Instituto Tecnológico y de Estudios Superiores de Monterrey
Campus Monterrey
School of Engineering and Sciences
Experimental Study of a Coiled Flow Inverter as Heat Exchanger for its use with TiO
2–Water Nanofluids and its Potential Implementation as an Enhanced Heat Exchanger for Desorption in a Modified Commercial
Direct–Fired Ammonia–Water Absorption Chiller
A dissertation presented by
Bárbara Arévalo Torres
Submitted to the
School of Engineering and Sciences
in partial fulfillment of the requirements for the degree of Doctor of Philosophy
In
Engineering Sciences
Monterrey Nuevo León. August 2nd, 2021
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v Dedication
This thesis is respectfully dedicated to Dr. José Luis López Salinas and his beautiful family.
vi Acknowledgements
Thank you to Professors: Dr. Alejandro Javier García Cuéllar, Dr. José Luis López Salinas, Dr. Carlos Iván Rivera Solorio, Dr. Miguel Ángel Gijón Rivera, and Dr. Nicolás Velázquez Limón.
Thank you to the institutions: Tecnológico de Monterrey and CONACyT.
vii Experimental Study of a Coiled Flow Inverter as Heat Exchanger for its use with TiO2– Water Nanofluids and its Potential Implementation as an Enhanced Heat Exchanger for
Desorption in a Modified Commercial Direct–Fired Ammonia–Water Absorption Chiller by
Bárbara Arévalo Torres
Abstract
The chaotic mixing occurring in the chaotic flow and developed inside a coiled flow inverter (CFI) produces the removal or weakening of thermal and hydrodynamic boundary layers, causing heat and mass transfer augmentation. The present work examines the synergy of incorporating a thermal improved working fluid (i.e. nanofluid) as a means of enhancing the in-tube forced convection heat transfer in a CFI. Also, this work considers a CFI as a potential enhanced desorber in an absorption refrigeration system.
Aqueous solutions of dispersed TiO2 nanometer-sized particles (i.e., TiO2–Water nanofluids) were prepared and characterized by measuring their physical and transport properties. The effects on forced convective heat transfer were experimentally investigated in a test rig including a CFI prototype. The convective heat transfer coefficients were obtained at Reynolds numbers (𝑁𝑅𝑒) and TiO2 nanoparticle volume concentrations ranging from 1400–9500 and 0–1.5 v/v%, respectively. The Nusselt number (𝑁𝑁𝑢) in the CFI containing 1.0% v/v nanofluid was 41%–52% higher than in the CFI containing pure base fluid (i.e., water), while the 1.5 v/v% nanofluid increased the 𝑁𝑁𝑢 by 4%–8% compared to water. Two new correlations to predict the 𝑁𝑁𝑢 of TiO2–water nanofluids in the CFI at Reynolds numbers of 1400 ≤ 𝑁𝑅𝑒 ≤ 9500 and nanoparticle volume concentrations ranges of 0.2–1.0 v/v% and 0.2–1.5 v/v% are proposed.
A fully instrumented test bench was built from a gas–fired activated commercial ammonia–
water absorption chiller of 3–ton (i.e. 10 kW cooling capacity). The test bench was charged with ammonia–water solution and startup trials were executed. They showed instabilities in the solution flow. A coiled flow inverter desorber prototype with s curvature ratio () value of 10 was constructed based on the models established on the doctoral thesis of Vargas Bautista [1] which precedes this thesis.
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Contents
Dedication . . . v
Acknowledgements . . . vi
Abstract . . . vii
List of Figures. . . x
List of Tables . . . xiii
Nomenclature . . . xiv
CHAPTER 1. INTRODUCTION 17 1.1 Background . . . 17
1.2 Justification . . . 19
1.3 Objectives . . . 20
1.4 Scope . . . 21
1.5 Thesis Outline . . . 21
CHAPTER 2. LITERATURE REVIEW 22 2.1 Chaotic Advection for Thermal Mixing . . . 22
2.1.1 Coiled Flow Inverter for Heat Exchange: A Special Case of Chaotic Advection . . . 22
2.1.2 Nanofluids in Chaotic Flows . . . 25
2.2 Low Temperature Heat Source in Absorption Refrigeration Systems . . . 27
CHAPTER 3. EXPERIMENTAL APPROACH AND DATA PROCESSING 3.1 Forced Convective Heat Transfer in a Coiled Flow Inverter Heat Exchanger Prototype using TiO2–Water Nanofluids . . . 34
3.1.1 Nanofluids Preparation . . . 34
3.1.2 Nanofluids Characterization . . . 34
3.1.3 Nanoparticles Characterization . . . 35
3.1.4 Test Rig and Instrumentation . . . 36
3.1.4.1 Calibration . . . 39
3.1.5 Trials . . . 40
3.1.6 Data Processing . . . 40
3.1.6.1 Uncertainty Analysis . . . 42
3.2 Test Bench from a Commercial Ammonia–Water Absorption Chiller and Manufacturing of the Desorption Unit as a Coiled Flow Inverter Heat Exchanger . . . 43
3.2.1 Test Bench Description . . . 43
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A) Ammonia–Water (Refrigerant–Absorbent) Loop . . . 45
B) Chilled Water Loop . . . 49
3.2.1.1 Instrumentation . . . 50
3.2.1.2 Leaking Test Trials . . . 57
3.2.1.3 Preparation of Ammonia–Water Solution and its Feeding into the Test Bench . . . 58
3.2.1.4 Thermal Oil Heater . . . 60
3.2.2 Coiled Flow Inverter Heat Exchanger as Desorption Unit 63 CHAPTER 4. RESULTS AND DICUSSION 70 4.1 Heat Transfer Performance of TiO2_Water Nanofluids Flowing in a Coiled Flow Inverter Prototype . . . 70
4.1.1 Nanoparticle Characterization by DLS, XRD, and SEM Analysis . . . 70
4.1.2 Nanofluid characterization
. . .
