THERMODYNAMIC AND COMBUSTION STUDY ON BLENDS OF AND AS ALTERNATIVE

A Thesis Presented

By

Sai Chandra Yelishala

to

The Department of Mechanical and Industrial Engineering

in partial fulfillment of the requirements for the degree of

Master of Science

in the field of

Thermo-fluids, Mechanical Engineering

Northeastern University Boston, Massachusetts

August 2019

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ACKNOWLEDGMENTS

Throughout my master’s degree, I received a boundless support and assistance from my advisors, collaborators, and colleagues. I would first like to thank my advisors, Professor Hameed Metghalchi and Professor Yiannis Levendis. Both of their expertise was invaluable in formulating the research and methodology. Since the first day, I began at Northeastern University; Professor Metghalchi has motivated and encouraged me to conduct valuable research. He also showed immense patience with me and has been supportive of my career goals. I am grateful for the invaluable guidance and mentorship provided by Professor Levendis. He has taught me the methodology to conduct and present research as clearly as possible. He encouraged me to take up new challenges and examine different research projects. Both the professors helped me with providing research assistantship during my masters, which gave me the financial freedom to concentrate on my work and courses. It was a great privilege to work with them and I am grateful for their insightful comments and encouragement.

I am thankful to all of those with whom I have had the pleasure to work during my degree. I want to acknowledge my colleagues from the lab for their incredible effort and helping me in various tasks, with a special mention to Ziyu Wang, Xiao Ma, Zhenyu Lu, Yu Cen Li, Guangying Yu, Ziwei Bai, Mohammed Alswat, and Moaz Omar Allehaibi. I am also thankful to the research team at Texas A&M University, especially Professor Reza Sadr and Dr. Kumaran Kannaiyan. I am grateful to all the professors who have taught me very challenging courses and provided me vast knowledge on thermofluids throughout my masters.

Finally, I thank my parents for their support and confidence in me on this path of higher education.

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ABSTRACT

Alternative refrigerants are an essential subset in reducing the effects of human-made climate change. Among the next-generation refrigerants, hydrocarbons (HCs) are natural and efficient, but their flammability restricts their widespread usage for fire-safety reasons. This research proposes a of hydrocarbons with carbon dioxide (CO2), as refrigerants which have low flammability risks. The objective of this research is to evaluate the thermodynamic performance of vapor compression refrigeration cycle using binary mixtures of HC + CO2 and to measure the laminar burning speed of one selected HC + CO2 mixture with air. Laminar burning speed is measured to evaluate the safety of these refrigerants during an accidental leak.

In this thesis, potential hydrocarbons were identified based on thermophysical properties. Thermodynamic performance parameters, like the coefficient of performance (COP) and volumetric refrigeration capacity, were compared for all selected blends at different operating conditions. Further analysis was performed to determine the non-dimensional irreversibilities in various components of the cycle. Different methods for performing thermodynamic analysis were discussed and evaluated, and a novel approach was developed, including temperature glide matching of zeotropic mixtures. The following four hydrocarbons can be classified as having best performance parameters in blends with CO2: Propylene, , , and .

Propane was selected for laminar burning speed study as it is a widely available and has excellent performance as a . Combustion took place at various CO2 concentrations (0% to 80%), different equivalence ratios (0.7 < ϕ < 1.2) and the laminar burning speed was measured over a range of temperatures (298 K to 420 K) and pressures (0.5 atm to 6.2 atm). Only laminar, spherical, and smooth flames were considered in measuring laminar burning speeds. Pressure rise data as a function of time during the flame propagation was the primary input of the multi-shell thermodynamic model for calculating the laminar burning speed of propane + CO2 + air mixtures. Results show that the addition of CO2 decreases the laminar burning speed of the blends and the COPs of blends are higher to the COPs of pure refrigerants at lower concentrations of CO2. iv

TABLE OF CONTENTS

1. Introduction ...... 1

1.1 Background ...... 1

1.2 Literature Review ...... 5

2. Evaluation of Alternate Refrigerants ...... 11

PART 1: THERMODYNAMIC ANALYSIS (PERFORMANCE STUDY) ...... 19

3 Constant Energy Transfer Fluid Temperature Method ...... 19

3.1 Analysis of Pure Refrigerants ...... 19

3.2 Analysis of Zeotropic Mixtures ...... 26

4. Variable Energy Transfer Fluid Temperature Method ...... 39

4.1 Temperature Glide Matching Models ...... 39

4.2 Performance Parameters ...... 46

PART 2: LAMINAR BURNING SPEED (SAFETY STUDY) ...... 52

5. Experimental Facility and Laminar Burning Speed Model ...... 52

5.1 Experimental Facilities ...... 52

5.2 Laminar Burning Speed Model ...... 53

6. Flame Instability and Results ...... 59

6.1 Flame Structure and Stability ...... 59

6.2 Results and Discussion...... 62

CONCLUSION ...... 68

REFERENCES ...... 69 v

LIST OF TABLES

Table 1 List of pure hydrocarbons after the first screening step ...... 13 Table 2 List of halogenated/oxygenated hydrocarbons after the first screening step ...... 14 Table 3 List of pure flame suppressants after the first screening step ...... 14 Table 4 Selected refrigerants after screening steps and their laminar burning speed values ...... 16 Table 5 Availability of properties for ASHRAE designated mixture refrigerants containing hydrocarbons ...... 18 Table 6 Performance comparison of different pure refrigerants for the two operating conditions ...... 24 Table 7 Critical Peclet number of propane + air mixtures at initial pressure of 2 atm ...... 61 Table 8 List of refrigerants screened out of screening step 1 in chapter 2 of this thesis ...... 80

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LIST OF FIGURES

Figure 1 Hydrofluorocarbons phase-down schedule (Kigali Amendment - 2016) ...... 2 Figure 2 Classification of refrigerants based on their flammability by ASHRAE [12] ...... 3 Figure 3 Energy (heat) of combustion and laminar burning speed for selected refrigerants [15] ...... 4 Figure 4 Schematic of a simple vapor compression refrigeration cycle ...... 12

Figure 5 Critical properties of the selected fluids with varying CO2 concentration ...... 18 Figure 6 Temperature – entropy diagram of a sub-critical simple VCRC using pure refrigerant ...... 20 Figure 7 Flow chart of the thermodynamic model using Constant ETF temperature method ...... 21 Figure 8 COP as a function of compressor discharge pressures in trans-critical cycles ...... 21 Figure 9 Volumetric refrigeration capacity and GWP of different pure refrigerants ...... 23 Figure 10 Non-dimensional irreversibilities for different components using pure refrigerants with operating conditions of set (1); The numbers indicate the position of the refrigerant in Table 6 ...... 25

Figure 11 Temperature-composition diagram for a zeotropic mixture of HC + CO2 ...... 26 Figure 12 Temperature-entropy diagram of sub-critical cycle using zeotropic mixture as refrigerant ...... 27 Figure 13 Temperature-entropy diagram of trans-critical cycle using zeotropic mixture as refrigerant ...... 28

Figure 14 COP of different refrigerant mixtures for varying CO2 mole fractions at set (1) operating conditions ...... 29

Figure 15 VRC of different refrigerant mixtures for varying CO2 mole fractions at set (1) operating conditions ...... 29

Figure 16 COP of different refrigerant mixtures for varying CO2 mole fractions at set (2) operating conditions ...... 30

Figure 17 VRC of different refrigerant mixtures for varying CO2 mole fractions at set (2) operating conditions ...... 30

Figure 18 Condenser pressures for varying CO2 mole fraction at set (2) operating conditions ...... 31

Figure 19 COP of HC + CO2 and HFO + CO2 mixtures for varying CO2 mole fractions at set (1) operating conditions ...... 32

Figure 20 VRC of HC + CO2 and HFO + CO2 mixtures for varying CO2 mole fractions at set (1) operating conditions ...... 32 vii

Figure 21 Comparison of COP for propane + CO2 and isobutane + CO2 mixtures with ‘CYCLE- D’ and ‘this study’ ...... 33 Figure 22 Temperature entropy diagram of Lorenz cycle [75] ...... 33 Figure 23 Temperature entropy diagram of modified Lorenz cycle ...... 34 Figure 24 Non-dimensional irreversibilities for different components at set (1) operating conditions ...... 34 Figure 25 Non-dimensional irreversibilities for different components at set (1) operating conditions ...... 35 Figure 26 Schematic of a VCRC with liquid-line/suction-line energy exchanger...... 35 Figure 27 Temperature entropy diagram of a VCRC with LL/SL-EX using pure refrigerant...... 36

Figure 28 COP of different refrigerant mixtures for varying CO2 mole fractions in a VCRC with LL/SL-EX at set (1) operating conditions ...... 36 Figure 29 Schematic of a VCRC with 2-stage compression ...... 37

Figure 30 COP of propane + CO2 mixture for varying CO2 mole fraction in different cycles ...... 37 Figure 31 Temperature-entropy diagram of a VCRC with constant ETF temperature using zeotropic mixture as refrigerant ...... 39 Figure 32 Temperature-entropy diagram of a VCRC with variable ETF temperature using zeotropic mixture as refrigerant...... 40 Figure 33 Temperature-entropy diagram of a VCRC with variable ETF temperature, using pure refrigerant ...... 40 Figure 34 Flowchart of variable ETF temperature method using minimum (DT) approach/model ...... 43 Figure 35 Temperature - entropy diagram representing the variable ETF method with minimum (DT) approach (input conditions) ...... 44 Figure 36 Temperature - entropy diagram representing the variable ETF method with minimum (DT) approach (quality check of state 4) ...... 44 Figure 37 Temperature entropy diagram representing the variable ETF method (sub-critical cycle) with minimum (DT) approach and linear temperature change assumption ...... 45 Figure 38 Temperature entropy diagram representing the variable ETF method (sub-critical cycle) with minimum (DT) approach and dividing temperature profile into N parts...... 45

Figure 39 COP of different refrigerant mixtures for varying CO2 mole fractions using variable ETF method and minimum (DT) approach with ∆푇푚𝑖푛 = 7, at set (3) operating conditions ...... 46

Figure 40 COP of different refrigerant mixtures for varying CO2 mole fractions using variable ETF method and minimum (DT) approach with different ∆푇푚𝑖푛 of 5, 7 and 10, at set (3) operating conditions...... 47 viii

Figure 41 VRC of propylene + CO2 mixture for varying CO2 mole fraction using variable ETF method and minimum (DT) approach with different ∆푇푚𝑖푛 of 5, 7 and 10, at set (3) operating conditions ...... 47

Figure 42 COP of propane + CO2 mixtures for varying CO2 mole fraction using variable ETF method and minimum (DT) approach with ∆푇푚𝑖푛 = 7, for different operating conditions ...... 48

Figure 43 Non-dimensional irreversibilities of different components for varying CO2

concentration in propane + CO2 mixture using variable ETF temperature method at set (3) operating conditions...... 48 Figure 44 Non-dimensional irreversibilities of different components for different ∆푇푚𝑖푛 in

propylene + CO2 (0.85 + 0.15) mixture using variable ETF temperature method at set (3) operating conditions...... 49

Figure 45 Temperature-entropy diagram of propylene + CO2 (0.4 + 0.6), at set (3) operating conditions ...... 49

Figure 46 Temperature-entropy diagram of propylene + CO2 (0.2 + 0.8), at set (3) operating conditions ...... 50

Figure 47 Comparison of COP for NIST and minimum (DT) model along CO2 mole fraction for

propylene + CO2 mixture at set (3) operating conditions ...... 50 Figure 48 Schematic of the experimental facilities ...... 53 Figure 49 Schematic of laminar burning speed model ...... 54

Figure 50 Images of propane/air/CO2 mixture flames at different pressures and different CO2

concentrations at pi of 2 atm, Ti of 298 K and an equivalence ratio of 1.2 ...... 60

Figure 51 Images of propane/air/CO2 mixture flames at different pressures and different

equivalence ratios at pi of 2 atm, Ti of 298 K and the CO2 concentration of 0.6 ...... 61

Figure 52 Laminar burning speed for propane/air/CO2 mixtures along isentropes at different CO2

concentrations, Ti = 298 K, φ = 1, pi = 1 atm ...... 62

Figure 53 Laminar burning speed for propane/air/CO2 mixtures along isentropes at different CO2

concentrations, Ti = 298 K, φ = 1, pi = 2 atm ...... 63

Figure 54 Laminar burning speed for propane/air/CO2 mixtures along isentropes at different CO2

concentrations, Ti = 298 K, φ = 1, pi = 0.5 atm ...... 63

Figure 55 Laminar burning speed for propane/air/CO2 mixtures along isentropes at different pi of

0.5 atm, 1 atm and 2 atm, Ti = 298 K, φ = 1.2, CO2 = 60% ...... 64

Figure 56 Laminar burning speed for propane/air/CO2 mixtures along isentropes at different

equivalence ratios of 0.7, 1, 1.1 and 1.2, Ti = 298 K, pi = 1 atm, CO2 =30% ...... 65

Figure 57 Laminar burning speed for propane/air/CO2 mixtures along isentropes at different

equivalence ratios of 0.8, 1 and 1.2, Ti = 298 K, pi = 0.5 atm, CO2 = 15% ...... 65 ix

Figure 58 Experimental and modeling results of laminar burning speed for propane/air/CO2

mixtures at different CO2 concentrations, Tu = 300 K, p = 1 atm, φ = 1 ...... 66

Figure 59 Experimental and modeling results of laminar burning speed for propane/air/CO2

mixtures through at different equivalence ratios, Tu = 300 K, p = 1 atm, CO2 = 30%, 60% ...... 66

Figure 60 Experimental and modeling results of laminar burning speed for propane/air/CO2

mixtures through at different unburned gas temperatures, φ = 1, p = 1 atm, CO2 = 60% ...... 67 x

LIST OF ABBREVIATIONS

AC Air Conditioning ASHRAE American Society of Heating, Refrigerating and Air-Conditioning Engineers ASFTE American Society of Thermal and Fluids Engineers COP Coefficient of Performance EOS Equation of State EPA Environmental Protection Agency ETF Energy Transfer Fluid FS Flame Suppressant GHG Greenhouse Gases GWP HAP Hazardous Air Pollutants HC/s Hydrocarbon/s HCFC/s Hydrochlorofluorocarbon/s HFC/s Hydrofluorocarbon/s H/O-HC/s Halogenated/Oxygenated- Hydrocarbon/s LL/SL-EX Liquid-Line/Suction-Line Energy (Heat) Exchanger LMTD Logarithmic Mean Temperature Difference NIST National Institute of Standards and Technology ODP Ozone Depletion Potential REFPROP Reference Fluids Properties Program TC Critical Temperature VCRC Vapor Compression Refrigeration Cycle VRC Volumetric Refrigeration Capacity 1

1. Introduction

This project is part of the research effort in finding alternative refrigerants to the current high global warming potential (GWP) refrigerants, for their use in a vapor compression refrigeration cycle (VCRC). The goal of this thesis is to assess the thermodynamic performance of a VCRC using binary mixtures of hydrocarbons

(HCs) and carbon dioxide (CO2) as refrigerant, and to determine the ideal blends with high performance and low laminar burning speed. The thesis can be broadly classified into two parts, part 1 is determining the performance parameters (thermodynamic analysis), and part 2 is determining the safety parameters in the case of a leak (laminar burning speed study).

The first section of this introduction chapter specifies the importance of finding a new alternative refrigerant and gives a background of the current situation on viable next-generation refrigerants. This section also discusses the relation between laminar burning speed and the safety aspect of it during the accidental leaks. The second section of the chapter includes an elaborate literature review for both thermodynamic analysis of the VCRC and laminar burning speed of the mixtures. This section also introduces more details of the research displayed in later chapters of the thesis.

1.1 Background

Research on alternative refrigerants, having high coefficient of performance (COP) is crucial to reduce the emission of greenhouse gases by refrigeration systems. A refrigerant with higher COP needs less energy to run the cycle/compressor than a refrigerant with lower COP, which means when using a higher COP refrigerant, a reduction of CO2 generation can be observed from the combustion-based power plants (assuming power is generated by fossil fuels). Another impact by refrigerants on climate change is due to the current use of synthetic refrigerants like hydrofluorocarbons (HFCs), which have high GWP. Consequently, these HFCs can cause severe damage to the environment if accidentally leaked into the atmosphere.

