260 280 300 320 Temperature (K) Ambient Temperature Power Requirement Nitrogen Methane Ethylene Ethane Propane Propylene i-Butane n-Butane Ammonia Chlorine
Figure 2. 13 : Operating ranges of some common refrigerants.
From this diagram, it is possible to infer that nitrogen and methane are selected for very low temperature applications, and that if these refrigerants, as well as ethylene or ethane, are utilized, and the heat rejection has to occur to an ambient temperature fluid, it is necessary to use a cascade system, so that heat is first transferred to another refrigerant, such as propane or propylene. The choice of this secondary refrigerant depends on the temperature difference in the heat exchanger linking the cycles, and on safety concerns.
The operating ranges of pure refrigerant fluids can be extended by using refrigerant mixtures, called mixed refrigerants, whose compositions and vaporization and condensation temperatures are not constant. In fact, they condense and evaporate over a temperature range, providing efficient cooling also in applications with high temperature variations [5].
2.2.1. Hydrocarbon refrigerants
Hydrocarbon refrigerants are naturally occurring and compatible with most lubricants (but not with silicone or silicate) and materials generally used in refrigeration systems. They have zero ozone depletion potential (ODP) and negligible global warming potential (GWP) [15]. The three most viable hydrocarbon refrigerants include propane, isobutene, and propylene, which have GWP values of 3 or less. On the other hand, hydrocarbons are highly flammable, so that they are considered unsafe in applications requiring large volumes of refrigerant [13]. Moreover, if a hydrocarbon is to be used as a replacement in a system that wasn’t originally designed for a flammable refrigerant, the equipment must be modified and associated with suitable safety systems [11].
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Methane (R-50)
Methane is used in many industrial chemical processes. Its major source is the extraction from natural gas fields and it can be transported as a refrigerated liquid (LNG). Methane can be used as a low-temperature refrigerant, being compressed and expanded in vapour-compression refrigeration cycles. Although it is highly flammable as other hydrocarbons, it offers advantages as alternative refrigerant, because it is inexpensive to produce, with zero ODP and very low GWP and toxicity [1].
Ethylene (R-1150)
Ethylene is a refrigerant suitable for very low temperature refrigeration applications, such as LNG liquefaction processes. It is non-toxic, with zero ODP and very low GWP, but it is also a flammable refrigerant and therefore not suitable for retrofitting existing fluorocarbon refrigerant systems [16]. Ethylene is usually used in the liquefaction process of ethylene itself, in a closed loop two-stage refrigeration cycle, which represents the low stage of a cascade refrigeration system (propane is used in the higher temperature stages). An open ethylene system, using the process ethylene as refrigerant, is not used in order to prevent contamination of the process ethylene by oil used in the ethylene compressor.
Propylene (R-1270)
Propylene is a refrigerant suitable for low and medium temperature refrigeration applications. It is non-toxic, with zero ODP and very low GWP (< 2), but also extremely flammable [17]. Propylene, such as propane, is a typical hydrocarbon refrigerant employed for the high stage of cascade refrigeration systems, typically working between about -30°C evaporating and +45°C condensing temperature. The higher vapour pressure of propylene, allows a 5 K lower evaporating temperature than propane. Propylene has been used, for example, in the high stages of a refrigeration system employed for the recovery of mixed gasoline and benzene vapour from air, at an oil refinery [18].
Propane (R-290)
Propane is a refrigerant suitable for low to high temperature refrigeration applications, with zero ozone depletion potential and low global warming potential (< 4). It has excellent thermodynamic properties, similar to those of ammonia, but it also has the advantage of being non-toxic. Propane is inexpensive and readily available, but its extreme flammability is its major disadvantage, which requires special safe design to avoid fire hazards. This disadvantage can be overcome by using propane as a refrigerant for the low temperature cycles [19]. Common safety measures are the use of semi-hermetic compressors and ventilation to prevent a flammable gas mixture occurring in case of refrigerant leakage. Due to these protection measures required, a conversion of an existing plant into a propane plant, is not always possible [20].
Compared to ammonia, propane has a lower boiling point, and thus it can be used for lower temperature applications. Furthermore more, it is compatible and not corrosive to copper (not suitable for ammonia systems), as well as to other common materials, such as aluminum, bronze and stainless steel [19].
Propane refrigeration systems are often required in gas processing to provide chilling and to condense the heavy components of a process gas stream, for example in NGL recovery. In this process, the natural gas stream is chilled with a propane refrigeration cycle, and then the condensed liquids are separated and stabilized in columns [19]. Furthermore, propane has a wide range of applications which includes
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commercial and industrial refrigeration, cold storage and food processing, transport refrigeration, and air conditioning systems [21]. It is also used in the high temperature stages of the refrigeration system applied for the liquefaction of ethylene [18].
Mixed refrigerant
A mixed refrigerant, is a blend of refrigerants which provide continuous cooling approaching the cooling curve of the process stream to be chilled or liquefied. Typically, these refrigerant mixtures are composed by several refrigerants such as light hydrocarbons and nitrogen. Compared to pure refrigerants, mixed refrigerants do not condense and evaporate at constant temperature but over a range of temperature. Thus, they allow to operate with low temperature difference in heat exchangers, reaching higher thermal efficiencies. In fact, adjusting the composition of these mixtures, so that refrigerants evaporate over a temperature range similar to that of the cooling curve, it is possible to maximize the efficiency of thermal exchanges, while optimizing the refrigerant flow rates and thus compressors dimensions and requirements.
2.2.2. Inorganic refrigerants
Early Oil & Gas facilities (LNG liquefaction plants [3]) were constructed in remote regions of the world and propane was generally selected as the pre‐cooling refrigerant since it was readily available and extracted from the feedstock natural gas, eliminating the need to supply refrigerant fluid. As LNG industry developed and migrated to locations with the nearby presence of chemical and petrochemical industries, plant energy efficiency became more important, both economically and environmentally. In fact, the increased focus on the impact of gas emissions on the global environment, led to an increased demand for environmental friendly refrigerants, such as inorganic refrigerants.
Nitrogen (R-728)
Nitrogen is used as in refrigerant in gaseous phase, and provides cooling after being compressed and expanded through a gas turbine, in low efficiency gas compression cycles (see section 3.1).
Carbon dioxide (R-744)
Carbon dioxide (CO2) is colourless, odourless and naturally occurring in the environment. It is heavier then air and thus can cause asphyxiation in the case of accidental loss, unless proper ventilation systems, detectors and alarms are installed [1]. It is a very attractive refrigerant, because it has zero ODP and a GWP of 1, by definition. It is inexpensive, readily available and also environmentally friendly cause it is nonflammable, nonexplosive and with low toxicity.
Carbon dioxide can be used as direct refrigerant in low efficiency transcritical operations, where it is not condensed at high pressure (the compressor outlet temperature and pressure are extremely high – above critical point (31.1 °C, 73.8 bar)). System equipment must be designed to be suitable to stand higher pressure than those typical of the other refrigeration systems, and this is costly. Carbon dioxide can also be employed in sub-critical systems, in which it operates at pressure levels much like conventional refrigeration systems. For example, in carbon dioxide/ammonia cascade systems, CO2 is expanded in order