724.1.2.1 The Thermal Conductivity of the Nanofluids . . . 72
4.1.2.2 The Dynamic Viscosity of the Nanofluids . . . 74
4.1.2.3 The Density of the Nanofluids . . . 76
4.1.2.4 The Specific Heat Capacity of the Nanofluids . . . 78
4.1.3 Forced Convective Heat Transfer Study . . . 80
4.2 Operation of the Test Bench Assembled from a Commercial Ammonia–Water Absorption Chiller and Operating with its Original Desorber . . . 90
4.2.1 Performed Trials . . . 90
CHAPTER 5. CONCLUSIONS, CONTRIBUTIONS, AND FUTURE WORK. . . 98
APPENDICES . . . . 101
References . . . 115
Curriculum Vitae . . . 122
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List of Figures
Figure 2.1. Basic components of a single effect absorption refrigeration system ...28
Figure 2.2. Generator absorber heat exchange (GAX) cycle. Figure taken from Herold, Radermacher, and Klein [67] ...31
Figure 3.1. (a) Schematic of the experimental test rig. (b) Photograph of the experimental setup.………..………….38
Figure 3.2. Schematic of the test section and the thermocouple locations. Arrows indicate the flow direction of the fluid. ...39
Figure 3.3 Commercial ammonia–water absorption chiller ACC60 Robur®’s ...44
Figure 3.4 ACC60 Robur®’s ammonia/water absorption chiller. Figure taken from robur.com. ...44
Figure 3.5 Components of the commercial ammonia–water absorption chiller ACC60 Robur®’s. * [=] The chilled water pump is part of the cooling loop, which is described on B) sub–section. ...46
Figure 3.6. Upper view (left) and side view (right) of the brazed plate heat exchangers used to replace the absorber and condenser heat exchangers in of the commercial ammonia–water absorption chiller ACC60 Robur®’s. ...47
Figure 3.7. Flanged connections of brazed plate heat exchangers used to replace the absorber and condenser heat exchangers in of the commercial ammonia–water absorption chiller ACC60 Robur®’s. ...47
Figure 3.8. Stainless steel rotary vane pumps used to replace the original diaphragm pump used to transfer solution from absorber to generator. ...48
Figure 3.9. Connection of the test bench components for the ammonia–water loop. ...48
Figure 3.10 Water–ethylene glycol circuit to operate ammonia–water loop. ...49
Figure 3.11. 5–ton R410a vapor compression system (York, model YCJD60S41S1). ...50
Figure 3.12. Turbine flowmeter placed in the chilled water loop. ...52
Figure 3.13. Pressure transducer placed in the ammonia–water loop. ...53
Figure 3.14. Coriolis flowmeter placed in the ammonia–water loop. ...54
Figure 3.15. Operation of coriolis flowmeter with water flowing through the test bench. ...55
Figure 3.16. Except for the coriolis flowmeters, the instruments in the test bench were connected to the data logger and operated simultaneously. ...55
Figure 3.17. Diagram of the test bench including instruments, accessories and process units. ...56
Figure 3.18. Pipe connection welded to one section of the desorber. ...57
Figure 3.19.Charging the ammonia–water loop with distilled water and subsequently with nitrogen. ...58
Figure 3.20. Sight glass placed between the condenser and evaporator in the ammonia–water loop (see Figure 3.17). The flow of distilled water is observed. ...58 Figure 3.21. 20 kg tank of ammonia used to prepare the 33 w/w% ammonia–water solution.59
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Figure 3.22. Solution preparation tank (see APPENDIX D for detailed procedure) ...60 Figure 3.23. (a) Thermal oil heating system model LV-100-G from Industrias Thermoflux S.A.
de C.V. (b) Installation of the system outdoors. ...61 Figure 3.24. The thermal oil heating system installed. ...62 Figure 3.25. Heated oil streams sent to the indoors fitted with J-type thermocouples and a turbine-type flow meter with viscosity compensation due to temperature variation. ...62 Figure 3.26. Desorber from the ACC60 Robur®’s chiller. ...63 Figure 3.27. CAD model of the tube side in the coiled flow inverter heat exchanger for its potential operation as desorber in the test bench. ...64 Figure 3.28. Leaking test on the coiled flow inverter heat exchanger tube side. ...65 Figure 3.29. CAD assembly of the shell side including the circular cut plate baffles placed between all CFI sets. ...66 Figure 3.30. Physical assembly including the CFI sets and the circular cut plate baffles. ...67 Figure 3.31. Top view of the physical assembly including the CFI sets, the circular cut plate baffles and the shell. ...67 Figure 3.32. Exterior view showing the shell side of the coiled flow inverter heat exchanger for its potential operation as desorber in the test bench. ...68 Figure 3.33. Incorporation of two separation stages at the outlet of the tube side to ensure the supply of pure refrigerant to the condenser. ...69 Figure 4.1. Particle size distribution of the TiO2 nanoparticles (NPs), obtained from the dynamic light scattering (DLS) analysis………..71 Figure 4.2. X-ray diffraction pattern of rutile TiO2. ...71 Figure 4.3. Scanning electronic microscopy (SEM) image of the TiO2 NPs. ...72 Figure 4.4. (a) Experimental thermal conductivity of TiO2–water (W) nanofluids (NFs) as a function of temperature and NP volume concentration. (b) Comparison of the measured thermal conductivity at 35 °C and that predicted from two of the classical models on effective thermal conductivity. ...74 Figure 4.5. (a) Experimental dynamic viscosity of TiO2–W NFs as a function of temperature and NP volume concentration. (b) Comparison of the measured dynamic viscosity at 35 °C and that predicted by the classical models of dynamic viscosity...76 Figure 4.6. (a) Experimental density of TiO2–W NFs as a function of temperature and NP volume concentration. (b) Comparison of the measured density at 35 °C and that predicted by the mixture rule principle [94]. ...78 Figure 4.7. (a) Experimental specific heat of TiO2–W NFs as a function of temperature and NP volume concentration. (b) Comparison of the measured specific heat at 35 °C and that predicted by the Xuan and Roetzel [97] model. ...80 Figure 4.8. Experimental 𝑁𝑁𝑢 versus 𝑁𝑅𝑒 in water and TiO2–W NFs. Temperature values along the CFI set (see Figure 3.2 for reference) are included in APPENDIX C. ...82