Present-day climate agreements like the Kigali amendment [1] to the Montreal protocol are enabling the change to reduce the use of these high GWP refrigerants. Figure 1 shows the HFC phase down schedule as per the amendment, in which the developed countries (Non article 5; main group) have agreed to reduce the use of high GWP refrigerants by 30% before 2024, 60% before 2029 and 70% before 2034. Similarly, other non-article 5 countries have agreed to follow the same path as the main group countries with a delay. 2

Non-Article 5 (main group) include USA, Australia, Canada, EU, Japan

Figure 1 Hydrofluorocarbons phase-down schedule (Kigali Amendment - 2016)

Paul Hawken [2] ranks “refrigerant management” as the first solution among the top 100 solutions

9 to reverse global warming. He mentions that a reduction of 89.7 gigatons (10 metric tons) of CO2 equivalent emissions by 2050 can be seen only through refrigerant management. He also estimates that the implementation of Kigali amendment will reduce global warming by one-degree Fahrenheit. The European F-gas regulation (EU), No 517/2014 of the European parliament and of the council of 16 April 2014 [3], states a target in the reduction of by 72-73% by 2030 and by 78% by 2050. This regulation 6 also mentions that the fluorinated GHG emissions were estimated at 90 million tons (10 metric tons) of CO2 equivalent in 2005 in European Union. A 60% reduction means that emissions would have to be reduced to approximately 35 million tons of CO2 equivalent by 2030. To achieve these goals, a quest for alternative refrigerants, which have low GWP, zero ozone depletion potential (ODP), and high efficiency is needed.

McLinden et al. [4] have studied a large number of components from PubChem [5] database to find pure refrigerants for sub-critical applications which have low GWP, zero ODP, low toxicity, and generous volumetric refrigeration capacity (VRC). They used these parameters for screening the fluids, and it resulted in 27 alternative refrigerants, including hydrocarbons & dimethyl ether, fluorinated alkanes, fluorinated (or ) and alkynes, fluorinated oxygenates, fluorinated & sulfur compounds and inorganic compounds. Among the 27 refrigerants, good COP and low toxicity refrigerants are available, but all of them are either slightly or highly flammable. Non-flammable candidates exist among the 27 refrigerants but have low volumetric refrigeration capacities. All the screened-out refrigerants have one or the other setback which can be flammability, low critical point, unknown risks of new synthetic refrigerants, and low VRC. They also conclude that the viable candidates for single component low-GWP alternatives for small AC systems are very limited, especially for refrigerants with VRCs like R-410A (CH2F2

{0.5} + CHF2CF3{0.5}) .

To find a trade-off between flammability and VRC, multi-component refrigerants with flammable component like hydrocarbons (HCs) and a flame suppressant component like carbon dioxide (CO2) can be a 3

way-forward as the potential alternative refrigerants. Other natural substances like [6], blends of CO2

+ NH3 [7], blends of CO2 + brine [8], and HCs [9] are also studied in literature as potential alternatives.

Hydrocarbons as pure refrigerants are considered natural, low GWP, efficient, non-toxic, and economically viable alternatives to the HFC refrigerants [10] However, the high flammability of HCs restricts their widespread usage. American Society of Heating, Refrigerating, and Air-conditioning Engineers (ASHRAE) classified the hydrocarbon refrigerants based on their flammability [11] as A3, extremely flammable. This classification is depicted in Figure 2, taken from reference [12]. This figure shows several flammability classes from non-flammable, class 1 to highly-flammable, class 3 depending on the refrigerant’s energy (heat) of combustion, laminar burning speed, and lower flammability limit. According to this classification, hydrocarbons are “class 3”, which makes them highly flammable. Where CO2 is ‘class 1’, and the mixtures of HCs and CO2, have not yet been designated any class for their flammability. A review of standards until 2008, concerning the usage of pure HCs as refrigerants, is discussed in reference [13]. They summarize that many of these standards have stringent restrictions against the usage of HCs, and some of the countries allow HCs as a refrigerant in low charge mass applications.

Figure 2 Classification of refrigerants based on their flammability by ASHRAE [12]

Carbon dioxide on its own is a with low GWP. However, due to its low critical temperature of 31 °C, its usage in tropical climates is restricted as the VCRC tends to be trans-critical, which will work less efficiently than a sub-critical cycle. The use of CO2 also leads to very high pressures in the condenser/gas cooler part of the cycle [14]. The blends of HC and CO2 behave as zeotropic mixtures, and they can weaken their shortcomings by decreasing the large pressures encountered in the case of pure CO2 and by reducing the fire-safety risks of pure HCs. These blends also have the advantage of having higher critical temperatures when compared to that of pure CO2, which tends the cycle with lower CO2 concentrations to be sub-critical. 4

ASHRAE assigns laminar burning speed as one of the safety parameters for the flammable refrigerants. Laminar burning speed is a thermo-physical property that is the direct measurement of the rate of energy released during the combustion process of any combustible mixture and is a direct function of pressure, temperature, equivalence ratio, diluent type, and fuel composition. Physically, laminar burning speed is the rate of expansion at which a planar, one-dimensional, adiabatic flame front travels relative to the unburned gas mixture.

Figure 3, taken from reference [15] shows laminar burning speed and energy (heat) of combustion for different refrigerants. This figure displays hydrocarbons like R-290 (propane) & R-600a (isobutane) as class 3 refrigerants, hydrofluoroolefins (HFOs) like R-1234yf (2,3,3,3-Tetrafluoropropene) & R-1234ze(E) (trans-1,3,3,3-tetrafluoroprop-1-ene) as class 2L refrigerants, and HFCs like R-32 () & R- 134a (1,1,1,2-Tetrafluoroethane) as class 2L refrigerants. A class 2L refrigerant is one which is defined as mildly flammable, with lower flammability limit more than 100 g/m3, energy (heat) of combustion less than 19,000 kJ/kg, and laminar burning speed less than 10 cm/s. When laminar burning speed is less than 10 cm/s, a stagnant flame like a candle is observed, as the average velocity of the air in an indoor environment is 10 cm/s. It also needs to be noted that the addition of CO2 will increase the lower flammable limits and decrease the upper flammable limits of the mixtures. Zabetakis [16] performed a detailed study on the flammability limits on the mixtures of propane, and other hydrocarbons with CO2 and air in 1965.

Figure 3 Energy (heat) of combustion and laminar burning speed for selected refrigerants [15]

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1.2 Literature Review

This section is mainly classified into two parts, viz. refrigeration cycle literature and laminar burning speed literature. Both parts explain previous studies and the outcomes from them. This section also introduces the research handled in the rest of the thesis, including the different methods for thermodynamic analysis, and experimental conditions for laminar burning speed measurement.

Refrigeration cycle literature:

There have been many theoretical and experimental studies on alternative refrigerants, but a limited number of studies on using the blends of hydrocarbons and carbon dioxide as refrigerants. The blends of HCs and CO2 are zeotropic mixtures as they have different boiling points. The key feature of a zeotropic mixture is the temperature glide, it is a difference of the mixture temperature during an isobaric phase change (vapor- liquid equilibrium). Further chapters in this thesis explain in more details about the zeotropic mixtures and its features. The thermodynamic analysis literature is shown next.

McLinden and Radermacher [17] have shown two thermodynamic analysis methods for comparing the performance of pure and mixed refrigerants. The first method has a constant energy (heat) transfer fluid (ETF) temperature and the second method has a variable ETF temperature in the energy (heat) exchangers. The latter method is performed to match the temperature glide of the zeotropic mixtures during phase change in counterflow energy (heat) exchangers and is commonly known as ‘temperature glide matching’. This reduces the irreversibility associated with the energy (heat) exchange process between the refrigerant and the energy transfer fluid. McLinden and Radermacher concluded that for a fair and complete comparison of the mixture and pure component performance, it is necessary to consider the change of ETF temperature. Preliminary results using the constant ETF temperature method was presented as an extended abstract at the ASTFE third thermal and fluids engineering conference [18]. Although NIST has developed a graphic user interface based software/platform named CYCLE_D [19] to perform the thermodynamic analysis for a system with constant temperature of ETF, it was not used in this study. This is because of data discrepancies at the transition point between sub-critical and trans-critical cycles, which is discussed in chapter 3.

For the variable ETF temperature method, a graphic user interface software called CYCLE_D-HX [20] by NIST performs thermodynamic analysis with the input of either mean-effective temperature difference (∆푇ℎ푥) or the overall energy (heat) transfer conductance (푈퐴ℎ푥) for both the counter flow energy (heat) exchangers in the cycle. Where either of these inputs are to be taken from experimental data or as an educated guess. From here on, this approach by NIST for variable ETF temperature method will be referred as ‘NIST model’. Apart from the requirement of experimental data, the NIST model is developed only for sub-critical VCRC, where the VCRC with blends of HC + CO2 can be either sub-critical or trans-critical. The concept of the mean effective temperature difference and the details on the NIST model is given in reference [21]. More specifics and logic on the initial version of Lorenz based cycle by NIST is given as CYCLE-11 6

and seen in reference [22]. Thermal machinery group at NIST has studied the temperature glide matching with binary and ternary zeotropic mixtures both experimentally and theoretically [23; 24] using the NIST model. Brignoli et. al [25] compared the CYCLE_D and CYCLE_D-HX model and validated CYCLE_D- HX model with experimental data and showed a maximum deviation of 7.4% for the values of coefficient of performance.

A different approach is observed in literature for variable ETF temperature method, which specifies the minimum temperature difference between the refrigerant and the ETF, ∆푇푚푖푛. This approach does not require any experimental input when compared to the NIST model, but the practicality of the energy (heat) exchanger area comes to the question. For example, an energy exchanger with low ∆푇푚푖푛 value will have large area and consequently when ∆푇푚푖푛 = 0, the area of the energy exchanger is infinite. From here on this approach to variable ETF temperature method will be referred as ‘minimum (DT) model’. A similar approach was observed in reference [26], where the ∆푇푚푖푛 was taken as 5 °C and the analysis was done for an energy

(heat) pump using various mixtures and one of them is dimethyl ether + CO2. Similarly [27] shows thermodynamic analysis using minimum (DT) model, where ∆푇푚푖푛 was taken as 7 °C for isobutane + CO2 and butane + CO2 mixtures with an assumption of a linear temperature profile of refrigerant in energy (heat) exchangers of energy (heat) pump cycle.

Many experimental studies in literature showcase the increase of COP when using zeotropic mixtures as refrigerants, in a VCRC with counterflow energy (heat) exchangers. Kim et al. [28] performed an experimental analysis on CO2 + C3H8 mixtures for cooling applications in trans-critical cycles and taking into the effect of temperature glide matching with a mass fraction of CO2 ranging from 0.6-1.0. They showed that COP of mixtures in trans-critical cycles has better COP than pure CO2 cycle. Li et al. [29] showed that a binary mixture of HFCs has maximum COP than a single compound HFC experimentally. Ju et al. [30] performed the experimental analysis for energy (heat) pump water heater using CO2 and propane blends with a mass fraction of CO2 ranging from 0.08-0.16. They showed that these blends have comparable performance to pure R22 (CHClF2). Kim and Kim [31] performed experimental and simulation studies of an auto-cascade refrigeration system using propane + CO2 mixture with CO2 concentration varying from 0.13-0.31. They showed that COP decreased with the addition of CO2 in the mixture albeit the increase cooling capacity. Exergetic analysis on alternative refrigerants was studied in reference [32].

Apart from the two models mentioned earlier i.e. NIST model and minimum (DT) model, Ali Hakkaki-Fard et al. [33] developed a numerical model with detailed simulation of air-source residential energy (heat) pump using refrigerant mixtures, including energy (heat) transfer characteristics by determining the pinch values of the energy (heat) exchangers.

Other factors like pressure drop in the VCRC and the change in circulation concertation have an impact on the performance. For example, Afroz et al. [34] experimentally studied the energy (heat) transfer coefficient and pressure drop of CO2 + dimethyl ether mixture. They showed that the pressure drops, and the 7

energy (heat) transfer coefficient decrease during condensation, with the addition of CO2 concentration in the mixture. A decrease of pressure drop signifies an increase in COP as the energy (work) input to the compressor reduces. Kim et al. [35] experimentally studied the change in circulation concentration of CO2 + propane mixtures in a VCRC. They reported that the cause for this change in concentration is the composition shift. They also showed that the circulation concentration of the mixture has a higher CO2 mole fraction of 0.03 than the initially charged mixture concentration.

Studies shown in references [36; 37; 38] perform thermodynamic analysis for cascade cycles.

Among them, Niu and Zhang [36] investigated the performance of a zeotropic mixture of CO2 + propane at a mole ratio of 71/29 in the low-temperature cycle of a cascade refrigeration system. This refrigerant mixture was proposed as an alternative to R-13 (CClF3). They reported that the COP, volumetric refrigeration capacity (VRC), and the operating pressures of the mixture were higher than those of the synthetic R-13 refrigerant.

Whereas [38] shows the usage of propane in a low-temperature circuit and CO2 in a high-temperature circuit. Onaka et al. [39] theoretically studied the performance of an energy (heat) pump using the zeotropic mixture of CO2 + dimethyl ether.

Nicola et al. [37] theoretically investigated the performance of a cascade refrigeration cycles operated with binary blends of CO2 + HC on the low-temperature side of the cycle. They evaluated several hydrocarbons like , propane, , propylene, and dimethyl ether. They reported that the COP of the cycle with the mixtures are slightly lower than the cycle using pure hydrocarbons. However, they recommended that more realistic models and experimental studies are needed to propose an optimal mixture.

Fan et al. [40] theoretically studied the performance of CO2 + DME mixture in an energy (heat) pump system and reported that the COP could be enhanced by about 8% when compared to a system with pure HCFC refrigerants. Same authors experimentally studied the performance of an energy (heat) pump system while using the CO2 + propane mixtures [41]. They reported that the addition of propane to CO2 helps to enhance system performance and reduce the discharge pressures from the compressor. Recently, Wu et al. [42] reported the performance of different combinations of CO2 + HC binary mixtures for geothermal power plant applications. They emphasized that all the CO2 + HC combination can lead to reducing the cost per net power of the plant when compared to that of pure CO2, although more energy (heat) transfer area is required for mixtures.

All the above studies highlight the interest in zeotropic mixtures of CO2 + HC for refrigeration and energy (heat) pump applications. However, most of the above studies are theoretical and focused on refrigeration systems in very low-temperature applications. Furthermore, the range of HCs evaluated for refrigeration system is also very limited. Although the zeotropic mixture of HC and CO2 helps to overcome the limitations of natural refrigerants, they introduce new challenges in the design and operation of vapor compression refrigeration systems [43]. The two significant challenges associated with zeotropic mixtures are the shift in mixture composition and the glide in mixture temperature during the phase change phenomena. 8

It is well established that the hydrocarbons have excellent refrigeration characteristics when compared to HFC refrigerants. As highlighted by Yang and Wu [44], their flammability properties must be suppressed to be able to use them in commercial refrigeration systems safely. Therefore, the thermodynamic analysis part of this study has three main objectives:

(1) ‘Identify a set of hydrocarbons that can be used as a zeotropic binary mixture with CO2 as working fluid in refrigeration and air-conditioning systems without deteriorating the performance.’ This was achieved using various screening parameters established based on thermophysical properties. These properties for the refrigerants were determined using REFPROP 10 (reference fluid thermodynamic and transport properties) [45], a database developed by national institute of standards and technology (NIST). This part of the study initially evaluates the available pure refrigerants in REFPROP 10 for binary mixtures. (2) ‘Further identification of hydrocarbons based on their performance parameters for different operating conditions.’ This was accomplished using a VCRC with constant ETF temperature, where the model required for thermodynamic analysis for selected refrigerants is developed in MATLAB. (3) ‘Enhancement of the cycle performance through the temperature glide matching of the refrigerant mixture and the energy (heat) transfer fluid.’ An analysis was conducted for the selected set of

hydrocarbons and CO2 in a VCRC with variable ETF temperature. This was performed using the minimum (DT) model, where a novel approach is developed in MATLAB using logical operators. The model developed in this study is good for both sub-critical and trans-critical cycles. To represent different counterflow energy (heat) exchanger designs, the thermodynamic analysis was performed for different

∆푇푚푖푛 values.

Other than the mentioned objectives, this part of the thesis also discusses the benefits and drawbacks of the developed novel approach ‘minimum (DT) model’ and was compared to the ‘NIST model’ used in CYCLE_D-HX. Further thermodynamic analysis for both constant ETF temperature method, and variable ETF temperature method shows non-dimensional irreversibilities and exergetic efficiencies. This was performed to determine the exergy loss contributions by each component in a simple VCRC.

Laminar burning speed literature:

To determine the safety of the mixtures of HC + CO2, laminar burning speed is measured as discussed in section 1 of this chapter. Propane was among the hydrocarbons with good coefficient of performance and widely available . Bearing that in mind, this study selected propane for the laminar burning speed measurement. Preliminary results on the laminar burning speed of propane + CO2 + air mixtures were presented as an extended abstract at the ASTFE third thermal and fluids engineering conference [46].

Metghalchi and Keck [47] have burned propane + air mixtures in a spherical constant volume chamber at equivalence ratios (φ) (0.8 < φ < 1.5), initial pressures (Pi) (0.4 atm < Pi < 7.6 atm), and initial temperatures (Ti) (298 K < Ti < 500 K). They have determined the laminar burning speeds for a pressure range of 0.4–40 atm and a temperature range of 298–750 K. Ebaid et al. [48] burned propane + air mixtures 9

in a pressure cylinder and measured the laminar burning speed at pressures 0.5–1.5 bar, temperatures 300–

350 K and at equivalence ratios (0.6 < φ < 1.5). Akram et al. [49] burned the propane + air + CO2 and propane

+ air + N2 mixtures in a preheated mesoscale diverging channel and measured the laminar burning speed of these mixtures at equivalence ratios (0.7 < φ < 1.3), atmospheric pressure, preheat temperatures 370–650 K, and diluent concentrations up to 40%.