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Figure 4.9. Deviation between the experimental and predicted 𝑁𝑁𝑢 values of TiO2–W flows. The predicted values were obtained by the proposed empirical correlations: Equation (14) in (a) and Equation (15) in (b). ...84 Figure 4.10. Plots of 𝑁𝑁𝑢 versus 𝑁𝑅𝑒. The symbols are the experimental data and the curves were calculated by Equation (14) in (a–a.iv) and Equation (135 in (b–b.iv). Cumulative results (a,b), and the results of NFs with a NP concentration of 1.5 v/v% (a.i,b.i), 1.0 v/v% (a.ii,b.ii), 0.5 v/v% (a.iii,b.iii), and 0.2 v/v% (a.iv,b.iv). Legends as shown in (a,b). ...86 Figure 4.11. Plots of𝑁𝑁𝑢/𝑁𝑃𝑟0.4 versus 𝑁𝑅𝑒. The symbols are the experimental data and the curves were calculated by Equation (14) in (a–a.iv) and Equation (15) in (b–b.iv). Cumulative results (a,b), and the results of NFs with an NP concentration of 1.5 v/v% (a.i,b.i), 1.0 v/v%
(a.ii,b.ii), 0.5 v/v% (a.iii,b.iii), and 0.2 v/v% (a.iv,b.iv). Legends as shown in (a,b). ...87 Figure 4.12. 𝑁𝑁𝑢 versus 𝑁𝑅𝑒 of 1.0 v/v% TiO2–W NF in the coiled flow inverter (CFI) set, obtained in the present experiment (symbols) and predicted by the present correlations (Equation (14) shown using a dashed curve and Equation (15) shown using a short dashed curve), 1.0 v/v% TiO2–W NF tested by Duangthongsuk and Wongwises [100] in a straight tube (Equation (16) shown using a solid curve), and water in a helical coiled tube evaluated by Shchukin [102] (Equation (17) shown using a dotted curve). ...89 Figure 4.13. Pressure measurements recorded from pressure transducers during the execution of the first trial of the test bench. ...91 Figure 4.14. Temperature measurements from the outlet of desorber recorded during the execution of the first trial of the test bench. ...92 Figure 4.15. Low solution level in the desorber. ...93 Figure 4.16. Correct solution level in the desorber. ...93 Figure 4.17. Burner from a 5–ton absorption chiller used as replacement for the original test bench burner. ...94 Figure 4.18. Replacement of the original burner under operation. ...94 Figure 4.19. Temperature measurements from the outlet of desorber recorded during the trial executed after installing the burner coming from a 5–ton absorption chiller (more robust than the original) and feeding the third batch of solution. ...96 Figure 4.20. Mass flow rate (kg/s) readings displayed by the coriolis flowmeter located at high pressure side in the test bench. ...96 Figure 4.21. Leaking solution from the vane pumps installed in the test bench. ...97 Figure A.1. Confirmation run after thermocouple calibration (the symbols indicate the
thermocouple position, see Figure 3.1and Figure 3.2 to identify the position of each
thermocouple). ...101 Figure C.1 Temperature values along the CFI set for water at 𝑁𝑅𝑒 8000………..…103 Figure C.2. Temperature values along the CFI set for TiO2–water nanofluid at 0.2 v/v% and at
𝑁𝑅𝑒 8500…...………103
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Figure C.3. Temperature values along the CFI set for TiO2–water nanofluid at 1.0 v/v% and at
𝑁𝑅𝑒 8500...………104
Figure C.4. Temperature values along the CFI set for TiO2–water nanofluid at 1.5 v/v% and at
𝑁𝑅𝑒 8500………104
Figure D.1 Brazed plate heat exchangers used as absorber. ...105 Figure D.2. Brazed plate heat exchangers used as condenser. ...106
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List of Tables
Table 2.1. Typical thermal conductivities for conventional heat transfer fluids and
nanoparticles ...25 Table 2.2. Examples of commercial ammonia–water single–effect absorption chillers of
cooling capacity up to 50 kW ...29 Table 3.1. Experimental conditions in the present study. ...40 Table 3.2. Specifications for brazed plate heat exchangers used to replace the absorber and condenser heat exchangers in of the commercial ammonia–water absorption chiller ACC60 Robur®’s. ...46 Table 4.1. Constant values of Equation (15) ...83 Table A.1. Thermocouple calibration Constants ...101 Table B.1. Uncertainty table for the calculation of the inner Nusselt number in the coiled flow inverter heat exchanger ...102
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Nomenclature
𝐴𝑜 external surface area of heat exchange (m2)
𝐶𝑝 specific heat capacity (of working fluid) (J kg−1 °C−1) 𝑑𝑖 inner diameter of the tube (m)
𝑑𝑜 outer diameter of the tube (m) 𝑑𝑐 coil diameter (m)
ℎ convective heat transfer coefficient (W m−2 °C−1) 𝑘 thermal conductivity (of working fluid) (W m−1 °C−1)
𝑘𝑇𝑆 thermal conductivity of the tube in the test section (W m−1 °C−1) 𝐿 length of the tube in the test section (m)
𝑚̇ mass flow rate of working fluid (kg s−1) 𝑛 empirical shape factor of nanoparticles 𝑁𝐷𝑒 Dean number (= 𝑁𝑅𝑒√𝑑𝑑𝑖
𝑐) 𝑁𝑁𝑢 Nusselt number
𝑁𝑃𝑒 Peclet number (= 𝑁𝑅𝑒𝑁𝑃𝑟) 𝑁𝑃𝑟 Prandtl number
𝑁𝑅𝑒 Reynolds number (=𝜌𝑣𝑑𝜇 𝑖) ppm parts per million
𝑄, 𝑄̇ heat transfer rate (W) 𝑉̇ volumetric flow rate (LPM)
𝑇0 bulk fluid temperature at the inlet (°C)
𝑇1–𝑇12 bulk fluid temperatures at the intermediate points along the test section (°C) 𝑇13 bulk fluid temperature at the outlet (°C)
𝑇𝑏 temperature of the heating bath water (°C)
∆𝑇𝑖𝑛 temperature difference at the inlet (°C)
∆𝑇𝑜𝑢𝑡 temperature difference at the outlet (°C)
∆𝑇𝐿𝑀𝑇𝐷 logarithmic mean temperature difference (°C) 𝑈 overall heat transfer coefficient (W m−2 °C−1) 𝑈𝑏 uncertainty in a calculated quantity
𝑈𝑋𝑖 uncertainty in a measured quantity 𝑣 inlet velocity (m s−1)
v/v % volume percent concentration (%) w/w % weight percent concentration (%)
𝑊 mechanical power (W)
𝑋𝑖 a measured quantity for Equation (16)
Subscripts
𝑎𝑏𝑠 absorber
𝑎𝑐𝑡 corrected (actual)