The present study examined the effect of CO2 on the laminar burning speed of propane + air mixtures. Thus, we burned the propane + air + CO2 mixtures in a cylindrical constant volume chamber at equivalence ratios (0.7 < φ < 1.2), initial pressures (0.5 atm < Pi < 2 atm), the initial temperature of 298

K, and at CO2 concentrations up to 80%. The laminar burning speed of these mixtures was determined for a pressure range of 0.5–6.2 atm and a range of temperatures of 298–420 K. The unburned mixture composition with air was defined as

휙{(1 − 훼)퐶3퐻8 + 훼퐶푂2} + 5(1 − 훼)(푂2 + 3.76푁2)

Where, α is the fraction of CO2 in the propane and CO2 mixture. The experiments at higher pressures and temperatures were performed as it has been observed in [18; 28] that a vapor compression refrigeration cycle with a blend of hydrocarbons and CO2 has higher pressures and temperatures at the end of compression. The experimental results were compared with the simulations results obtained via a steady one-dimensional (1D) laminar premixed code using the CANTERA [50] open source. The laminar burning speed results presented in this thesis are published as a journal article, as seen in reference [51].

Leak analysis literature:

The laminar burning speed is only relevant when there is a leak and bearing it in mind, this part of the section gives details on the literature regarding the leak study for a better understanding of the scenario. Few researchers have studied how the concentration of a zeotropic mixture changes and how a flammable gas/dense gas leaks over time. For a mixture of HC + CO2 undergoing phase change, the low volatile component (CO2) will leak out first for a vapor leak and a high volatile component (HC) will leak out first for a liquid leak. It is because, less volatile component condenses faster than the more volatile component and vice versa in evaporation.

Kim and Didion [52] have developed a thermodynamic model to determine the composition shift and pressure change in the refrigerant tank with time and is used in a graphic user interface platform/software called REFLEAK [53] to evaluate composition shift in this situation. A modified version of this thermodynamic model is proposed by Clodic and Riachi [54] which considers finite mass and the effects of room conditions on the leak. Nagasoa [55], has developed a numerical model with constant inlet velocity and constant inlet concentration with time for a single component dense gas leak. He used a laminar model, contrary to the other studies; He also states that Schmidt number ~ 1; Kinematic viscosity (ν) = (D) mass diffusivity for the dense gas. Similarly, many others [56; 57] have performed a numerical/CFD analysis of a single component leak using turbulent model. 10

Afzal et al. [58] performed a simulation and an experimental study for the leak of propane, CO2, and air into a room filled with air. Their results show a difference of 20% between experimental and simulation values. Afzal et al. assumed that the concentrations of propane and CO2 in the leaking mixture (propane, CO2, and air) as very small. They also get the unsteady inlet velocity input for simulations from the experiments.

This assumption does slightly represent a similar scenario, but in this study, only propane and CO2 are leaking, i.e., without air, and there is a composition shift. It has not been observed of any numerical or analytical analysis of a multicomponent leak which represents the case of zeotropic mixtures. Similarly, like Afzal et al., Jia et al. [59] performed an experimental study for the leak of R-32 and Li [60] for the leak of Propane.

This thesis does not study the leak analysis of the zeotropic mixtures or pure components in further to mentioned earlier.

11

2. Evaluation of Alternate Refrigerants

For determining the thermophysical properties, a database called reference fluid thermodynamic and transport properties (REFPROP 10) [45] developed by the NIST was used. In REFPROP, there are codes written to solve an equation of state (EOS) for a given fluid/mixture. For solving these codes (EOS), an input of two fluid properties like pressure and temperature are required , which defines a state. After defining a state, the output of the code displays the calculated thermophysical properties. The saturated states for liquid- vapor equilibrium are defined with a single state property as an input. The purpose of choosing REFPROP is due to its experimental validation, its extensive use in industrial applications, and its use in several thermodynamic textbooks and research articles.

The EOS’s used in REFPROP for hydrocarbons and carbon dioxide are expressed in the form of Helmholtz free energy with independent variables of density and temperature (in addition to the molar composition of fluids for a mixture). More details on the Helmholtz energy-based EOS which are used in REFPROP can be found in the references [61; 62; 63].

Vapor compression refrigeration cycle (VCRC):

A simple vapor compression refrigeration cycle (VCRC) includes four components; compressor, condenser/gas cooler, expansion valve, and evaporator, as shown in Figure 4. A VCRC can be operated either as sub-critical or trans-critical. Sub-critical cycle indicates a refrigerant phase change (vapor-liquid equilibrium) process happening in both the energy (heat) exchangers, i.e., condenser and evaporator. Whereas, trans-critical cycle has refrigerant phase change process happening only in one of the energy (heat) exchangers, i.e. evaporator. In the other energy exchanger, only gas cooling process takes place (condenser is replaced by gas cooler). The state points in the VCRC are defined as follows; process 1-2 is compression, process 2-3 is condensation/gas cooling, process 3-4 is expansion, and process 4-1 is evaporation.

Screening steps:

Refrigerants for VCRC should be environmentally friendly, capable of having an easy phase change, and good energy (heat) transfer characteristics. The thermophysical properties that are relevant to the energy (heat) transfer phenomena ( critical temperature, normal ), environmental impact metrics (ODP, GWP), hazardous air pollutant classification, and toxicity, were used as criteria to narrow down on the list of pure refrigerants. Further screening is done based on the mixture data availability in REFPROP. In order to determine the effective binary mixture refrigerants among the 155 pure refrigerants available in REFPROP 10, a four-step screening process was considered and is explained as follows : 12

Figure 4 Schematic of a simple vapor compression refrigeration cycle

Screening Step 1: Environmental Factors

• All the 155 pure refrigerants are checked for their environmental factors as mentioned: Low GWP

(<100), calculated as 100 year CO2 equivalent; Zero ODP (~ 0), calculated as R-11 equivalent; Non- toxic; Not classified as hazardous air pollutant (HAP) by environmental protection agency (EPA) [64]; Not extremely flammable like . • The initial list of fluids after the first screening parameter step mentioned above include 37 pure HCs as shown in Table 1, 20 halogenated/oxygenated HCs (H/O-HC) as shown in Table 2, and 9 flame suppressants (FS) as shown in • • Table 3. The total number of fluids from these three tables is 66. Attached appendix to this thesis includes the refrigerants which are not shown in tables (1-3), i.e. the fluids which were screened out in this step. The thermophysical properties of refrigerants shown in the following tables (1-3) are critical temperature and critical pressure. It also lists ASHRAE designation and of the refrigerant. In addition to these, the tables also show coefficient of performance (COP) and volumetric refrigeration capacity (VRC). The performance parameters shown in these tables are calculated for air conditioning applications in tropical climates. The details on the calculation of these performance parameters is shown in chapter 3 and these tables are cross-referred when required in the further chapters of this thesis.

Among the tables (1-3), the following symbols showcase

‘-‘ No REFPROP designation ‘c’ indicates that the chemical structure is cyclic

‘S’ indicates the sublimation temperature ‘*’ evaporator exit temperature (푇1) < triple point (푇푡푝)

‘#’ evaporator exit temperature (푇1) > critical temperature (푇푐) 13

Table 1 List of pure hydrocarbons after the first screening step

N Chemical name Chemical formula ASHRAE Tc Pc COP VRC o designati-- (K) (MPa) (1) (MJ/m3) on (1) 1  CH4 R-50 191 4.6 # # 2 Acetylene CH≡CH - 308 5.0 1.9 11.7 3 Ethene CH2=CH2 R-1150 282 5.0 # # 4 Ethane CH3-CH3 R-170 305 4.9 1.5 9.1 5 .Propadiene C3H4 - 398 5.2 4.7 3.0 6 .Prop-1-yne CH3-C≡CH - 402 5.6 4.9 2.7 7 CH2=CH-CH3 R-1270 364 4.6 4.3 4.1 8 .Cyclopropane c-CH2-CH2-CH2- R-C270 398 5.6 4.8 3.1 9 Propane CH3-CH2-CH3 R-290 370 4.3 4.3 3.4 10 .1-Butyne CH3-CH2-C≡CH - 432 4.1 4.9 0.9 11 .Cyclobutene c-CH=CH-CH2-CH2- - 448 5.2 5.0 1.1 12 1-butene CH3-CH2-CH=CH2 - 419 4.0 4.7 1.3 13 .Cis-butene CH3-CH=CH-CH3 - 436 4.2 4.9 1.0 14 .Trans-butene CH3-CH=CH-CH3 - 429 4.0 4.8 1.1 15 Isobutene (CH3)2-C=CH2 - 418 4.0 4.7 1.3 16 Butane CH3-CH2-CH2-CH3 R-600 425 3.8 4.7 1.1 17 Isobutane (CH3)2-CH-CH3 R-600a 407 3.6 4.6 1.4 18 .1-Pentene CH2=CH(CH)2-CH3 - 465 3.6 4.8 0.4 19 Cyclopentane c-CH2-3(CH2)- CH2- - 512 4.6 5.1 0.2 20 CH3-3(CH2)-CH3 R-601 470 3.4 4.8 0.3 21 (CH 3) 2-CH-CH2-CH 3 R-601a 460 3.4 4.8 0.4 22 Neopentane C-(CH3)4 R-601b 434 3.2 4.6 0.7 23 Cyclohexane c-CH2-6(CH2)- CH2- - 554 4.1 5.0 0.08 24 Isohexane CH3-CH- - 498 3.0 4.8 0.1 (CH3)(CH2)2CH3 25 2,2- (CH3)3C-CH2-CH3 - 490 3.1 4.7 0.2 Dimethylbutane 26 2,3- (CH3)2-CH-CH(CH3)2 - 501 3.2 4.8 0.2 Dimethylbutane 27 3- (CH3-CH2)2-CH-CH3 - 506 3.2 4.8 0.1 Methylpentane 28 Methylcyclohe- C6H11-CH3 - 572 3.5 5.0 0.04 -xane 29 Heptane CH3-5(CH2)-CH3 - 540 2.7 4.8 0.03 30 Octane CH3-6(CH2)-CH3 - 569 2.5 4.9 0.01 31 Propylcyclohe- C6H11-(CH2-CH2-CH3) - 631 2.9 4.9 0.003 -xane 32 Nonane CH3-7(CH2)-CH3 - 595 2.3 4.9 0.003 33 Decane CH3-8(CH2)-CH3 - 618 2.1 4.9 0.001 34 Undecane CH3-9(CH2)-CH3 - 639 2.0 4.9 0.0003 35 Dodecane CH3-10(CH2)-CH3 - 658 1.8 4.9 0.00001 36 Hexadecane C16H34 - 722 1.5 * * 37 Docosane C22H46 - 792 1.2 * *

14

Table 2 List of halogenated/oxygenated hydrocarbons after the first screening step

N Chemical name Chemical formula ASHRAE Tc Pc COP VRC o designation (K) (MPa) (1) (MJ/m3) (1) 1  CH3-F R-41 317 5.9 1.9 12 2 Ethanol CH3-CH2-OH - 514 6.3 5.1 0.05 3 Dimethyl ether CH3-O-CH3 R-E170 400 5.3 4.7 2.6 4 .Fluoroethane CH3-CH2-F R-161 375 5.0 4.6 3.8 5 Propan-2-one CH3-(C=O)-CH3 - 508 4.7 5.1 0.2 6 CH3-CH2-O-CH2-CH3 R-610 467 3.6 4.8 0.4 7 Monoethanolami- HOCH2CH2NH2 671 8.1 * * -ne 8 .Trifluoroethylene CF2=CHF R1123 332 4.5 3.3 5.6 9 Dimethyl CH3-O-(C=O)-O-CH3 - 557 4.9 5.1 0.04 carbonate 10 2,3,3,3- CH2=CF-CF3 R-1234yf 368 3.4 4.2 2.4 Tetrafluoropropene 11 Trans-1,3,3,3- CHF=CH-CF3 R- 383 3.6 4.4 2 tetrafluoroprop-1- 1234ze(E) ene 12 Methyl Palmitate C17H34O2 - 755 1.4 * * -9 13 Methyl linolenate C19H32O2 - 772 1.4 4.9 10e -9 14 Methyl linoleate C19H34O2 - 799 1.3 4.9 10e -9 15 Methyl oleate C19H36O2 - 782 1.3 4.9 10e 16 Methyl stearate C19H38O2 - 775 1.2 * * 17 .Cis-1,3,3,3- CHF=CH-CF3 R1234ze(Z 423 3.5 4.8 0.9 tetrafluoroprop-1- ) ene 18 .3,3,3- CH2=CF-CF3 R1243zf 377 3.5 4.4 2.3 Trifluoropropene 19 .1,1,1,4,4,4- CF3CH=CHCF3(Z) R1336mzz 445 2.9 4.7 0.4 Hexafluoro-2- (Z) butene 20 Dodecafluoro-2- CF3CF2(C=O)CF(CF3)2 - 442 1.9 4.0 0.2 methylpentan-3-one

Table 3 List of pure flame suppressants after the first screening step

No Chemical Chemical ASHRAE Tc Pc COP VRC name formula designation (K) (MPa) (1) (MJ/m3) (1) 1 • He R-704 5 0.2 # # 2 • Ne R-720 45 2.7 # # 3 •Water H2O R-718 647 22 4.6 0.02 4 •Heavy water D2O - 644 22 4.5 0.02 5 •Nitrogen N2 R-728 126 3.4 # # 6 • Ar R-740 151 4.9 # # 7 Carbon dioxide CO2 R-744 304 7.4 1.9 13 8 • Kr - 210 5.5 # # 9 •Xenon Xe - 290 5.8 1.3 11 15

Screening Step 2: Flame Suppressants

In addition to CO2, there are a few compounds available in REFPROP, which can be used as flame suppressants. They are not included in this study due to the following reasons:

• Flame suppressants like helium, neon, argon, krypton, and nitrogen are neglected since they have very low critical temperatures and normal boiling points. These low temperatures indicate that for a phase change in air-conditioning (AC) applications, these compounds need very high operating pressures, which will make the refrigerant mixtures supercritical and in turn lead to low COP.

• Flame suppressants like water and heavy water are neglected because they have very high critical temperatures and normal boiling points. Consequently, the phase change at room cooling temperatures will be at shallow pressures, which in turn results in complex system design.

• Only CO2 and xenon will have the phase change at moderate pressures for AC applications. However,

when compared to CO2, xenon is expensive, and the hydrocarbon-xenon binary mixture properties are

not available in REFPROP 10. Therefore, only CO2 is the viable ideal flame suppressant for further analysis.

Fluids filtered out in this step have a symbol ‘•’ placed beside their chemical names in

Table 3.

Screening Step 3: Critical Temperature

Critical temperature of a refrigerant plays crucial role in the energy (heat) transfer performance. For example, critical temperature of the refrigerant should be above the temperature of the cooling space to operate the evaporator with a phase change. Another effect of critical temperature is that refrigerants with higher critical temperatures tend to have lower volumetric refrigeration capacity. Based on this, a screening parameter on critical temperature was established and is shown below. McLinden et. al [4] showed a similar screening parameter of critical temperature, but with a different range.

• An acceptable critical temperature (푇푐) range of 320 K < 푇푐 < 500 K for the refrigerants was considered.

The lower and upper limits of the range are higher than critical temperature of pure CO2, as the critical

properties will be affected by the change of CO2 concentration in the mixture and will be discussed later in Figure 5. • Fluids that have been filtered out from this step are indicated by the symbol ‘’ besides their chemical name in Table 1 and Table 2. 16

After this step, it can be seen from Table 1 that HCs with more than 6 carbons are dropped out and will not be considered in further analysis. The resultant number of fluids from Table 1 are 17, and from Table

2 are 10, which gives a total of 28 fluids including CO2.

Screening Step 4: Mixture Data Availability

In this step the availability of the binary interaction parameters for CO2 based mixtures in REFPROP 10 was checked. This assessment was performed for all the refrigerants resulting from step 3. Only the refrigerant mixtures with binary interaction parameters offered in REFPROP 10 are selected for further analysis. Selected refrigerants in this chapter (2) are used in chapter 3 for determining the performance parameters for constant ETF temperature system. Refrigerants filtered out in this screening are indicated by a symbol ‘.’ besides their chemical name in Table 1 and Table 2.

• The resultant number of refrigerants after this step including CO2 is 15, which means that 14 different

binary mixtures with CO2.

The final 15 refrigerants screened out from the four screening steps are given in Table 1. This table also lists their respective laminar burning speed (Sl) values with air at standard conditions. It can be seen from this table that HCs and H/O-HCs have a significant laminar burning speed when compared to HFOs. The resultant 15 refrigerants from screening steps include 2 hydrofluoroolefins (HFOs), which are R1234yf, and R1234ze(E). The list also includes an engineered fluid from 3M, which is Novec 649. These refrigerants have zero ODP, low GWP, low laminar burning speeds (< 10 cm/s), and are being commercialized as an alternative solutions. However, HFOs have other environmental/performance issues as highlighted below: • Low volumetric refrigeration capacities, • Hydrochlorofluorocarbons as an ingredient in the making of some HFOs,

• Hydrofluoroolefins have the potential to release a toxic byproduct called Trifluoroacetic acid (C2HF3O2), which can have a detrimental effect on aquatic environments [65; 66], • Products from the combustion of HFOs, include poisonous hydrogen fluoride (HF) and carbonyl fluoride

(COF2) gases [66].

Above issues may prevent them as a long-term alternative. Even though, HFOs are mildly flammable or non-flammable, the key reason for mixing them with CO2 is to increase their VRC and in turn change the COP.

Table 4 Selected refrigerants after screening steps and their laminar burning speed values

No Refrigerant 푺풍 (cm/s) No Refrigerant 푺풍 (cm/s) e 1 Carbon Dioxide # 9 Isobutene (C4H8) 35.2 a f 2 Dimethyl ether (C2H6O) 42.1 10 Butane (C4H10) 37.0 b f 3 Propylene (C3H6) 39.3 11 Isobutane (C4H10) 33.4 c g 4 Propane (C3H8) 38.6 12 Diethyl ether (C2H5)2O 40.3 d 1 5 2,3,3,3-Tetrafluoropropene 1.5 13 Pentane (C5H12) 42.6 (C3H2F4) 17

6 Trans-1,3,3,3-tetrafluoroprop-1- # 14 Isopentane (C5H12) ene (C3H2F4) e 7 1-Butene (C4H8) 40.3 15 Neopentane (C5H12) 8 Dodecafluoro-2-methylpentan-3- # one (C6F12O) # Non-flammable at standard conditions a Experimental data, from reference [67] b Experimental data, from reference [68] c Experimental data, from reference [51] d Experimental data, from reference [69] e Experimental data, from reference [70] f Experimental data, from reference [71] g Experimental data, from reference [72] 1 Simulation data, using mechanism [73]

Mixtures with more than two refrigerants

Up until now only binary mixtures were considered, i.e. a flammable fluid with carbon dioxide.

However, ternary mixtures of 2 flammable fluids and CO2, quaternary mixtures of 3 flammable fluids and

CO2, and similarly a mixture with ‘n’ components including CO2 is also possible. The necessity for using mixtures with more than two fluids is to increase the efficiencies by linearizing the temperature profile in the energy exchanger. Linearization of temperature profile by adding a third component is shown experimentally in reference [24], where the third component has intermediate boiling point to the initial two components of the mixture. ASHRAE has designated a few mixtures of hydrocarbons (without carbon-dioxide) as refrigerants and Table 5 shows some of these refrigerants, which are also available in REFPROP. This table also lists the critical temperature (Tc), global warming potential (GWP), and its mixture data availability with carbon dioxide (CO2) in REFPROP. In this thesis/study only binary mixtures are studied, possibility of ternary and higher mixtures are mentioned for better understanding of the scenario.

Critical properties of mixtures:

The critical properties determined using REFPROP are shown in Figure 5. It displays the values of critical temperature and critical pressure with varying CO2 mole fractions for blends of dimethyl ether + CO2, propylene + CO2, and R1234yf + CO2. This figure depicts that with increasing CO2 concentration, the mixture critical temperature decreases, and the critical pressure increases at lower CO2 concentration and then decreases at higher CO2 concentrations. In Figure 5 legends show the ‘HC’ used in HC + CO2 mixture. If the state of the refrigerant is having both the temperature and pressure above the critical point, then the refrigerant is said to be in supercritical state. This is vital as the properties of state 3 with respect to critical properties in a VCRC, determines if the cycle is sub-critical or trans-critical.

18

Table 5 Properties and mixture data availability with CO2 in REFPROP for ASHRAE designated mixture refrigerants containing hydrocarbons

ASHRAE Mixture composition (mass fraction) Tc Mixture data GWP

designation (K) availability with CO2 (100yr) R-432A Propene + dimethyl ether (0.8 + 0.2) 370 Yes 1.6 R-433A Propane + propene (0.7 + 0.3) 368 Yes 2.9 R-433B Propane + propene (0.95 + 0.05) 370 Yes 3.2 R-433C Propane + propene (0.75 + 0.25) 368 Yes 2.9 R-436A Propane + isobutane (0.56 + 0.44) 389 Yes 3.2 R-436B Propane + isobutane (0.52 + 0.48) 391 Yes 3.1 R-441A Ethane + propane + isobutane + butane 392 Yes 3.6 (0.031 + 0.548 + 0.06 + 0.361) R-443A Propene + propane + isobutane 369 Yes 3.0 (0.55 + 0.4 + 0.05) R-510A Dimethyl ether + isobutane 399 Yes 1.2 (0.88 + 0.12) R-511A Propane/dimethyl ether (0.95 + 0.05) 370 Yes 3.2

Figure 5 Critical properties of different mixtures with varying CO2 mole fraction

19

PART 1: THERMODYNAMIC ANALYSIS (PERFORMANCE STUDY)

3 Constant Energy Transfer Fluid Temperature Method

This chapter evaluates the pure refrigerants and mixtures of HCs and CO2 in a VCRC with constant temperature of energy transfer fluid (ETF) in both evaporator and condenser/gas cooler. The first part of this chapter shows the analysis of different pure refrigerants, which include the model, equations, and performance parameters like coefficient of performance, volumetric refrigeration capacity, and non- dimensional irreversibilities. The second part of this chapter shows the analysis of different mixtures with

CO2 and their results. At the end of this chapter, the refrigerants are refined based on the performance and the selected refrigerants from this chapter are used for further analysis in chapter 4.

3.1 Analysis of Pure Refrigerants

Thermodynamic analysis of pure refrigerants, which resulted from screening steps, was initially conducted to evaluate their performance parameters, before discussing their blends with CO2. Assumptions for the analysis of a simple vapor compression refrigeration cycle as shown in Figure 4 are : • Isobaric evaporation and condensation/gas cooling, • Compressor isentropic efficiency of 80 %, • Isenthalpic expansion, • Constant ETF temperature in both evaporator and condenser/gas cooler, • Steady-state conditions, • Negligible energy (heat) losses to the environment from all components, • Unit mass flow rate of refrigerant throughout the cycle. Some of these assumptions are depicted in the temperature-entropy diagram of a sub-critical VCRC using a pure refrigerant, as shown in Figure 6. In this figure the solid line portrays refrigerant states, dashed line depicts energy transfer fluid states, and the dotted line is saturation curve. The simple cycle discussed in this study is semi-ideal and does not consider many of the practical scenarios like: ➢ Volumetric efficiency and electric-motor efficiency of the compressor, ➢ Pressure drop in the energy (heat) exchangers, suction line, and discharge line, 20

➢ Auxiliary power for fan and controls, ➢ Superheating and subcooling Considering the aforementioned parameters into the method will make the analysis more real and practical. The study performed in this thesis is to compare the refrigerant mixtures from pure, and other mixtures and to determine the best compositions with high performance parameters.

Figure 6 Temperature – entropy diagram of a sub-critical simple VCRC using pure refrigerant

The method for determining the states of a refrigerant and performing thermodynamic analysis for a VCRC with constant ETF temperature is shown in Figure 7, as a flowchart. This method assumes that the ETF temperatures in evaporator and condenser/gas cooler, the evaporator outlet temperature, and the condenser/gas cooler outlet temperature are fixed. In this approach, the temperature in condenser/gas cooler at state-3, the saturation temperature in the evaporator at state-1, and the refrigerant used in VCRC are the key inputs. After defining the temperatures, the other states are calculated based on the type of cycle, i.e., sub-critical cycle or trans-critical cycle. The selection of saturation states gives lower (evaporator) and higher (condenser) pressures for a sub-critical VCRC with constant ETF temperature. In a sub-critical cycle, state 4 is calculated by isobaric assumption (4-1) and isenthalpic assumption (3-4), and state 3 is calculated by isobaric assumption (2-3) and isentropic compressor efficiency assumption (1-2).

For a trans-critical cycle, state 3 is not saturated liquid, and therefore, the discharge pressure from the compressor is optimized such that the cycle attains the maximum COP. The optimization is performed by calculating the change in COP by varying the discharge pressures (P2-3) from compressor. Figure 8 shows

COP as a function of compressor discharge pressure for different CO2 concentrations in propylene + CO2 mixtures. It can be seen that first COP increases as a function of pressure and then it decreases as reported in the literature [18; 19]. The corresponding pressure where the COP has the highest value is considered for calculations. 21

Figure 7 Flow chart of the thermodynamic model using Constant ETF temperature method

Figure 8 COP as a function of compressor discharge pressures in trans-critical cycles 22

After determining the state points either of a sub-critical cycle or trans-critical cycle, performance parameters for pure refrigerants like coefficient of performance (COP), volumetric refrigeration capacity (VRC), exergetic efficiency and non-dimensional irreversibilities (IRR) are calculated.

Coefficient of performance (COP) is a dimensionless quantity which determines the performance of a refrigeration cycle, it is defined as the ratio between refrigeration effect and energy (work) needed to run the compressor/cycle, as shown in equation (1):

푅푒푓푟𝑖푔푒푟푎푡𝑖표푛 퐸푓푓푒푐푡 (푅퐸) ℎ − ℎ 퐶푂푃 = = 1 4 (1) 퐸푛푒푟푔푦 푟푒푞푢𝑖푟푒푑 푡표 푟푢푛 푉퐶푅퐶 ℎ2 − ℎ1

Where, ℎ is the specific enthalpy of the refrigerant and the subscripts 1, 2, 3 and 4 indicate different states.

Volumetric refrigeration capacity (VRC) is defined as the refrigeration effect per unit volume of vapor entering the compressor, as shown in equation (2):

푅푒푓푟𝑖푔푒푎푟푡𝑖표푛 퐸푓푓푒푐푡 푉푅퐶 = = (ℎ − ℎ )/휈 (2) 퐶표푚푝푟푒푠푠표푟 𝑖푛푝푢푡 푠푝푒푐𝑖푓𝑖푐 푣표푙푢푚푒 1 4 1

Where, 푣1 is the specific volume of the refrigerant at state 1.

VRC has a significant influence on the physical size of the system, and the larger value of VRC corresponds to a compact system. For a system with high VRC, there will be smaller pressure drop, as reported in reference [34]. Smaller pressure drop, will indirectly improve the COP of a VCRC by reducing the pressure ratio, which reduces the energy needed to run the compressor. Mulroy et al. [24] mentioned that lower VRC refrigerants tend to show reduced COP as the compressor mechanical losses become a significant proportion of the total capacity [74]. Figure 9 shows GWP and VRC for different pure refrigerants. From this figure, it can be seen that hydrocarbons have low GWP and generous VRC, when compared to HFCs and HCFCs. Figure 9 shows that HFOs have very low GWP as HCs, but their VRCs are also very low. This figure displays that pure CO2 has the highest VRC and lowest GWP. Figure 9 shows the dotted lines at GWP = 100 and VRC = 2.5 MJ/m3 to define a good zone at top left quadrant of the figure.

Exergetic efficiency (휂푒푥) can be defined as the ratio between the energy output (refrigeration effect) of an actual cycle to the energy output of a reversible cycle when the energy input is same for both the cycles, as shown in the equation (3). The exergetic efficiency can also be written as the ratio of COP of an actual cycle to the COP of a reversible cycle. A reversible cycle is a cycle with the maximum cooling effect, for a given energy input while satisfying the laws of thermodynamics. Exergetic efficiency determines the effect of irreversibilities on the cycle performance. For a pure refrigerant, the reversible cycle is the Carnot cycle.

퐶푂푃푎푐푡푢푎푙 푞푎푐푡푢푎푙 휂푒푥 = = (푤ℎ푒푛 푒푛푒푟푔푦 𝑖푛푝푢푡 𝑖푠 푠푎푚푒) (3) 퐶푂푃푟푒푣푒푟푠푖푏푙푒 푞푚푎푥푖푚푢푚 23

Figure 9 Volumetric refrigeration capacity and GWP of different pure refrigerants

Table 6 shows COP, VRC, and exergetic efficiency for the selected 15 pure refrigerants from chapter 2 and the HFC refrigerant, R134a, as a reference to the currently used HFCs. These parameters are shown for two different sets of operating conditions and are given as follows:

(1) 푇3 = 50 °퐶 푎푛푑 푇1 = 10 °퐶

(2) 푇3 = 30 °퐶 푎푛푑 푇1 = 0 °퐶

Where, 푇1 is the saturation temperature at state 1 and 푇3 is the temperature at state 3. In this article, the uppercase ‘T’ indicates the temperature of refrigerant and the lowercase ‘t’ indicate the temperatures of the ETF.

The first set of operating conditions represent the AC application in tropical climates, and the second set represents food refrigeration application in household/commercial systems. These two sets of operating conditions are used for all the analysis in chapter 3. Table 6 lists that CO2 has the lowest COP and highest

VRC, which implies that VRC of the blends of HCs with CO2 decreases when compared to CO2 and increases when compared to HCs. Among the HCs, COP does not change significantly within each other, and VRC decreases with the increase in the number of carbons in HCs. On an average, the COP of HCs is almost (or) more than twice the value of the COP of pure CO2. The COP values are higher for operating conditions of the set (2) as the saturation temperature at state 3 is lower when compared to operating conditions of set (1).

The irreversibilities (𝑖) for the four components in the simple cycle can be calculated by using the entropy balance Comparison of the exergy loss caused by each component in the cycle can be made using the following non-dimensional irreversibilities. The non-dimensional irreversibility due to compression is given by,

푇 (푠 − 푠 ) 𝑖 = 표 2 1 (4) 푐표푚푝 푤 24

Where, 푇표 is the environment temperature, 푠2 is the specific entropy of the refrigerant after the compression, 푠1 is the specific entropy of the refrigerant before compression, and 푤 is the energy needed to run the compressor.

Table 6 Performance comparison of different pure refrigerants for the two operating conditions

No Refrigerant COP COP VRC (1) VRC (2) 휼풆풙 (1) 휼풆풙 (2) 푴푱 푴푱 (1) (2) ( ) ( ) (%) (%) (Operating conditions) 풎ퟑ 풎ퟑ

1 Carbon dioxide 1.92 3.61 13.15 12.34 27.1 39.6 2 Dimethyl ether 4.74 6.5 2.58 2.11 66.9 71.1 3 Propylene 4.30 6.15 4.13 3.70 60.8 67.6 4 Propane 4.33 6.17 3.43 3.06 61.1 67.8 5 R1234yf 4.15 6.04 2.42 2.17 58.7 66.4 6 R1234ze(E) 4.41 6.23 2.00 1.70 62.3 68.5 7 1-Butene 4.74 6.50 1.32 1.06 67.0 71.3 8 Isobutene 4.71 6.46 1.34 1.08 66.5 70.1 9 Butane 4.70 6.45 1.07 0.86 66.3 70.9 10 Isobutane 4.57 6.35 1.44 1.21 64.5 69.8 11 Diethyl ether 4.80 6.53 0.36 0.26 67.8 71.7 12 Pentane 4.80 6.53 0.34 0.25 67.8 71.7 13 Isopentane 4.75 6.50 0.44 0.33 67.1 71.3 14 Neopentane 4.57 6.35 0.73 0.60 64.6 69.8 15 Novec 649 3.96 5.82 0.15 0.12 55.9 63.9 16 R134a* 4.43 6.26 2.68 2.26 62.6 68.7

‘*’ Commonly used HFC refrigerant R134a, CH2FCF3 (shown here for comparison)

The non-dimensional irreversibility due to condensation/gas cooling is given by,

푇2 훿푞푐 푞푐 푇표 {(푠3 − 푠2) + ∫ } 푇표 {(푠3 − 푠2) + } 푇3 푡 푡 𝑖 = = 2 (5) 푐표푛푑/푔푐 푤 푤

Where, 푠3 is the specific entropy of the refrigerant after condensation/gas cooling, and 푞푐 is the energy (heat) transferred to the ETF in the condenser/gas cooler. For constant ETF temperature system, 푡2 =

푡3 as seen in Figure 6.

25

The non-dimensional irreversibility due to evaporation is given by,

푇4 훿푞푒 푞푒 푇표 {(푠1 − 푠4) − ∫ } 푇표 {(푠1 − 푠4) − } 푇1 푡 푡 𝑖 = = 3 (6) 푒푣푎 푤 푤

Where, 푠4 is the specific entropy of the refrigerant before evaporation, and 푞푒 is the energy (heat) gained from the ETF in the evaporator. For constant ETF temperature system, 푡1 = 푡4 as seen in Figure 6.