𝐵𝐹 base fluid
𝑐𝑜𝑛𝑑 condenser 𝑑𝑒𝑠 desorber
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𝑒𝑣𝑎𝑝 evaporator
𝑔𝑎𝑔𝑒 gage (measured from reading)
𝑖 inner
𝑁𝐹 nanofluid 𝑁𝑃 nanoparticle
𝑜 outer
𝑝 pump
𝑟𝑒𝑐𝑡 rectifier Greek symbols
𝛼0, 𝛽0, 𝛾 constants in Equation (15)
curvature ratio (=𝑑𝑑𝑐
𝑖)
𝜇 dynamic viscosity (kg m−1 s−1) 𝜌 density (kg m−3)
volume fraction of nanoparticles
𝐿 nanoparticles volume fraction enhancement limit for the present study ( 𝐿 = 0.015)
𝜓 sphericity of nanoparticle Acronyms
CFI coiled flow inverter DLS dynamic light scattering EG ethylene glycol
GNP graphene nanoplate
HVAC-R heating, ventilation, air conditioning, and refrigeration ICSD inorganic crystal structure database
LPM liters per minute
MWCNT multi-walled carbon nanotube NF(s) nanofluid(s)
NP(s) nanoparticle(s)
PALS phase analysis light scattering PHE plate and frame heat exchanger PTFE polytetrafluoroethylene
PVC polyvinyl chloride
SEM scanning electronic microscopy STHE shell and tube heat exchanger XRD X-ray diffraction
W water
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This chapter is divided into five sub–sections. In the first one, background information related to enhanced heat exchangers is provided. Also, in that sub–section, an explanation of why an absorption refrigeration cycle can be coupled to enhanced heat exchangers is presented. The subsequent three sections includes the justification, the objectives and the scope of the present thesis. This chapter concludes with the outline of this work.
1.1 Background
Heat exchangers play a vital role in a wide range of industries, such as the pharmaceutical, food and beverage, chemical, petrochemical, oil and gas, power generation, HVAC-R (heating, ventilation, air conditioning, and refrigeration), and other fields. The processes involved may require either streams at specific operation temperatures, the removal/addition of heats of reaction, mixing, adsorption, etc., or the unit operations for drying, boiling, and condensing, among other steps. Industry-specific requirements have led to the development of enhanced heat exchangers capable of transporting high heat fluxes without compromising on practical sizing aspects. In literature, the improvement of the thermal performance of heat exchangers is referred to as heat transfer enhancement, augmentation or intensification. In general, heat transfer augmentation entails increasing the heat transfer coefficient. The consequence of this may be viewed either as an increase of the heat flux or as a reduction of the size of the equipment needed for a specified heat load, or a combination of both.
Enhancement techniques may be categorized as passive or active, depending on whether they need external power. Passive techniques do not need external power. In order to outperform the individual technique, a variety of permutations and combinations of active and passive techniques have been used and they are identified as compound enhancement techniques. Examples of active enhancement methods are well stirring the fluid or vibrating the surface. Examples of passive enhancing methods are: (a) treated surfaces, (b) rough surfaces, (c) extended surfaces, (d) displaced enhancement devices, (e) swirl flow devices, (f) coiled tubes, (g) surface tension devices, (h) additives for fluids, and many others. As mentioned, compound enhancement techniques involve the combination of two or more of the techniques exemplified above in order to produce an enhancement synergy compared to any of the techniques operating separately. The effectiveness of any of them is strongly dependent on the mode of heat transfer, which may range from single-phase free convection to dispersed-flow film boiling. In heat exchangers, the primary modes of heat transfer occurring inside them can be accurately “generalized” in conduction and external/internal forced convection (radiation typically removes only a small portion of the overall thermal load). Thus, in enhanced heat exchangers, the design strategies followed to reduce the
Chapter 1 Introduction
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thermal resistance imposed by boundary layers in external/internal forced convection generally fall into either passive or active mechanisms, or even a combination of both, being the former the most commonly used due to their lower cost, their easiness to implement and longer operating life.
The utilization of coiled tubes as passive enhancement technique is one of the most popular geometrical features for compact heat exchangers. They are appropriate because by coiling a tube it is produced a secondary flow consisting of a pair of longitudinal vortices which are consequence of the centrifugal forces within the fluid. Hence, heat transport occurs not only by diffusion on the radial direction but also by convection. The contribution of that secondary convective transport dominates the overall process and enhances the heat transfer per unit length of tube compared to a straight tube of equal length. In addition, the secondary flow decreases entrance lengths and reduces the difference between laminar and turbulent heat transfer rates.
In the same context of achieving further thermal performance in coiled tubes, the concept of “chaotic advection” first introduced by Aref [2] was also integrated into these geometries and the coiled flow inverter came up as a novel an enhanced heat exchanger (in sub–section 2.1.2. this equipment is described in detail). The fundamental approach behind it has to do with getting multiple flow inversions achieved by changing the direction of centrifugal force in helically coiled tubes. These inversions in consequence cause that the radial vortices (generated by the centrifugal forces present on the secondary flow) rotate around the axis of the tube. This fact results in further mitigation of the radial velocity gradients.