The non-dimensional irreversibility due to expansion is given by,

푇 (푠 − 푠 ) 𝑖 = 표 4 3 (7) 푒푥푝 푤

Reference [27] shows a similar approach as seen in this article for calculating irreversibilities. The aforementioned equations are used to calculate the non-dimensional irreversibilities for selected pure refrigerants and is shown in Figure 10. In this study, the constant ETF temperatures are assumed as t1 = T1 +

7 and t3 = T3 - 7 like reference [27], but it needs to be noted that the temperature difference between the refrigerant and ETF depends on the design of the energy (heat) exchangers. It is seen from Figure 10, that the highest non-dimensional irreversibility is in the compressor and then in the expansion valve for all refrigerants except CO2. In the case of CO2, the expansion valve has the highest non-dimensional irreversibility due to the high discharge pressures after compressor (P2-3). All the refrigerants other than CO2 have nearly the same total non-dimensional irreversibility. The contribution of different components to the total non-dimensional irreversibility changes with refrigerant.

Figure 10 Non-dimensional irreversibilities for different components using pure refrigerants at operating conditions of set (1); The numbers indicate the position of the refrigerant in Table 6 26

3.2 Analysis of Zeotropic Mixtures

A mixture of two or more single compound fluids can be classified into three types based on their boiling points and dew points. The three types are, zeotropic mixtures, near-azeotropic mixtures, and azeotropic mixtures. A zeotropic mixture is a mixture of two or more single compound fluids having different boiling points. A temperature composition diagram of a binary zeotropic mixture of HC and CO2 is shown in Figure 11. Dew curve and boiling curve in this figure is drawn as a locus of dew points and boiling points at different mixture concentrations. The dew curve and boiling curve of a zeotropic mixture do not intersect at any location other than the ends, as shown in Figure 11. An azeotropic mixture also have different pure boiling points, but its dew and boiling curves intersect at least one point other than the ends. Examples of azeotropic mixture are R-500 (mixture of R-12 {CCl2F2} and R-152a {C2H4F2}) and R-503 (mixture of R-23{CHF3} and R-12{CCl2F2}). A near-azeotropic mixture is when the difference between the boiling points and the dew points along the change in composition is very less, but it does not intersect. Most mixtures behave as zeotropic mixtures when compared to azeotropic mixtures and near-azeotropic mixtures. Didion and Bivens [75] showed a very detailed explanation on the classification of types of mixtures, and the features of zeotropic mixtures.

Figure 11 Temperature-composition diagram for a zeotropic mixture of HC + CO2

Blends of hydrocarbons and CO2 are zeotropic mixtures with a significant difference between the boiling points of pure substances. One of the differentiating features of zeotropic mixtures from an azeotropic mixture or a pure refrigerant is that the phase change occurs at a varying temperature in the former. The temperature glide (TG) for a zeotropic mixture is defined as the difference between boiling point and dew point at a constant pressure during the phase change for a given composition. Figure 11 also shows that temperature glide has a maximum value at a specified composition and converges to zero for pure refrigerants. 27

Another important feature of zeotropic mixtures is composition shift. It can be observed in the two- phase region of the cycle, due to the differential hold up of individual component relative to its volatility, i.e. there will be a difference in mixture concentrations at different stages of the VCRC. More study on the effect of temperature glide and composition shift on energy (heat) exchangers on the cycle is performed by Leelananda Rajapaksha [43]. The composition shift is experimentally studied by Kim et al. [35].

To perform thermodynamic analysis of a VCRC using zeotropic mixtures, a similar method to the pure refrigerants was used, except that phase change in the evaporator and condenser are not isothermal. The TG of a zeotropic mixture in a VCRC can be seen in a temperature - entropy diagram, like Figure 12. This figure shows a sub-critical cycle using zeotropic mixture as a refrigerant with constant ETF temperature in the energy exchangers. In this figure the solid line depicts refrigerant states, dashed line depicts energy transfer fluid states, and dotted line is the saturation curve. It can be seen from this figure that the temperature changes non-linearly with entropy in the energy (heat) exchangers. Figure 13 shows the temperature-entropy diagram of a trans-critical cycle using zeotropic mixture as a refrigerant with constant ETF temperature in the energy exchanger. It is seen from this figure that the non-linearity of temperature with entropy in the gas cooler is dominant than the non-linearity of temperature with entropy in the evaporator. From Figure 12 and Figure 13, it is observed that non-linearity of temperature with entropy is significant in trans-critical cycle than that of the sub-critical cycle. A VCRC with a constant ETF temperature, having substantial non-linearity in temperature with entropy tends to have lower COP due to higher irreversibilities in the energy exchangers.

Figure 12 Temperature-entropy diagram of sub-critical cycle using zeotropic mixture as refrigerant

Performance parameters like COP and VRC for all the selected 14 refrigerants mixtures from section 2 was calculated, but only the best four HC + CO2 blends, with high-performance parameters, are plotted. The best four mixtures are dimethyl ether + CO2, propylene + CO2, propane + CO2, and isobutane + CO2. Among the 14 different refrigerant mixtures, pentane + CO2, isopentane + CO2, diethyl ether + CO2, and neopentane + CO2 have the lowest performance parameters respectively in both sub-critical and trans-critical 28

cycles. Among the mixtures of butane + CO2, isobutane + CO2, isobutene CO2, and 1-butene + CO2 higher performance parameters have been observed for isobutane + CO2 mixtures.

Figure 13 Temperature-entropy diagram of trans-critical cycle using zeotropic mixture as refrigerant

Figure 14 shows the values of COP as a function of CO2 mole fraction for the best four HC + CO2 mixtures at operating conditions of set (1). This figure displays that COP decreases with the addition of CO2 in the mixture concentrations. For all the trends in the figure, a dip at higher concentrations of CO2, indicates the VCRC transition from sub-critical to trans-critical. Among the trends in Figure 14, mixtures of dimethyl ether + CO2 have the highest COP, followed by mixtures of propylene + CO2 and propane + CO2 for the sub- critical part of the trends. For the trans-critical part, mixtures of propylene + CO2 has the highest COP, followed by mixtures of propane + CO2, and isobutane + CO2.

Figure 15 shows values of VRC as a function of CO2 mole fraction for the best four HC + CO2 mixtures at the operating conditions of set (1). This figure depicts that VRC increases with the addition of

CO2 for all the mixtures. A similar dip, as seen in Figure 14, is observed in this figure, demonstrating the transition of the VCRC from sub-critical to trans-critical. Among the trends, mixtures of propylene + CO2 has the highest VRC, followed by mixtures of propane + CO2, and dimethyl ether + CO2 in the subcritical part of the trends. For the trans-critical part of the trends, mixtures of propylene + CO2 has highest VRC, followed by mixtures of propane + CO2, and isobutane + CO2.

Performance parameters for a VCRC at operating conditions of set (2) are calculated and plotted in similar to plots of a VCRC at operating conditions of set (1). Figure 16 shows values of COP as a function of CO2 mole fraction for the best four HC + CO2 mixtures at operating conditions of set (2). This figure displays a similar pattern of COP trends as seen in Figure 14, with mixtures of dimethyl ether + CO2 having the highest COP at the lower concentrations of CO2 and mixtures of propylene + CO2 having the highest

COP at higher concentrations of CO2. Figure 17 shows values of VRC as a function of CO2 mole fraction for the best four HC + CO2 mixtures at operating conditions of set (2). This figure has the similar pattern as that 29

of Figure 15, where mixtures of propylene + CO2 has the highest VRC for all concentrations of CO2. These figures illustrate that COP for operating conditions of set (2) are higher than set (1), as discussed in Table 6. The trends in Figure 16 and Figure 17 does not have a significant dip, showing the transition of sub-critical to trans-critical as seen in Figure 14 and Figure 15. This is because for all concentrations of CO2 the VCRC is sub-critical for operating conditions of set (2).

Figure 14 COP of different refrigerant mixtures for varying CO2 mole fractions at set (1) operating conditions

Figure 15 VRC of different refrigerant mixtures for varying CO2 mole fractions at set (1) operating conditions 30

Figure 16 COP of different refrigerant mixtures for varying CO2 mole fractions at set (2) operating conditions

Figure 17 VRC of different refrigerant mixtures for varying CO2 mole fractions at set (2) operating conditions

Figure 18 shows the higher pressures in condenser/gas cooler for varying CO2 concentrations at operating conditions of set (2). It is seen from this figure that pressures increase with the addition of CO2 concentration in the mixture. Higher pressures indicate a more complex system leading to an increase in the capital and maintenance cost. It also means that the system is more likely to leak the refrigerant into the environment at compressor and condenser when compared to lower pressure cycles, as maintaining high 31

pressures in the systems is tough. It is also seen that higher pressures have a direct effect on the non- dimensional irreversibility of the expansion valve, which will be discussed later in this section.

Figure 18 Condenser pressures for varying CO2 mole fraction at operating conditions of set (2)

Performance parameters like COP and VRC of the CO2 mixtures with two selected HFOs (R1234yf and R1234ze(E)) from section 2 are plotted. Mixtures of CO2 + Novec 649 are not shown since they have very low performance parameters when compared to other mixtures. Figure 19 shows values of COP as a function of CO2 mole fraction for HC + CO2 and HFO + CO2 mixtures at operating conditions of set (1). It illustrates that COP values for mixtures of R1234yf + CO2 and R1234ze(E) + CO2 are comparable to the mixtures of dimethyl ether + CO2 and propylene + CO2. Figure 20 shows values of VRC as a function of CO2 mole fraction for HC + CO2 and HFO + CO2 mixtures at the operating conditions of set (1). Similarly, in this figure VRC of mixtures of R1234yf + CO2 and R1234ze(E) + CO2 are comparable to the VRC of the mixtures of dimethyl ether + CO2, but propylene + CO2 mixtures have a considerable advantage in VRC. Bearing in mind the low flammability of HFOs, lower carbon dioxide concentrations in the mixtures would be beneficial with a slight increase in VRC when compared to those of pure HFOs.

The COP for propane + CO2 and isobutane + CO2 mixtures was calculated using CYCLE_D (NIST) [19] program to compare with the model developed in this study. Both models follow the same approach of assuming temperatures at the end of condenser/gas cooler and at the end of evaporator. Figure 21 shows the values of COP as a function of CO2 concentration using both CYCLE-D (NIST) and the model developed in this study at the operating conditions of set (1). It is seen from the figure that both the models agree with each other, except at the transition point of the sub-critical to trans-critical cycle. At this point, CYCLE-D assumes an uncalled pressure rise after compression without any energy input (work), which is not possible. So, at this transition point, CYCLE-D results cannot be considered due to some error in the software. 32

Figure 19 COP of HC + CO2 and HFO + CO2 mixtures for varying CO2 mole fractions at set (1) operating conditions

Figure 20 VRC of HC + CO2 and HFO + CO2 mixtures for varying CO2 mole fractions at set (1) operating conditions

The exergetic efficiencies for zeotropic mixtures are calculated using Lorenz cycle instead of Carnot cycle, which considers temperature glide in reversible cycle during phase change as shown in Figure 22 [76].

In Lorenz cycle, the temperature glide in both the energy (heat) exchange processes is the same, i.e. 푇1 −

푇4 = 푇2 − 푇3, and the expansion process (3-4) & compression process (1-2) are isentropic. In the real case evaporation and condensation/gas cooling have different pressures and, the evaporation does not start from the saturated liquid. So, a modified Lorenz cycle (reversible) is used, where 푇1 − 푇4 = 퐷푇푒 < 푇2 − 푇3 =

퐷푇푐. Where 퐷푇푒 푎푛푑 퐷푇푐 are calculated using the assumption of isobaric condensation and evaporation 33

along with isentropic compression and expansion. Modified Lorenz cycle is shown in Figure 23 and a similar approach, to modified Lorenz cycle is seen in reference [77]. For this cycle to be possible, it requires infinite energy (heat) exchangers and energy (heat) source/sink.

Exergetic efficiency for zeotropic mixtures is calculated using equation (8), where reversible cycle here is modified Lorenz cycle.

퐶푂푃푎푐푡푢푎푙 푞푎푐푡푢푎푙 휂푒푥 = = (푤ℎ푒푛 푒푛푒푟푔푦 𝑖푛푝푢푡 𝑖푠 푠푎푚푒) (8) 퐶푂푃푚표푑푖푓푖푒푑 퐿표푟푒푛푧 푞푟푒푣푒푟푠푖푏푙푒

Figure 21 Comparison of COP for propane + CO2 and isobutane + CO2 mixtures with ‘CYCLE-D’ and ‘this study’

Figure 22 Temperature entropy diagram of Lorenz cycle [76] 34

Figure 23 Temperature entropy diagram of modified Lorenz cycle

To determine the exergy loss contribution by individual component in the cycle, the non- dimensional irreversibilities are calculated. For zeotropic mixtures the equations are same as (4)-(7), like pure refrigerant analysis (section 3) with 푡1 = 푇1 + 7 and 푡3 = 푇3 − 7. Figure 24 shows the non-dimensional irreversibility of different components for the increasing CO2 concentration in propylene + CO2 mixture. In

Figure 24, the irreversibility of the expansion valve is highest among other components for all CO2 concentrations and is observed to increase with an increase in CO2 concentration. Similarly, non-dimensional irreversibilities of different components for the increasing CO2 concentration in dimethyl ether + CO2 mixture is shown in Figure 25. In this figure, the condenser seems to have the highest irreversibility for CO2 concentrations less than 80% (sub-critical cycle), and for the CO2 concentration of 80% (trans-critical cycle) the expansion valve seems to have the highest non-dimensional irreversibility.

Figure 24 Non-dimensional irreversibilities for different components at set (1) operating conditions 35

.

Figure 25 Non-dimensional irreversibilities for different components at set (1) operating conditions

Further study using with an additional component of liquid-line/suction-line energy (heat) exchanger (LL/SL-EX) to a simple VCRC was performed to see the effect on performance parameters. Figure 26 shows the schematic of a VCRC with LL/SL-EX. The temperature-entropy diagram of this cycle is shown in Figure 27, which is using pure refrigerant. For the thermodynamic analysis, an energy (heat) exchanger effectiveness (휀) of 0.8 is used for the LL/SL-EX. The COP for this cycle calculated using selected HC +

CO2 mixtures for varying CO2 is shown in Figure 28, using the operating conditions of the set (1). Similar trends of COP to the simple VCRC without LL/SL-EX, as shown in Figure 14 is observed.

Figure 26 Schematic of a VCRC with liquid-line/suction-line energy exchanger 36

Figure 27 Temperature entropy diagram of a VCRC with LL/SL-EX using pure refrigerant

Figure 28 COP of different refrigerant mixtures for varying CO2 mole fractions in a VCRC with LL/SL-EX at set (1) operating conditions

Similarly, like a LL/SL-HX addition to a simple cycle, a liquid-vapor separator can be added for a two-stage compression cycle, whose schematic is shown in Figure 29. Figure 30 shows the comparison of

COP for propane + CO2 mixture with varying CO2 concentration. This figure compares simple cycle, LL/SL- EX cycle, and 2-stage compression cycle at operating conditions of set (1). From, Figure 30 it is seen that the 2-stage compression cycle has the highest COP followed by LL/SL-EX cycle. The difference of COP between simple cycle and LL/SL-EX cycle is minimal when compared to either of its difference with 2-stage cycle. Addition of such devices does increase COP but comes with an increase in capital cost and an increase in the complexity of the cycle. 37

Figure 29 Schematic of a VCRC with 2-stage compression

Figure 30 COP of propane + CO2 mixture for varying CO2 mole fraction in different cycles

Addition of other components to the simple VCRC can increase the performance of the overall cycle. This chapter showed a couple of such components, i.e. LL/SL-EX and two-stage compression. Sarkar et al. [78] reviewed and compared different modifications to a simple trans-critical cycle such as adding an internal energy (heat) exchanger, expansion turbine, 2-stage compression, ejector-expansion, vortex-tube expansion, and parallel compression economization. He showed that adding an expansion turbine to the simple cycle has the highest COP improvement among all others mentioned earlier. Other relevant techniques to increase COP include mass flow rate change of the ETF to match the non-linear temperature profile of the refrigerant. This technique is experimentally shown in reference [79].

38

Chapter 2 and 3 conclusions:

Different pure refrigerants available in REFPROP 10 were evaluated using specific screening parameters to identify viable refrigerants for binary mixtures with CO2. These steps resulted in 14 different mixtures with

CO2. The thermodynamic analysis was performed using pure refrigerants and the zeotropic mixtures with

CO2 in the vapor compression refrigerant cycle with constant energy transfer fluid temperature in both evaporator and condenser/gas cooler. The analysis was performed for different operating conditions to represent air-conditioning and food refrigeration applications. Non-dimensional irreversibilities of different components of a vapor compression refrigeration cycle were determined to evaluate their exergy loss contribution.

• Results show that the COP decreases and VRC increase with increasing concentrations of CO2 in the

mixture. On an average, for HC + CO2 , a change in CO2 concentration from 0% to 20% reduced the

COP by 35% and a change in CO2 concentration from 60% to 80% reduced the COP by 15%. For HC

+ CO2 mixtures, a change in CO2 concentration of 20% in the sub-critical cycle increased the VRC by

12% and the same change in CO2 concentration in a trans-critical cycle increased the VRC by 40%.

• Among the HC mixtures, dimethyl ether + CO2 mixtures have highest COP for lower CO2

concentrations, and propylene + CO2 has highest COP for higher CO2 concentrations.

• The best four refrigerant mixtures in terms of performance are propylene + CO2, dimethyl ether + CO2,

propane + CO2, and isobutane + CO2. Among the HFO mixtures, R-1234yf + CO2 and R-1234ze(E) +

CO2 show comparable performance parameters with other mixtures. It is also observed that operating conditions of the set (2) has a higher COP than operating conditions of the set (1).

• The total non-dimensional irreversibilities is observed to increase with the increase in CO2 concentrations. An increase of expansion valve contribution to the total exergy loss is observed with

increasing CO2 concentration.

39

4. Variable Energy Transfer Fluid Temperature Method

4.1 Temperature Glide Matching Models

For a fair comparison between the proposed zeotropic refrigerant mixtures and current refrigerants, it is necessary to match the temperature glide of the zeotropic mixtures with energy transfer fluid (ETF) temperature profile in both energy (heat) exchangers. Use of counter flow or cross-counter flow energy (heat) exchangers will align the refrigerant and ETF temperature profiles and reduces the irreversibilities in the energy exchangers, which enhances the performance as seen in the literature review section of chapter 1. This method of aligning the refrigerant and ETF temperature profiles is called “glide matching”. Due to the non-linear change of temperature in energy (heat) exchangers, simple methods like logarithmic mean temperature difference (LMTD) cannot be used for the thermodynamic analysis of zeotropic mixtures.

Figure 31 and Figure 32 shows the difference between the two methods of performing thermodynamic analysis with the help of temperature - entropy plots. Figure 31 depicts the thermodynamic analysis using constant ETF temperature method, as shown in chapter 3 and Figure 32 depicts the variable ETF temperature method. The closer the gap between the ETF and the refrigerant temperature profiles, the lesser the irreversibility and higher the COP.

Figure 32 and Figure 33 show the temperature - entropy plots for zeotropic and pure refrigerants respectively along with variable ETF temperature profiles. In pure refrigerant cycle, one can estimate that the ∆푇푎푣푔 (average temperature difference between the refrigerant and ETF for both the energy (heat) exchangers is greater than the ∆푇푎푣푔 of the zeotropic refrigerant cycle.

Figure 31 Temperature-entropy diagram of a VCRC with constant ETF temperature using zeotropic mixture as refrigerant 40

Figure 32 Temperature-entropy diagram of a VCRC with variable ETF temperature using zeotropic mixture as refrigerant

Figure 33 Temperature-entropy diagram of a VCRC with variable ETF temperature, using pure refrigerant

The two approaches for performing thermodynamic analysis as discussed in chapter one are NIST model and minimum (DT) model. In this section NIST model is discussed first and briefly. The minimum (DT) model is explained in detail and the logic involving the thermodynamic analysis is discussed in this section.

NIST model:

The inputs for a simple cycle include refrigerant mixture composition, ETF temperature profiles for both condenser and evaporator, net system cooling capacity (Qe), mean effective temperature difference ΔThx or the overall energy (heat) transfer conductance (UAhx) in the evaporator and condenser. It also needs to be noted that the NIST model does not work for trans-critical cycles and so, the cycle should be sub-critical only. The energy (heat) transfer between the refrigerant and the ETF in the evaporator is given by equation (9), 41

푄푒 = 푄ℎ푥 = 푈퐴ℎ푥(∆푇ℎ푥) (9)

Where, ∆푇ℎ푥 is a harmonic mean weighted with a fraction of energy (heat) transferred in individual sections, where universal energy (heat) transfer coefficient (푈) for each section is same.

푄1 푄2 1 푄푖 ∆푇ℎ푥 = + … = ∑ (10) 푄ℎ푥∆푇1 푄ℎ푥∆푇2 푄ℎ푥 ∆푇푖

Now, the model bisects the energy (heat) exchanger until the ∆푇ℎ푥 values obtained from two consecutive evaluations agree with each other to a convergence input parameter. CYCLE-D_HX, a NIST software performs the thermodynamic analysis using the NIST model and have the options to include real parameters like pressure drop, compressor mapping, superheating, subcooling, auxiliary fan losses, energy (heat) exchanger type, dimensions and circuits. The main drawback for this model is that the value for

∆푇ℎ푥 표푟 푈퐴ℎ푥 need to be obtained by experiments only. Alternatively, an educated guess is needed for these values, this brings us to the second method, as outlined below.

Minimum (DT) model:

In this approach, a script in MATLAB is written utilizing the logical operators such that the minimum temperature difference between the refrigerant and the ETF is fixed (usually greater than 5). Inputs for a simple cycle in this method include: refrigerant mixture composition, ETF temperature profiles for both condenser/gas cooler and evaporator, minimum temperature difference between the refrigerant (ΔTm) for both condenser/gas cooler and evaporator. In the literature a minimum temperature difference of 7°C is observed in [27] and 5°C is observed in [26]. The minimum temperature difference depends on the energy (heat) exchanger design, so in this analysis we selected 5°C ,7°C and 10°C. The energy exchanger temperature profile was divided into 100 parts (or more) based on enthalpy difference. The state points of the refrigerant are changed such that the minimum temperature difference between the refrigerant and the ETF is ΔTm. Using this approach, we can calculate the efficiencies, ΔThx using equation (10), and several other parameters by having less input conditions than the NIST model. This model does not require any experimental input. It can be used to get an educated guess for ∆푇ℎ푥 for its use in NIST model. From Figure 33, it is seen that for pure refrigerant ∆푇푚 for the evaporator (4 - 1) is always at state 4 and for the condenser (2 - 3) it is always at the vapor-liquid saturation point (dew point). Whereas the ∆푇푚 for the zeotropic mixture cycle keeps changing its position based on CO2 concentration and operating conditions.

Many assumptions were made for the VCRC using the variable ETF temperature method with minimum (DT) model and are shown below. These assumptions are similar to the constant ETF temperature method shown in chapter 3.

➢ Isobaric evaporation and condensation/gas cooling, ➢ Compressor isentropic efficiency of 80 %, ➢ Isenthalpic expansion, 42

➢ Constant specific heat for energy (heat) transfer fluid (ETF) in both evaporator and condenser/gas cooler, ➢ Steady-state conditions, ➢ Negligible energy (heat) losses to the environment from all components, ➢ Unit mass flow rate throughout the cycle.

Figure 34 shows the logic flowchart for determining the states of the refrigerant and performing the thermodynamic analysis using the minimum (DT) model. In this figure and in further sections of this chapter the lowercase alphabets indicate ETF properties and uppercase alphabets indicate refrigerant properties.

Among the input conditions, 푡1 is the inlet temperature and 푡4 is the exit temperature of the ETF in the evaporator. In the condenser/gas cooler 푡2 is the inlet temperature and 푡3 is the exit temperature of the ETF.

Other than ETF temperatures the other input condition is the minimum temperature difference between refrigerant and ETF in both the energy (heat) exchangers, ∆푇푚푖푛. Figure 35 displays the input conditions, which shows the vapor saturation curve and ETF temperatures in the counterflow energy (heat) exchangers. Now, a sub-critical cycle is defined when (푡3 + min(퐷푇) < 푇푐) so that the state 3 can be a saturated state. For both sub-critical and trans-critical cycles, the first step is to guess the temperature of the ′ ′ refrigerant at state 1 and state 4 i.e. 푇1 푎푛푑 푇4 with an initial guess of 푇1 = 푡1 − ∆푇푚푖푛 푎푛푑 푇4 = 푡4 −

∆푇푚푖푛. Using these temperatures, state 4 can be determined utilizing the isobaric assumption of the refrigerant in the evaporator. Second step involves a quality check of state 4 as shown in Figure 36, where 푇1′′ is the change of the guess of 푇1 such that state 4 is a two-phase mixture. The first two steps in sub-critical and trans- critical cycle is same and the further steps vary for each type of the cycle. 43

Figure 34 Flowchart of variable ETF temperature method using minimum (DT) approach/model 44

Figure 35 Temperature - entropy diagram representing the variable ETF method with minimum (DT) approach (input conditions)

Figure 36 Temperature - entropy diagram representing the variable ETF method with minimum (DT) approach (quality check of state 4) Sub-critical cycle:

In sub-critical cycle state 3 can be calculated using the assumption of isenthalpic expansion process

(3 - 4) such that state 3 is saturated. If the state 3 temperature, 푇3 < 푡3 + ∆푇푚푖푛, the initial guess of ′ temperature 푇1 is changed such that 푇3 ≥ 푡3 + ∆푇푚푖푛 = 푇3. After determining state 3, state 2 can be calculated utilizing the assumption of isobaric condensation (2 - 3) and isentropic efficiency compression (1 ′′′ - 2). This step is shown in Figure 37, where 푇1 is the change of the guess of 푇1 to satisfy the condition. As it is seen in this figure, it was assumed that the phase change temperature is linear with entropy/enthalpy but in a real scenario, the temperature changes non-linearly. To consider this non-linearity the enthalpy difference between (2 - 3) and enthalpy difference between (1 - 4) is divided into ‘N’ (푁 ≥ 100) parts, where each part is assumed to have linear temperature change. After this division, each part is checked for their temperature difference, ∆푇 between the refrigerant and ETF. If any of the divisions have their ∆푇 < ∆푇푚푖푛, the temperature guesses of the refrigerant are changed such that ∆푇 > ∆푇푚푖푛 for all divisions. Similarly, if the 45

minimum ∆푇 among all the divisions is not ∆푇푚푖푛, then the guesses of refrigerant temperature is changed.

This process is repeated until min(∆푇푐표푛푑푒푛푠푒푟) = min(∆푇푒푣푎푝표푟푎푡표푟) = ∆푇푚푖푛. This process can be seen in ′′′′ ′′ Figure 38, where 푇1 denote the final temperature of state 1, 푇4′′ denote the final temperature of state 4, 푇3 ′ denote the final temperature of state 3 and 푇2 denote the final temperature of state 2. This logic is clearly shown in terms of flowchart in sub-critical part of Figure 34.

Figure 37 Temperature entropy diagram representing the variable ETF method (sub-critical cycle) with minimum (DT) approach and linear temperature change assumption

Figure 38 Temperature entropy diagram representing the variable ETF method (sub-critical cycle) with minimum (DT) approach and dividing temperature profile into N parts Trans-critical cycle:

After the steps shown in Figure 36, state 3 needs to be calculated. The only known property of state 3 is the enthalpy from isenthalpic expansion process (3 - 4). Since, it is not a saturated state in trans-critical cycle one other property is required to define state 3. Which would be a guess temperature for 푇3 = 푡3 +

∆푇푚푖푛. Now, a phase check for state 3 is performed such that it is a supercritical fluid and not in solid phase. 46

If state 3 is solid phase, the initial temperature guess of 푇1 is changed. After determining the state 3, state 2 can be calculated by the assumption of isobaric gas cooling (2 - 3) and isentropic efficiency compression (1 - 2). After, calculating state 2, the enthalpy difference in gas cooler and evaporator is divided into ‘N’ (푁 ≥ 100) parts to account for the non-linearity, like the sub-critical cycle. After the division, the initial guesses of refrigerant temperature are changed such that min(∆푇푔푎푠 푐표표푙푒푟) = min(∆푇푒푣푎푝표푟푎푡표푟) = ∆푇푚푖푛, similar to the sub-critical cycle analysis model. This logic is clearly shown in terms of flowchart in trans- critical part of Figure 34.

4.2 Performance Parameters

For the calculation of performance parameters, input conditions of ETF temperature are taken as, t3 = 45 + 273.15 K; condenser/gas cooler ETF inlet t2 = 60 + 273.15 K; condenser/gas cooler ETF outlet t1 = 30 + 273.15 K; evaporator ETF inlet t4 = 15 + 273.15 K; evaporator ETF outlet

The above conditions are used for figures in this chapter and are defined as operating conditions of set (3). Figure 39 shows the COP for different refrigerant mixtures for varying CO2 mole fraction with

∆푇푚푖푛 = 7. In this figure, it is seen that dimethyl ether + CO2 mixture has the highest COP for all CO2 concentration, except at the transition point of sub-critical to trans-critical. From this figure it is also seen that the highest COP for dimethyl ether and propylene mixtures is at a CO2 mole fraction of 0.15, for propane the maximum COP is at a CO2 mole fraction of 0.1 and for isobutane it is at a CO2 mole fraction of 0.05.

Figure 39 COP of different refrigerant mixtures for varying CO2 mole fractions using variable ETF method and minimum (DT) approach with ∆푇푚푖푛 = 7, at set (3) operating conditions 47

Figure 40 and Figure 41 shows the performance parameters for propylene + CO2 mixtures for different ∆푇푚푖푛푖푚푢푚 of 5, 7 and 10 using the operating conditions of set (3). Figure 40 shows the change of

COP with CO2 mole fraction and Figure 41 shows the change of VRC with CO2 mole fraction. It can be seen from this figure around CO2 mole fraction of 0.15, we have the maximum COP and the minimum COP is observed at pure CO2. VRC increases with addition of CO2 in the mixture until it reaches it maximum at pure

CO2. From this figure, it is observed that lower the ∆푇푚푖푛, higher the performance parameters are.

Figure 40 COP of different refrigerant mixtures for varying CO2 mole fractions using variable ETF method and minimum (DT) approach with different ∆푇푚푖푛 of 5, 7 and 10, at set (3) operating conditions

Figure 41 VRC of propylene + CO2 mixture for varying CO2 mole fraction using variable ETF method and minimum (DT) approach with different ∆푇푚푖푛 of 5, 7 and 10, at set (3) operating conditions

Figure 42 shows the COP for different operating conditions, which are mentioned in the table on the figure. It can be seen from this figure that the maximum COP over the range of CO2 concentration changes with operating conditions. It is seen from this study that zeotropic mixtures increase COP and VRC when compared to pure refrigerants, but they also have possible disadvantage as mentioned by [52] like system design changes, difficulty in charging and servicing, and compositions changes during an accidental leak. 48

Set t1 (°C) t2(°C) t3(°C) t4(°C) A 30 60 45 15 B 30 60 45 10 C 30 60 40 15

Figure 42 COP of propane + CO2 mixtures for varying CO2 mole fraction using variable ETF method and minimum (DT) approach with ∆푇푚푖푛 = 7, for different operating conditions

The irreversibilities for the variable ETF method can be calculated using the equations (4 - 7), where the temperature of the ETF is changing in the integral part of the non-dimensional irreversibility equations for condenser (equation 5) and evaporator (equation 6). Figure 43 shows the irreversibility for different CO2 concentrations for propane + CO2 mixture and at ∆푇푚푖푛 of 7 and operating conditions of set (3). It is seen from this figure that 15% CO2 has lowest irreversibility (highest COP) and 60% CO2 has the highest irreversibility (lowest COP). This figure agrees with COP change of propane with CO2 concentrations as shown in Figure 39.

Figure 43 Non-dimensional irreversibilities of different components for varying CO2 concentration in propane + CO2 mixture using variable ETF temperature method at set (3) operating conditions 49

Non-dimensional irreversibilities for different components with different ∆푇푚푖푛 is shown in Figure

44 for propylene CO2 mixture. It is seen from this figure that the irreversibility is lowest when ∆푇푚푖푛 is 5, which agrees with the COP shown in part (a) of Figure 40.

Figure 44 Non-dimensional irreversibilities of different components for different ∆푇푚푖푛 in propylene + CO2 (0.85 + 0.15) mixture using variable ETF temperature method at set (3) operating conditions

Figure 45 and Figure 46 show the temperature-entropy diagrams of propylene + CO2 mixtures at operating conditions of set (3). It is seen from Figure 45 that the process (2-3) is a combination of gas cooling and condensation. Figure 45 shows mainly condensation process and Figure 46 shows only gas cooling process. The cycle that behaves as shown in Figure 45 is defined as ‘semi trans-critical cycle’ in this thesis where the process (2-3) has a non-linear temperature change with entropy from state 2 to saturated vapor state. Similarly, for a complete sub-critical cycle as seen in Figure 32, the process (2-3) has a near linear temperature change with entropy from state 2 to saturated vapor state.