On the other hand, the improvement of the thermophysical properties of working fluids, as a passive enhancement technique, offsets the low thermal conductivity of conventional heat transfer fluids (e.g., water, ethylene glycol, and engine oil). For instance, the thermal properties that affect convection heat transfer include thermal conductivity, specifc heat capacity, density and viscosity. As a result, every heat transfer model must include precise details on thermal properties. Extensive research have been conducted to improve thermal conductivity of these fluids by suspending nanoparticles of diverse materials in heat transfer fluids, known as nanofluids (see sub–section 2.1.3). In fact, modern technology room provides opportunities to process and produce particles further below 50 nm. It may also expected and desirable that nanofluids provide not only higher heat transfer rate, but also good stability of the suspension by eliminating possible agglomeration and sedimentation to permit long-term application.
Considering that chaotic advection in coiled tube geometry and nanofluids are both enhancement techniques that improve heat transfer efficiency passively, as a first part of this experimental work, the benefits of incorporating this compound enhancement technique to increase the thermal performance in a exchanger are explored and confirmed, taking into consideration that an experimental study of forced convective heat transfer in a coiled flow inverter using nanofluids as the working fluid is yet to be seen in literature.
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As a second part of this experimental work, the redesign of the desorption stage in a commercial absorption refrigeration system is performed. In the absorption refrigeration system, a physicochemical process replaces the mechanical process of the conventional vapor compression refrigeration system by using energy in the form of heat rather than mechanical work. Thus, absorption refrigeration system has been selected since it is integrated by four heat exchangers and the utilization on enhanced heat exchangers (as coiled flow inverter heat exchangers are) in a system like this becomes thermally attractive.
Particularly, the coiled flow inverter heat exchangers have been thought as an immediate solution for intensifying heat exchangers, however, its application on cooling systems has not been explored.
One substantial part of this second part in this thesis has to do with the construction and habilitation for the testing bench of the cycle already mentioned. It may be worth to point out that this testing bench is basically an existing absorption commercial chiller which its components are intended to be pulled apart and then place needed instruments for measurement purposes and manipulation flowing in streams.
1.2 Justification
Current research on experimental studies using coiled flow inverters as an enhanced device for thermal process intensification is limited. Particularly, studies of forced convective heat transfer in coiled flow inverters using nanofluids (NFs) as the working fluid is scarce. Furthermore, extensive heat transfer research has focused on Al2O3–Water (W) and CuO–W systems [3] and, in comparison, there are limited studies on TiO2–W systems.
TiO2 nanoparticles (NPs) possess excellent physical and chemical properties, such as chemical stability, good dispersivity, and non-toxicity. In fact, TiO2 NPs exhibit better dispersion than other metal oxide NPs [4]. Hence, the first part of this thesis (subsections 3.1 and 4.1) includes experimental studies of forced convective heat transfer in a coiled flow inverter using TiO2–W NFs at volume concentrations of 0.2, 0.5, 1.0, and 1.5 v/v%
within Reynolds numbers of 1400 ≤ 𝑁𝑅𝑒 ≤ 9500. Also, physical and transport properties of the NFs needed to be measured as the classical two-component theoretical models to predict the physical and transport properties of NFs do not take into account the contribution of the dispersant [5].Heat transfer correlations to predict the Nusselt number of the studied NFs were then proposed.
On the other hand, the investigations in regards with the utilization of enhanced heat exchangers in NH3-Water vapor absorption refrigeration systems is relatively limited compared to those for LiBr-Water. Particularly, there are no studies on the mass and heat transfer effects for an ammonia–water desorption in a coiled flow inverter until now (see sub–section 2.2 for detailed information. The only related research work on using a coiled flow inverter to characterize the transport of a fluid in two phases is limiting to observe the flow pattern of an heterogeneous mixture water–air and quantify the volume of air occupied in water without there being heat transfer involved [6]. The work of Kumar et al. [7] is in
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fact the one that utilized a coiled flow inverter as a heat exchanger but its thermal performance was evaluated only in single–phase.
When thinking on coiled flow inverter as a novel alternative for the heat transfer intensification in a double–phase process and considering that such heat exchanger has not been previously used in that application, it would be convenient to take into consideration the work of Vashisth and Nigam [6] already referred, in where it is indicated an optimum curvature ratio =10. Higher or lower values lead to increased liquid hold- up and reducing the gas void fraction. The prototype constructed by Kumar et al. [7] had a value of 20.
The second part of this thesis (subsections 3.2 and 4.2) aims to extensively modify a commercial direct-fired ammonia-water absorption chiller to allow fluid pressures, temperatures and flowrates to be measured at strategic points in the system. This fully instrumented test bench is intended to be useful for future work with an experimental coiled flow inverter desorber prototype (with =10), which was constructed as part of the present work and whose construction was based on the models established on the doctoral thesis of Vargas Bautista [1] which precedes this thesis.
1.3 Objectives
The objectives of this thesis are contextualized on the utilization of a coiled flow inverter as an enhanced heat exchanger following two lines of experimental research: the one that covers the use of nanofluids as working fluids and the one that considers it as a potential low-temperature heat exchanger in an absorption refrigeration system.
1.3.1 General
I. To perform an experimental study on a coiled flow inverter using TiO2_Water nanofluids as inner working fluids.
II. To consider a coiled flow inverter as a potential desorber in a modified commercial ammonia-water absorption chiller.
1.3.2 Specific
i. To build up a lab scale forced convective test rig including a coiled flow inverter heat exchanger prototype.
ii. To prepare TiO2_Water nanofluids at volume concentrations of 0.2, 0.5, 1.0, and 1.5 v/v% and measure them the physical and transport properties that have an effect on forced convection heat transfer.
iii. To execute steady–state trials in the forced convective test rig at different flow rate conditions using the TiO2_Water nanofluids prepared.