Figure 45 Temperature-entropy diagram of propylene + CO2 (0.4 + 0.6), at set (3) operating conditions 50

Figure 46 Temperature-entropy diagram of propylene + CO2 (0.2 + 0.8), at set (3) operating conditions

Figure 47 Comparison of COP for NIST and minimum (DT) model along CO2 mole fraction for propylene + CO2 mixture at set (3) operating conditions

Figure 47 compares the COP from the NIST model and minimum (DT) model. In minimum (DT) model the temperature profiles of both the energy (heat) exchangers is divided into N parts, it is done such 푄 푄 that each part has the same enthalpy difference (푑푄 = 푒 푎푛푑 푑푄 = 푐). This is applicable for both sub- 푒 푁 푐 푁 critical and trans-critical cycles. After attaining the states and ending the logic loop ∆푇1 푡표 ∆푇푁 can be calculated to obtain ∆푇ℎ푥 as per equation (10). This calculated ∆푇ℎ푥 can be used as an educational guess for the input conditions in NIST model. Using this we can compare the performance parameters calculates using the minimum (DT) model and the NIST model. From Figure 47, it can be seen that both the models agree with each other in sub-critical cycles. Whereas NIST model (CYCLE-D_HX) is not compatible with either semi trans-critical cycle or trans-critical cycle. Hence, COP cannot be compared for higher CO2 concentration. 51

Chapter 4 conclusion:

The selected best four refrigerant mixtures dimethyl ether + CO2, propylene + CO2, propane + CO2, and isobutane + CO2 from sections 3 were evaluated in this chapter using the variable ETF temperature method. Different models and their approaches were discussed and a novel approach or minimum (DT) model is developed. The thermodynamic analysis was performed using selected zeotropic mixtures with CO2 in the vapor compression refrigerant cycle with variable energy transfer fluid temperature in both evaporator and condenser/gas cooler. The analysis was performed for air-conditioning applications. Non-dimensional irreversibilities of different components of a vapor compression refrigeration cycle were determined to evaluate their exergy loss contribution.

• Results show that COP increases initially on additions of CO2 concentration till it reaches its maximum

COP and then decreases till pure CO2. An increase up to 40% COP is observed from pure hydrocarbons.

The maximum COP is observed between 10% CO2 to 20% CO2 for air-conditioning applications. On

average for CO2 concentrations up to 50% the COP of mixtures is higher than pure hydrocarbons. It is

observed that smaller ΔTmin values have higher total COP, when compared to higher ΔTmin values.

• Among the HC mixtures, dimethyl ether + CO2 mixtures have higher COP for lower CO2 concentrations,

and propylene + CO2 mixtures have higher COP for higher CO2 concentrations.

• Volumetric refrigeration capacity is observed to increase with addition of CO2 concentration, till it

reaches its maximum at pure CO2 concentrations. A similar pattern is observed in chapter 3 of this thesis. • Among the non-dimensional irreversibilities expansion valve has the highest value, indicating a large

contribution towards exergy loss. It is observed that smaller ΔTmin values have lower total non-

dimensional irreversibilities when compared to higher ΔTmin values.

52

PART 2: LAMINAR BURNING SPEED (SAFETY STUDY)

5. Experimental Facility and Laminar Burning Speed Model

In chapters 3 and 4, this thesis discussed the determination of performance parameters using different methods and approaches. Now, in chapters 5 and 6, this thesis will discuss the experimental measurement of laminar burning speed of propane + CO2 + air mixtures. The first section of this chapter discusses the experimental facilities and the second section of the chapter discusses the multi-shell thermodynamic model that is used for attainment of the laminar burning speed values from the experimental- pressure rise input.

5.1 Experimental Facilities

Experiments were conducted using a cylindrical chamber to measure the laminar burning speed and to study the flame stability. The stainless-steel cylindrical chamber has an inner diameter of 13.5 cm, a length of 13.5 cm and can withstand pressures up to 75 atm. It is capped with fused-quartz windows of 7.62 cm thick on both ends and is sealed with high-temperature elastomer O-rings. The purpose of the glass windows is to capture the flame propagation, but they limit the maximum pressure for the vessel. The chamber was set up in a Z-shaped Schlieren system coupled with a high-speed CMOS camera, which was used to capture the evolutionary behavior of flames up to 4000 frames per second [80; 81; 82; 83]. K-type thermocouples were used to measure the inside gas temperature and the interior wall temperature of the combustion chamber. Two extended spark plugs which have a spark gap of 0.9 mm were used for central point ignition. For minimizing the spark discharge effect on flame propagation, the spark energy must be set close to the minimum ignition energy [84; 85]. The combustion vessel was filled by the partial pressures method via a gas manifold system consisting of high accuracy pressure sensors, strain gages, valves, pipes, gas tanks, and a Welch ChemStar vacuum pump. Figure 48 shows the schematic of the experimental facility, where the cylindrical chamber is at the center, and the dashed lines through the chamber indicate the light reflections of the Z-shaped Schlieren system. A Kistler Quartz pressure sensor which is suitable for high frequencies (up to 400 kHz) is used to measure the pressure inside the chamber as a function of time. Since the flame shape is naturally spherical, it was affected by the cylindrical shape of the chamber at the later stages of expansion. When the spherical flame hits the chamber wall, the unburned gas at the corner pockets of the chamber gets ignited and becomes burned gas. This combustion of the corner gas increases the pressure even after the flame hits the chamber wall [86]. Hence, only the pressure-time data of the flame before hitting the inner wall of the combustion 53

chamber was considered for calculating the laminar burning speed. So that the flame is laminar, smooth, spherical and stable, which was also confirmed by the flame pictures taken through the Schlieren system. Before filling the gases for an experiment, the chamber was initially vacuumed to less than 50 millitorr, and then the vessel was initially filled with propane and CO2 then air and the mixture was left free for 3 minutes to achieve quiescence and uniform temperature [87]. After 3 minutes, a spark was generated at the center of the vessel, and the mixture is ignited. Each experimental condition was repeated at least three times for a higher confidence level. More information regarding the experimental equipment can be found in previous publications [88; 89; 90; 91; 92; 93].

Figure 48 Schematic of the experimental facilities

5.2 Laminar Burning Speed Model

The experimental pressure rise data as a function of time during the flame propagation in the cylindrical vessel was the primary input of the multi-shell thermodynamic model, which was used to measure the laminar burning speed of smooth flames. This model was developed by Metghalchi, Keck, and co-workers [94; 95], and it has been modified to include several correction factors. Which include the effects of radiative and conductive energy losses to the chamber walls, effects of temperature gradients in the burned gas, effects of preheat zone, wall and 54

electrode boundary layers and as well as the effect of conductive energy loss to spark electrodes. Figure 49 shows the multi-shell model and several assumptions that were made. As shown in the figure, the gases are divided into two regions, burned gas region and unburned gas region separated by a reaction layer of negligible thickness. The burned gas zone is further divided into ‘n’ number of shells, where ‘n’ is directly proportional to the combustion duration. The temperatures of each shell differ from each other, yet all burned gas remains in local equilibrium at their respective shell temperatures.

The burned gases are surrounded by a preheat zone of a thickness 훿푝ℎ having variable temperatures due to the energy transfer from the reaction layer to the unburned gas zone. Core unburned gas with uniform temperature surrounds the preheat zone. The core unburned gases are surrounded by a thermal boundary layer of thickness 훿푤푏 which separates the unburned gas from the chamber wall. Similarly, to the wall boundary layer, the effect of energy transfer from the unburned and burned gas to spark electrodes is determined using a thermal electrode boundary layer of thickness 훿푒푏. Further assumptions include that the pressure throughout the chamber is uniform, both burned and unburned gases always behave as ideal gases and are compressed isentropically, the unburned gas is frozen in its initial composition, and the energy flux due to the temperature gradient in the burned gas is negligible. More details on the model is given in Refs. [96; 97; 98; 99; 100; 101; 102; 103; 104; 105; 106].

Nomenclature for the Figure 49:

Rb = Burned gas radius; Vb = Burned gas volume

Rc = Chamber radius; Vu = Unburned gas volume

Tb = Burned gas temperature near the reaction layer

Tu = Unburned gas temperature (uniform)

Tw = Wall temperature

δph = Preheat zone surrounding burned gas

δwb = Thermal boundary layer at the chamber wall

Figure 49 Schematic of laminar burning speed model

For spherical flames, distribution of temperature in the burned and unburned gas regions is given by the ideal gas equation of state 55

푝푣 = 푅푇 (11)

Where, 푝 is pressure, 푅 is specific gas constant, 푇 is temperature and 푣 is specific volume.

The mass conservation equation for the unburned and burned gas regions is given by

푝푖(푉푐 − 푉푒) 푚 = 푚푏 + 푚푢 = 휌푏푉푏 + 휌푢푉푢 = (12) 푅푇푖

Where, 푚 is total mass of the gas in the chamber, 푚푢 is the mass of the unburned gas, 푚푏 is the mass of the burned gas. Subscripts 푢, 푏, 𝑖 denote unburned gas, burned gas and initial conditions respectively.

푉푐 is the volume of the chamber, 푉푒 is the volume of the electrode, 휌 is the average density.

The total volume of the gas in the combustion chamber is

푉푖 = 푉푐 − 푉푒 = 푉푏 + 푉푢 (13)

Where,

푚푏 푚푏 ′ ′ ′ ′ ′ ′ 푉푏 = ∫ 푣 (푇 , 푝)푑푚 = ∫ 푣푏푠(푇 , 푝)푑푚 − 푉푒푏 (14) 0 0 is the volume of the burned gas and 푉푏푠 is the specific volume of the isentropically compressed burned gas and

( ) 푉푒푏 = ∫푒푏(푣푏푠,푒푏 − 푣푏,푒푏)푑푚 = 퐴푒푏훿푒푏 15 is the displacement volume of the electrode boundary layer, 퐴푒푏 is the area of the electrode and 훿푒푏 is the displacement thickness of the electrode boundary layer,

푚 ′ ′ ′ 푉푢 = ∫ 푣 (푇 , 푝)푑푚 = 푚(1 − 푥푏)푣푢푠 − 푉푤푏 − 푉푝ℎ (16) 푚푏

푚 is the volume of the unburned gas, 푥 = 푏 is the mass fraction of the burned gas, 푣 is the specific volume 푏 푚 푢푠 of the isentropically compressed unburned gas,

푉푤푏 = ∫ (푣푢푠,푤푏 − 푣푢,푤푏) 푑푚 = 퐴푤푏훿푤푏 (17) 푤푏 is the displacement volume of the wall boundary layer, 퐴푤푏 is the area of the area of the chamber wall, 훿푤푏 is the displacement thickness of the wall boundary layer,

푉푝ℎ = ∫ (푣푢푠,푝ℎ − 푣푢,푝ℎ)푑푚 = 퐴푝ℎ훿푝ℎ (18) 푝ℎ is the displacement volume due to the preheat zone, 퐴푝ℎ is the area of preheat zone, 훿푝ℎ is the displacement thickness of the preheat zone. 56

The displacement thickness equations were derived by Keck and Metghalchi [47].

0.5 2 훼푏푟 푇푏 훿푒푏 = ( ) ( − 1) (19) 3 푟̇ 푇푤

0.5 1 푧 1 푧 −0.5 훼 휏 − 푢 푟 ′ ′푟 ′′ ′′ 훿푤푏 = ( ) 푧 푢 ∫ (푧 − 푧 푢) (∫ 푧 푑푧 ) 푑푦 (20) 휋 0 푧′

훼푢 푇푏 푇푏 훿푝ℎ = − ( ) ( − 1) ln (21) 푟 푇푢 푇푢

Where, 훼푏 is the thermal diffusivity of burned gas, 푟 is the flame radius, 푇푏 is the temperature of the burned gas, 푇푤 is the wall temperature, 훼푢 is the thermal diffusivity of unburned gas, 휏 is the characteristic 푝 푡 burning time, 푧 = is the dimensionless pressure, 푦 = is the dimensionless time, 푇푏 is the temperature of 푝푖 휏 the burned gas.

The energy conservation equation is

퐸푖 − 푄푒 − 푄푤 − 푄푟 = 퐸푏 + 퐸푢 (22)

Where, 퐸푖 is the initial energy of the gas, 푄푒 is the energy loss to electrodes, 푄푤 is the energy loss to walls, 푄푟 is the radiation energy loss to the vessel wall,

푚푏 푚푏 ′ ′ ′ ′ ′ ′ 퐸푏 = ∫ 푒 (푇 , 푝)푑푚 = ∫ 푒푏푠(푇 , 푝)푑푚 − 퐸푒푏 (23) 0 0 is the energy of the burned gas, 푒푏푠 is the specific energy of the isentropically compressed burned gas,

푝푉푒푏 퐸푒푏 = ∫ 푐푣(푇푏푠,푒푏 − 푇푏,푒푏)푑푚 = (24) 푒푏 훾푏 − 1 is the energy defect due to the electrode boundary layer,

푚 ′ ′ ′ 퐸푢 = ∫ 푒 (푇 , 푝)푑푚 = 푚(1 − 푥푏)푒푢푠 − (퐸푤푏 + 퐸푝ℎ) (25) 푚푏 is the energy of the unburned gas, 푒푢푠 is the specific energy of the isentropically compressed unburned gas,

푝푉푤푏 퐸푤푏 = ∫ 푐푣(푇푢푠,푤푏 − 푇푢,푤푏)푑푚 = (26) 푤푏 훾푢 − 1 is the energy defect due to wall boundary and

푝푉푝ℎ 퐸푝ℎ = ∫ 푐푣(푇푢푠,푝ℎ − 푇푢,푝ℎ)푑푚 = (27) 푝ℎ 훾푢 − 1 is the energy defect due to preheat zone. 57

The equation (24), (26) and (27) were developed using the ideal gas assumption

푝푣 푒 − ℎ = (28) 푓 훾 − 1

푐푝 Where, ℎ푓 is the enthalpy of formation of the gas at zero degrees Kelvin and 훾 = is the specific heat ratio. 푐푣

푝 퐴푒푏훿푒푏푐푝푝 푄푒 = − 퐴푒푏 ∫ 훿푒푏푑푝 (29) 푅 0

푝 퐴푤푏훿푤푏푐푝푝 푄푤 = − 퐴푤푏 ∫ 훿푤푏푑푝 (30) 푅 0

The radiation energy loss is calculated by fitted statistical narrowband correlated-k (FSNB-CK) model [107; 108]

푁−푏푎푛푑 푁 푀

푄푟 = 푉 ∑ ∆푣 ∑ 휔푛푘푛(∑ 퐼푗,푛,푚 − 4휋퐼푏,푗) (31) 푗=1 푛=1 푚=1

Where V is the volume of the energy source, N−band is the number of narrow bands, ν is the wave number, N is the number of Gaussian quadrature points, ωn is the weight function, kn is the k-distribution of absorption coefficient, M is the number of directions, I is the radiation intensity. The effects of spectral radiation absorption were considered in this model. The energy loss by radiation decreases the laminar burning speed.

Finally combining equations (3), (4) and (6) gives,

푥푏 ∫ 푣푏푠(푇, 푝)푑푥 + (1 − 푥푏)푣푢푠 = 0 푉 − 푉 푉 + 푉 + 푉 푐 푒 + 푒푏 푤푏 푝ℎ (32) 푚 푚 moreover, combining equations (12), (13) and (15) gives,

푥푏 ∫ 푒푏푠(푇, 푝)푑푥 + (1 − 푥푏)푒푢푠 = 0 푝푉 훿 훿 푝ℎ − 퐴 ∫ 푤푏 푝푑훿 − 퐴 ∫ 푒푏 푝푑훿 − 푄 퐸 훾 − 1 푤푏 0 푒푏 0 푟 + 푢 (33) 푚 푚

The above equations have been solved for two unknowns: burned mass fraction, 푥푏(푡), and the burned gas temperature of the last layer, 푇푏(푟, 푡). Given pressure, 푝(푡), as a function of time, they can be solved numerically by using the method of shells.

58

Ultimately, the laminar burning speed is defined as

푚̇ 푏 푚푥̇푏 푆푙 = = (34) 휌푢퐴푏 휌푢퐴푏

Where, 2 2 퐴푏 = 4휋푟 − 2휋푟푒 (35)

is the area of the burned gas sphere, 푟 is the radius of the flame, 푟푒 is the radius of the electrode, 푥̇푏is the rate of mass fraction burned, which is calculated by numerical differentiation of 푥푏.

Error Analysis:

The combined standard uncertainty can be obtained as follows [109]

푛 2 휕푓 2 퐸(푆푙) = √∑ ( ) 퐸푖 (푥푖) (36) 휕푥푖 푖=1

Where, 퐸(푠푙) is the standard uncertainty of 푆푙 with 푥푖(푃, 푇, 휙, 퐷, 푀표푑푒푙) and sensitivity coefficient

휕푓/휕푥푖. In the present work, all measured values were assumed to be uncorrelated which means no covariance term exists in Eq. (36). Moreover, most of the sensitivity coefficients are considered to be unity.

The Omega PX419 pressure sensor has a precision of ±0.08%. The DP41-S pressure strain gage has an accuracy of ±0.005%. The uncertainty from 5010B dual mode amplifier is ±0.5%, and the error in the digital conversion is ±0.1%. The error generated in filtering the noise out of the signals after the digital conversion is less than 0.5%. The human error in filling process value is about ±0.005 of the digital readout which translates into 0.5% uncertainty of the pressure value. Therefore, the combined standard uncertainty for the burning speed measurements was estimated to be 1% using Equation (36).