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iv. From the experimental data of the TiO2_Water nanofluids tested in the forced convective test rig at different flow rate conditions, to execute applicable data processing to obtain heat transfer empirical correlations to predict the Nusselt number.
v. To build-up a testing bench of an ammonia-water absorption refrigeration by:
v.i. Spliting-up the operational units of a commercial ammonia–water absorption chiller model ACC60 (3–ton) from Robur®.
v.ii. Incorporating instrumentation at strategic points in the system to measure fluid pressures, temperatures and flowrates.
vi. To perform base line trials in the testing bench described in specific objective v.
vii. To generate a CAD model and engineering drawings of the coiled flow inverter experimental prototype for desorption purposes defined at Chapter 7 of Vargas- Bautista [1] doctoral thesis.
viii. To follow up the manufacturing stages involved in the fabrication of the prototype described on specific objective vii.
1.4 Scope
The scope of this work extents to:
a) Obtain correlations to predict Nusselt number in a coiled flow inverter using as inner working fluids TiO2_Water nanofluids at volume concentrations of 0.2, 0.5, 1.0, and 1.5 v/v% in the Reynolds number range 1400 ≤ 𝑁𝑅𝑒 ≤ 9500.
b) Construct a fully instrumented ammonia-water absorption system from a direct–fired 3–ton commercial chiller and manufacture a coiled flow inverter desorber with curvature ratio =10.
1.5 Thesis Outline
The subsequent chapters in this thesis have been organized in the following manner:
Chapter 2 presents the state of art in the fields of chaotic advection used for enhanced thermal mixing purposes and, in the other hand, this section also includes the approaches taken for the utilization of low-grade thermal energy in absorption refrigeration systems (these being scenarios where enhanced heat transfer techniques could be integrated). In Chapter 3 methods for the experimental procedures and the processing of data obtained from the present study are presented and explained. In Chapter 4, the results obtained from the data analysis are reported with interpretation. In Chapter 5, a summary of results, conclusions and future work from the present study are presented.
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This chapter reviews previous research in the area of enhanced heat exchangers by incorporating two passive techniques for heat transfer improvement: 1) Chaotic advection in a particular case of coiled heat exchangers and 2) nanofluids. On the other hand, this section also describes those studies related to the utilization of low temperature heat sources in absorption refrigeration systems. This as a prelude to consider incorporating an enhanced heat exchanger into a system like this.
This chapter aims to summarize the development in the aforementioned fields, underscore its engineering significance, and the underlying significance of our contribution.
2.1 Chaotic Advection for Thermal Mixing
Advection implies the transport of a particle with the mean fluid flow. Aref [2] introduced the concept of chaotic advection by modelling the motion of a particle in an ideal stirred tank. The flow was composed of two blinking vortices and by blinking each vortex one at a time there were obtained chaotic fluid particle trajectories showing showed that chaotic flow is created by the discontinuity of the flow. This phenomenon is called chaotic advection. In short, chaotic advection means chaotic motion of passive particles in a non- randomness velocity field. Aref [2] was not only the first to coin the term chaotic advection, but he also established the clear connection between the flow domain and the phase space of dynamical systems, thus, creating a clear and powerful link between fluid mechanics and dynamical systems. The result of this work had two effects: 1) a great increase in the study of flows using dynamical systems, and therefore the Lagrangian approach and 2) the key involvement of chaotic advection within the field of fluid mechanics and mixing. As it might be intuitive, chaotic advection can enhance heat transfer by reducing temperature gradients becoming on temperature profiles more uniform. Two main categories of flow geometries are used in chaotic mixing: 1) Active mixers: those that use rotating or translating elements as cylinders [8]–[11] and 2) passive mixers: those that use multiple periodic changes in the geometry [12]–[20].The former require external power to bring about the mixing effect while passive ones do not utilize energy input.
2.1.1 Coiled Flow Inverter for Heat Exchange: A Special Case of Chaotic Advection Helically coiled tubes, classified as a passive technique to promote thermal mixing, are a simple and effective way of augmenting heat transfer: by coiling a tube it is produced a secondary flow consisting of a pair of longitudinal vortices which are consequence of the centrifugal forces within the fluid. The secondary flow along coiled
Chapter 2 – Literature Review
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geometries promotes the transport of colder material from the center of the tube to the hotter regions close to the wall (or vice versa). The streamlines of both Dean vortices are closed curves, and the fluid particles moving among them do not mix [21]. In addition, the fluid parcels near the center of the vortices do not approach the tube walls and heterogeneous temperature fields appear in the radial direction [22]. Chaotic mixing produced by chaotic advection is then conceived as a practical way to perturb this secondary flow and produce chaotic trajectories for the particles to achieve greater mixing within the fluid [21].
The coiled flow inverter (CFI) introduced by Saxena and Nigam [23] is a particular class of curved geometry that promotes chaotic mixing through a combination of coils and bends. The centrifugal forces acting on the secondary flow change direction at each equally spaced 90° bend along the length of a straight helical coiled tube. The bent coils in this pioneering work comprised different numbers of bends, bend angles, and spacing between the bends to achieve the narrowest residence time distribution.
It was found that CFI geometry exhibited up to a 20 times reduction in the dispersion number compared to that in a helical coil geometry.
Typically, bent coils are made up by four helical “arms” arranged in a square shape.
These four-arm CFI units, which are used in different research fields, including heat transfer enhancement, are henceforth referred to as CFI sets.
In the laminar regime, Kumar and Nigam [24] numerically characterized the hydrodynamics and forced convection in a CFI set, and compared them to those occurring in a straight coil configuration. The fully developed Nusselt number (𝑁𝑁𝑢) was 20–30% higher in the bent coil configuration (CFI set) than in the straight coil case. In a later work, Kumar and Nigam [25] analyzed a comparable scenario but with uniform heat flux at the wall. Convective heat transfer was evaluated in the same CFI geometry and fluid parameters as their earlier study. The 𝑁𝑁𝑢 values were 25–36%
higher in the bendy coil than in the coil without bends. The study also included the effect of temperature-dependent viscosity on the hydrodynamics and heat transfer performance. When evaluated in a CFI set, the 𝑁𝑁𝑢 was higher in 14% than in the constant viscosity case. From the numerically obtained𝑁𝑁𝑢, one might foresee the temperature dependency of the fluid properties. Mridha and Nigam [26] studied the fluid flow and heat transfer characteristics in a CFI set under turbulent flow conditions.
They numerically solved the three-dimensional differential governing equations of mass, momentum, and energy. The velocity and temperature profiles revealed that increasing the number of bends in the CFI set increased the uniformity of both secondary fields by radial mixing. In the simulation data, the heat transfer enhancement was 4–13% relative to the coiled tube without flow inversions. In addition, the 𝑁𝑁𝑢 increased with Dean number (𝑁𝐷𝑒) in all three of their evaluated scenarios: straight tube, straight helical coil, and the CFI set.
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Most of the experimental research in heat transfer characteristics for CFI sets has been performed using water as the working fluid. For instance, Kumar et al. [7]
evaluated the flow dynamics and thermal performance in a heat exchanger similar to a shell and tube (STHE) with eight CFI sets at the tube side. The shell side included circular cut plate baffles placed between all CFI sets, and a central rod for support.
Experiments were conducted under both laminar and turbulent flow conditions in the counter-flow arrangement, with flowing hot water through the CFI side (tube side) and either tap water or ambient air through the shell side. The overall heat transfer coefficient (𝑈) increased with increasing internal flow. The fully developed 𝑁𝑁𝑢 values in the CFI geometry were compared with those in a straight coiled tube predicted by existing correlations. The CFI geometry boosted the ℎ𝑖 by 25% and 12% in the ranges 1,000 < 𝑁𝑅𝑒< 10,000 and 10,000 < 𝑁𝑅𝑒< 16,000, respectively. Fitting the simulated curves to the experimental data, Kumar et al. [18] obtained empirical correlations between the friction factor (𝑓) and 𝑁𝑁𝑢.
Mandal et al. [27] compared the performances of the CFI heat exchanger used by Kumar et al. [7] with those of the conventional plate heat exchanger (PHE) and the STHE. These three devices were rated under equivalent heat transfer areas and process conditions. In the CFI evaluation, the test bench of Kumar et al. [7] was modified to ensure the turbulent flow condition at the tube side. The PHE and STHE were not tested under this flow condition due to technical constraints in the experimental setup. The number of transfer units calculated in the laminar flow range was approximately 3.7–7.5 and 2.0–2.2 times higher in the CFI heat exchanger than in the STHE and PHE, respectively. Moreover, based on the hydrodynamic and heat transfer results, the turbulent flow inside the CFI sets increased the 𝑓 by 7–10% and augmented the 𝑁𝑁𝑢 by 12–14%, relative to the flows in the STHE and PHE.
Correlations were again obtained by fitting the simulations to experimental data.
Singh and Nigam [28] evaluated the pressure drops and heat transfer performances of three CFI heat exchangers with different heat transfer areas. The geometric features of the prototypes were similar to those of the heat exchanger evaluated by Kumar et al. [7], differing only in the number of CFI sets inside the shell. In Singh and Nigam [20], the tube side of the prototypes contained one, two, or four CFI sets. Experiments were conducted predominantly under turbulent flow conditions by varying both the cold- and hot-stream mass flow rates. The hot water or compressed hot air flowed inside the flow inverter, while the cooling water or ambient air flowed through the shell side. In the setup with one CFI set (𝐴 = 0.22 m2), the 𝑈 was up to 3.6 and 4.5 times higher, respectively, than in the PHE and STHE (𝐴 = 1.76 m2) obtained by Mandal et al. [27]. Moreover, the measured pressure drop per unit area in the prototypes with one and two CFI sets were comparable to those obtained in the PHE.
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2.1.2 Nanofluids in Chaotic Flows
In addition to geometric changes, improving the thermal properties of the working fluids is another technique for enhancing heat transfer. This improvement can offsets the low thermal conductivity (𝑘) of conventional heat transfer fluids (e.g., water, ethylene glycol, and engine oil). Among these working fluids, water has the highest 𝑘 value, although it amounts to 0.6 𝑊/𝑚𝐾 under room temperatures, it is many orders lower than most metals or metal oxides, as shown in Table 2.1.
Table 2.1. Typical thermal conductivities for conventional heat transfer fluids and nanoparticles
Numerous theoretical and experimental studies of 𝑘 for suspensions containing solid particles were conducted since Maxwell presented a theoretical basis for predicting the 𝑘 of suspensions more than 100 years ago [37]. However, the typical approach followed on those studies was confined to millimeter- or micrometer-sized particles.
This conventional approach has two major technical problems: (1) conventional millimeter or micrometer-sized particles settle rapidly in fluids, (2) the values of 𝑘 for these suspensions are low at low particle concentrations. Taking this this into consideration, Choi & Eastman [38] conceived the novel concept of nanofluids by hypothesizing that the much larger relative surface area of nanophase powders, when compared with conventional micrometer-sized powders, might markedly improve the heat transfer capabilities and stability of the suspensions. It was reported in that work [38] the possibility of doubling convection heat transfer coefficients by using nanofluids, a result that would otherwise require a tenfold increase in pumping power.
The high surface area of nanoparticles enhances the heat conduction of nanofluids since heat transfer occurs on the surface of the particle. In addition to thermal
Substance Thermal conductivity
( 𝑊/𝑚𝐾 ) Reference
Conventional heat transfer fluid
Water ~0.598-0.609 [29]
Ethylene Glycol ~0.251 [30]
Engine Oil ~0.145 [31]
Mineral Oil ~0.115 [29]
Metallic Nanoparticle
Aluminum 237 [32]
Copper 398 [32]
Gold 315 [32]
Nonmetallic Nanoparticle
Alumina (Al2O3) 31–41 [33]
Copper Oxide (CuO) 77 [34]
Titania (TiO2) 8–11 [35]
Zinc Oxide (ZnO) 13–29 [36]
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properties of NFs, the mechanical, optical, electrical, and magnetic properties are superior to those of conventional bulk materials with coarse grain structures. NFs are then considered as the new generation of heat transfer fluids. They are engineered by homogeneously suspending nanometer-sized materials or structures in conventional base fluids such as water, ethylene glycol or oils. NFs have been found to possess enhanced thermal conductivity and convective heat transfer coefficients to those of base fluids. In addition to those advantages, the following are worth mentioning: better stability, higher specific surface area and lower particle clogging than colloidal suspensions.