59

6. Flame Instability and Results

In this chapter the flame instability and results from the thermodynamic model are discussed. The first section of this chapter discusses the flame instability using the images obtained from the Schlieren system. This section shows the snapshots of the flames at different initial pressures, equivalence ratios, and different CO2 concentrations. The second section of this chapter discusses the results and figures of laminar burning speed for different initial pressures and equivalence ratios, and different CO2 concentrations. This section also compares the experimental data with a chemical kinetic model.

6.1 Flame Structure and Stability

The Z-shaped Schlieren system allows capturing the images of flame propagation to study the flame structure and stability at different conditions. Flame propagation pictures were taken at a range of equivalence ratios of 0.7 to 1.2, for different initial pressures of 0.5 atm, 1 atm, and 2 atm and at varying concentrations of CO2. Hydrodynamic and diffusive-thermal effects are the two significant reasons for the cell formation and instability that occur in premixed flames. Hydrodynamic instability results from the gas expansion, which is due to the energy released by chemical reactions. This expansion causes a flow that induces any flame perturbation further away from its original structure [110; 111]. It is also associated with the thermal

푑푇 expansion ratio (휎 = 휌푢/휌푏) and the flame thickness, which is defined by 훿푓 = (푇푎푑 − 푇푢)/ ( ) , where 푑푥 푚푎푥

푇푎푑 is the adiabatic flame temperature and 푇푢 is the unburned gas temperature, 휌푢 is the density of the unburned gas and 휌푏 is the density of the burned gas. The diffusive-thermal instability results from the disparity of mass diffusion and thermal diffusion [112; 113]. Lewis number, a non-dimensional parameter which is defined as the ratio of thermal diffusivity of the reactant mixture to the mass diffusivity of the limiting reagent, distinguishes the thermal diffusive instability [114; 115].

Figure 50 shows snapshots of the flame for different CO2 concentrations at different pressures

(푝/푝푖), where 푝 is the pressure at that state and 푝푖 is the initial pressure for the respective experiment; for this figure pi=2 atm. It can be seen in this figure that for a given CO2 concentration the flame instability increases with increasing pressure and the flame becomes more cellular. This is due to the decrease in Lewis number and flame thickness and the increase of the expansion ratio. It also can be observed that with the increase in

CO2 diluent concentration the flame instability gets delayed. 60

Figure 50 Images of propane/air/CO2 mixture flames at different pressures and different CO2 concentrations at pi of 2 atm, Ti of 298 K and an equivalence ratio of 1.2

Figure 51 shows snapshots of the flame for different equivalence ratios for a CO2 concentration of 60%. From this figure, we can observe that, as the mixture becomes richer, the flame instability increases. This instability is due to the decrease in Lewis number, flame thickness and the increase of expansion ratio.

During the first stage of the flame propagation, the cracks generate a slight deformation on the flame, while the instability associated with cell formation appears suddenly and covers the whole surface of the flame at a certain radius and pressure. The critical Peclet number (푃푒푐푟 = 푟푐푟/훿푓) is defined as the ratio of critical flame radius (rcr), to the flame thickness (δf), and gives an estimation parameter for the flame instability. Here, the critical flame radius is defined as the radius at which the cell formation occurs spontaneously over the flame surface, which is an onset of flame instability.

Table 7 shows the critical Peclet number for different CO2 concentrations and equivalence ratios for an initial pressure of 2 atm. With the increase in diluent concentration, the critical Peclet number increases for a given state which delays the flame instability. Increasing the pressure or increasing the equivalence ratio from stoichiometric to rich mixtures, the critical Peclet number decreases which promote early flame instability. In the table smooth for the equivalence ratio of 0.8 denotes that the flame was always smooth, and no cells were observed. 61

Figure 51 Images of propane/air/CO2 mixture flames at different pressures and different equivalence ratios at pi of 2 atm, Ti of 298 K and the CO2 concentration of 0.6

Table 7 Critical Peclet number of propane + air mixtures at initial pressure of 2 atm

풑풊 = ퟐ 풂풕풎 φ = 0.8 φ = 1 φ = 1.2

푪푶ퟐ = ퟎ% Smooth 519 257

푪푶ퟐ = ퟏퟓ% Smooth 530 276

푪푶ퟐ = ퟑퟎ% Smooth 541 284

푪푶ퟐ = ퟒퟎ% Smooth 555 292

푪푶ퟐ = ퟓퟎ% Smooth 559 306

푪푶ퟐ = ퟔퟎ% Smooth 616 343

62

6.2 Results and Discussion

Laminar burning speed was measured only for smooth flames having flame radii greater than 4 cm to minimize the stretch effect and the spark discharge effect [116; 117]. Spark discharge creates a shock wave followed by an extremely high-temperature region which affects the initial flame kernel propagation. This effect drops as soon as the temperature of the hot region drops, which happens at a higher flame radius.

Figure 52, Figure 53, and Figure 54 show the change in laminar burning speed with unburned gas temperature for different CO2 concentrations at different initial pressures of pi = 1 atm, pi = 2 atm and pi = 0.5 atm respectively. All the three figures correspond to stoichiometric mixtures (휙 = 1), at an initial chamber temperature (Ti) of 298 K. It can be observed from these figures that the laminar burning speed increases with the unburned gas temperature for all CO2 concentrations. It can also be seen that with increasing CO2 concentration as a diluent to the propane/air mixture the laminar burning speed decreases.

From Figure 52 we can see that the laminar burning speed at 0% CO2 is between 40 and 45 cm/s and blending with CO2 decreases the laminar burning speed to 15 cm/s before the flammability limit is reached at 82.5% CO2. The requirement of the laminar burning speed values changes according to the application, i.e., for an internal combustion engine, higher laminar burning speed is preferred for better performance. For a refrigerant application lower laminar burning speed is preferred for the safety during an accidental leak.

Figure 52 Laminar burning speed for propane/air/CO2 mixtures along isentropes at different CO2 concentrations, Ti = 298 K, φ = 1, pi = 1 atm 63

In Figure 53, we can observe that for a higher initial pressure of 2 atm, the laminar burning speed at

0% CO2 is around 32-34 cm/s and it decreases to 23 cm/s at 60% CO2. Laminar burning speed values at high initial pressure might be of interest for engine applications as the internal combustion takes place at high pressures. Similarly, from Figure 54, we can see that at a lower initial pressure of 0.5 atm, the laminar burning speed has high values when compared to the laminar burning speed for similar conditions at higher initial pressures.

Figure 53 Laminar burning speed for propane/air/CO2 mixtures along isentropes at different CO2 concentrations, Ti = 298 K, φ = 1, pi = 2 atm

Figure 54 Laminar burning speed for propane/air/CO2 mixtures along isentropes at different CO2 concentrations, Ti = 298 K, φ = 1, pi = 0.5 atm 64

Figure 55 shows the laminar burning speeds along isentropes for three different initial pressures at a CO2 concentration of 60%, the initial temperature of 298 K and the equivalence ratio of 1.2. From this figure, it can be interpreted that the laminar burning speed decreases with an increase in initial pressure. It should also be noted that experimental data for rich mixtures and high pressures is not easily achievable as the flames at these conditions tend to be more cellular (not stable) than the flames at lean mixtures and low pressures and this effect was shown in Section 3 [118].

Figure 56 and Figure 57 show the laminar burning speed along isentropes at different equivalence ratios. For figure 9 the experiments were conducted at an initial temperature of 298 K, the CO2 concentration of 30% and an initial pressure of 1 atm. Results show that the highest laminar burning speed was obtained at an equivalence ratio of φ = 1.1 and it can also be seen that for both φ = 1 and φ = 1.2 the laminar burning speed data overlaps for all unburned gas temperatures. In the same figure, we also observe that laminar burning speed for φ = 0.7 has lower values when compared to other equivalence ratios.

Figure 57 shows results collected at an initial temperature of 298 K, the CO2 concentration of 15% and an initial pressure of 0.5 atm. At these conditions the highest laminar burning speed was obtained at an equivalence ratio of φ = 1, followed by that at φ = 1.2 and, finally, that at φ = 0.8, for all unburned gas temperatures.

Figure 55 Laminar burning speed for propane/air/CO2 mixtures along isentropes at different pi of 0.5 atm, 1 atm and 2 atm, Ti = 298 K, φ = 1.2, CO2 = 60% 65

Figure 56 Laminar burning speed for propane/air/CO2 mixtures along isentropes at different equivalence ratios of 0.7, 1, 1.1 and 1.2, Ti = 298 K, pi = 1 atm, CO2 =30%

Figure 57 Laminar burning speed for propane/air/CO2 mixtures along isentropes at different equivalence ratios of 0.8, 1 and 1.2, Ti = 298 K, pi = 0.5 atm, CO2 = 15%

The laminar burning speeds for propane/air/CO2 mixtures was also determined numerically using the detailed H2/CO/C1-C4 kinetics model (111 species, 784 reactions) of Wang et al. [119] via a steady one- dimensional laminar premixed flame code from the CANTERA software program (simulation).

Figure 58, Figure 59, and Figure 60 show the comparison between the laminar burning speed measured based on experimental data and the laminar burning speed predicted by the kinetic model. Results shown in Figure 58 is for different CO2 concentrations (0% - 80%) at an unburned gas temperature of 300 K, 66

a pressure of 1 atm and at an equivalence ratio, φ = 1. From this figure, we can see that the experimental results are in excellent agreement with the kinetic model at the stoichiometric condition. In Figure 59 laminar burning speeds were experimentally-determined and numerically-simulated at different equivalence ratios, for the unburned gas temperature of 300 K, a pressure of 1 atm and at CO2 concentrations of 30% and 60%. In the figure, the laminar burning speed from the experiments and the simulations vary a little at non- stoichiometric conditions, but they are still in good agreement.

Figure 58 Experimental and modeling results of laminar burning speed for propane/air/CO2 mixtures at different CO2 concentrations, Tu = 300 K, p = 1 atm, φ = 1

Figure 59 Experimental and modeling results of laminar burning speed for propane/air/CO2 mixtures through at different equivalence ratios, Tu = 300 K, p = 1 atm, CO2 = 30%, 60%

Figure 60 displays experimentally-determined and numerically-simulated laminar burning speeds at different unburned gas temperatures, at a pressure of 1 atm, at the equivalence ratio of φ = 1 and a CO2 67

concentration of 60%. From figures 11-13, it can be seen that the experimental data and the simulation data are in overall good agreement.

Figure 60 Experimental and modeling results of laminar burning speed for propane/air/CO2 mixtures through at different unburned gas temperatures, φ = 1, p = 1 atm, CO2 = 60%

Chapters 5 and 6 conclusions:

Experimental studies have been conducted to investigate the laminar burning speed and the flame structure/stability of propane/air/CO2 premixed flames. The effects of pressure, equivalence ratio and CO2 (diluent) concentration on the flame stability were studied using Schlieren photography during the experiment. It was observed that with an increase in pressure or with an increase in equivalence ratio the flame tends to be more unstable and prone to cellularity. It was also noted that with an increase in CO2 concentration the flame acquires delayed instability. The effect of critical Peclet number with flame stability was also tabulated.

It was observed that the laminar burning speed increases with unburned gas temperature and the increase in pressure has an inverse effect on the laminar burning speed. It was also seen that with the rise in

CO2 concentration, the laminar burning speed decreases. With a change in CO2 concentration from 0% to

80% by 10% increment, decreases the laminar burning. That is a change in CO2 concentration from 0% -

10% reduces the laminar burning speed by 2%, and a change in CO2 concentration from 70% to 80% decreases the laminar burning speed by 33%. Flame ceases to exist at CO2 concentrations above 82.5%.

Similarly, for a refrigeration system, a mixture with a high percentage of CO2 is suggested to reduce the flammability risk of the accidental leaks. The laminar burning speed was found to be maximum at an equivalence ratio of 1.1 among the tested conditions. Experimental laminar burning speed results are in good agreement with simulation results by using Wang’s Mechanism for stoichiometric conditions, but the simulation results deviate a little from the experimental results for non-stoichiometric conditions. 68

CONCLUSION

The mixtures of hydrocarbons and carbon-dioxide were evaluated as alternate refrigerants by conducting a performance study (thermodynamic analysis) and safety study of selected mixture (laminar burning speed measurement).

Initially the refrigerants from REFPROP were screened according to their thermophysical properties to identify viable hydrocarbons for the mixtures with carbon dioxide. Further evaluation was performed using a vapor compression refrigeration cycle with constant ETF temperature system to determine the best four refrigerants. The selected refrigerant mixtures were then later used for thermodynamic analysis of a VCRC with variable ETF temperature system.

For, constant ETF temperature system, the COP is observed to decrease with addition of CO2 concentration in the mixture of HC + CO2. This is due to the additional exergy loss caused in the energy (heat) exchangers due to the non-isothermal phase change of zeotropic mixtures. Thermodynamic analysis for variable energy transfer fluid temperature method showed that the COP of mixtures is higher for than pure hydrocarbons for lower concentrations of carbon dioxide. This is due to the lower exergy loss observed in the energy exchangers in VCRC.

Laminar burning speed measurement for propane + CO2 + air was performed for different initial pressures, equivalence ratios, and CO2 concentrations. It is observed that laminar burning speed decreases with increase in CO2 concentration. For a 10% to 20% CO2 concentration in the HC + CO2 mixture, where

COP is maximum the laminar burning speed has only 10% decrease from pure propane. For a CO2 concentration of 50%, where the COP is same as pure hydrocarbon the laminar burning speed is decreased by 40%. This comparison depicts that a average CO2 concentration of 40% to 60% in the HC + CO2 mixture is a good mixture which has comparable performance with pure hydrocarbons and significant reduction in laminar burning speed. 69

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APPENDIX

A.1 Screened out refrigerants in step 1 of chapter 2

In this tables, HAP indicates that the fluid has been named as a hazardous air pollutant by the environment protection agency of the United States of America. Other factors include global warming potential (GWP) greater than 100, ozone depletion potential (ODP) greater than zero, toxicity, extremely flammable, and difficult phase change.

Table 8 List of refrigerants screened out of screening step 1 in chapter 2 of this thesis

S. Chemical name Reason S. Chemical name Reason No No 1 D4 ODP>0 44 R245fa GWP = 1030 2 D5 ODP>0 45 R32 GWP = 675 3 D6 ODP>0 46 R365mfc GWP = 794 4 Deuterium Low mass 47 R40 GWP = 13; ODP = 0.02 5 Fluorine GWP>100 48 RC318 GWP = 9,540 6 Hydrogen Flammable 49 RE143a GWP = 756 7 MD2M ODP>0 50 RE 245cb2 GWP = 654 8 MD3M ODP>0 51 RE 245fa2 GWP = 812 9 MD4M ODP>0 52 RE 347mcc GWP = 530 10 MDM ODP>0 53 Sulfur GWP = 22,200 hexafluoride 11 MM ODP>0 54 Trifluoro GWP>100 iodomethane 12 Nitrogen trifluoride GWP = 16,000 55 Oxygen Difficult phase change 13 GWP = 298 56 SO2 Difficult phase change 14 Orthohydrogen Extreme 57 Carbonyl sulfide Flammable flammable

15 Parahydrogen Extreme 58 Hydrogen sulfide Flammable flammable 16 GWP = 8860 59 Chlorine ODP>0 17 Perfluoropentane GWP = 9160 60 Chlorobenzene Mildly toxic and possible effect of chlorine on ODP 18 R11 GWP = 4750 61 1,2 Mildly toxic and possible Dichloroethane effect of chlorine on ODP 19 R113 GWP = 6130 62 Perfluoro hexane GWP = 3200 20 R114 GWP = 10,000 63 Mild toxic and possible effect of chlorine on ODP 21 R-115 GWP = 7370 64 R1224yd(Z) ODP = 0.00012 22 R-116 GWP = 12,200 65 Diethanolamine HAP 23 R-12 GWP = 10,900 66 Hexane HAP 81

24 R-1216 GWP = 17,340+ 67 R1234zd(E)-trans ODP>0 25 R123 GWP = 77; ODP 68 Trifluoroiodomet Toxic = 0.02 hane 26 R1233zd(E) GWP = 1; ODP = 69 Ethylbenzene HAP 0.00034 27 R124 GWP = 609 70 m-xylene HAP 28 R125 GWP = 3500 71 o-xylene HAP 29 R13 GWP = 13,396 72 p-xylene HAP 30 R134a GWP = 1,430 73 Isooctane HAP 31 R14 GWP = 7,390 74 Toluene HAP 32 R-141b GWP = 725 75 Methanol HAP 33 R-142b GWP = 2,310 76 Hydrogen Sulfide Toxic 34 R-143a GWP = 4,470 77 Propylene oxide HAP 35 R152a GWP = 124 78 Carbonyl Sulfide Toxic 36 R-21 GWP = 151 79 Ethylene glycol HAP 37 R218 GWP = 8830 80 Benzene HAP 38 R22 GWP = 1810 81 1,3 Butadiene HAP 39 R227ea GWP = 3,220 82 Ethylene oxide HAP 40 R23 GWP = 14,800 83 Carbon monoxide Toxic 41 R236ea GWP = 1370 84 Toxic 42 R236fa GWP = 9810 85 Hydrogen Toxic Chloride 43 R245ca GWP = 693