Advances in nanotechnology have allowed new suspensions to be formulated by suspending various types of nanoparticles in conventional working fluids [5], [39]–[41].
Many researchers have experimentally determined the 𝑘 of different nanofluids with volume concentrations in the 0.5–4.0 volume percent concentration (v/v%) range.
Typically, k is 15–40% higher in the nanofluids than in pure base fluid [42]. Various nanofluids have been numerically, analytically, and experimentally studied in diverse heat-exchanger configurations, including tube-in-tube [43], [44], shell and tubes [45], [46], plates [47], [48], plate-fin [49], tube-fin [50], microchannels [51], [52], and helical pipe coils [53]–[56].
Despite the numerous studies on convective heat transfer using different nanofluids in conventional heat exchanger, only few investigations have integrated nanofluids and chaotic advection (as heat transfer enhancement techniques) and they are numerical:
Tohidi et al. [57] investigated the effects of nanofluid flow on heat transfer in a C- shaped channel to determine the possible benefits of using nanofluids in a chaotic flow. The chaotic geometry consisted of a C-shaped segment in the upper plane and a straight section in the lower plane. This pattern was periodically replicated in the longitudinal direction. The cross-section of these two rectangular channels was 1 x 1 mm. The results showed that the addition of CuO and Al2O3 nanofluids at 1-3 v/v% in the C-shaped channel could lead to an improvement in thermal efficiency of 4-14%
and 4-18%, respectively, when compared to the C-shaped channel with base fluid (e.g.
water). Singh, Choudhary, and Nigam [58] numerically solved the governing equations of mass, momentum, and energy in straight tubes, straight helical coils, and an CFI set. They investigated the effects of Al2O3–water(W) nanofluid at different volume concentrations (1, 3, and 4v/v%) under laminar flow conditions. The computational 𝑁𝑁𝑢 values were compared with the experimental values in water reported by Kumar et al. [7]. Relative to the base fluid (water), the 1, 3 and 4 v/v% Al2O3–W augmented the 𝑁𝑁𝑢 by 24%, 33%, and 42% respectively in the CFI set, and by 15%, 25%, and 35% respectively in the straight helical coil. The superior heat transfer was attributed to chaotic mixing induced by flow inversion in the CFI set. Tohidi et al. [59] numerically analyzed the hydrodynamics and heat transfer characteristics of Al2O3–W and CuO–
W nanofluids at three volume concentrations (1, 2, and 3 v/v%) in a particular chaotic
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geometry under the laminar regime. A chaotic configuration was created by assembling consecutive sections of two regular helical segments with different pitches.
Internal flow inversion was generated by the geometrical perturbation at the bend of each helical segment. The results of the chaotic configuration were compared to those in the straight helical coil. The convective heat transfer coefficient (ℎ) in the chaotic configuration was increased by 18%, 19%, and 21% for the Al2O3–W nanofluid added at 1, 2, and 3 v/v%, respectively, and by 18%, 21%, and 25% respectively for the CuO–W nanofluid at the same concentrations. Bahiraei and Hangi [60] and Bahiraei, Mazaheri, and Alighardashi [61] evaluated the hydrothermal characteristics of TiO2– W nanofluids at 1, 3 and 5 v/v% and 1, 3 and 4 v/v%, respectively, in a chaotic geometry as the one used by [57]. From both investigations, it was found that convective heat transfer and pressure drop in the chaotic channel increased by increasing concentration and reducing particle size. Also, it was reported that temperature distribution in the C-shaped channel was more uniform than that in the ordinary straight channel as a result of intense mixing in the chaotic geometry.
2.2 Low Temperature Heat Source in Absorption Refrigeration Systems
With a heat source, absorption refrigeration systems (ARSs) are able to produce a cooling effect. So, a primary benefit of ARSs is their potential possibility to use low temperature heat sources. A single–effect ARS is constituted by four heat exchangers, a pump and two flow restrictors, as shown in Figure 2.1. It is a fact that heat exchangers are intrinsically the crucial part on ARSs. This cycle is completed by the following steps [62]: A flow of pure refrigerant enters the evaporator in the form of a cooled and low-pressure liquid. In there, heat is transferred from the targeted medium to be cooled down to the liquid refrigerant, causing it to evaporate. Then, absorber plays the role of suctioning the vapor of refrigerant in an absorbent liquid which is contained into the heat exchanger. During this step, refrigerant vapor is converted to liquid by exothermic condensation and then, subsequent absorption into the liquid solution occurs releasing thermal energy. The amount of this energy (per unit mass) is comparable to the sum of the enthalpy of vaporization of the refrigerant and the enthalpy of solution. The refrigerant rich solution is pumped to the generator (also named desorber), which together with condenser constitutes the high- pressure side of the system. In the generator, vapor refrigerant is desorbed from the absorbent by the addition of heat that is supplied by direct or indirect firing. The former implies hooking the system to a fuel line to power up the cycle within its own self-contained fuel unit in generator. The latter considers steam, hot water or a low-grade heat source.
As the refrigerant is boiled away in the desorber, the remaining mixture becomes poor in refrigerant and passes through a restrictor in order to flow back to the absorber to capture again refrigerant vapor that comes from the evaporator. High-pressure vapor refrigerant migrating from the desorber flows to the condenser where it condenses to then enter the
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restrictor which reduces its pressure. Then, the refrigerant enters to the evaporator and the cooling effect occurs.
Figure 2.1. Basic components of a single effect absorption refrigeration system Water–lithium bromide and ammonia–water are the most widely used absorption working solutions, especially because their respective refrigerants (water and ammonia, respectively) have a high value of latent heat of vaporization: H2O=2,257kJ/kg and NH3=1,357 kJ/kg [63], so the need of refrigerant flow rate in the cycle can be minimized.
In the LiBr machine, evaporator and absorber requirements are attained to the thermodynamic properties of the refrigerant (water). As a first conception, for a temperature of 5°C in the evaporator, the corresponding vapor pressure for water is approximately 0.009 atm. At this pressure, the specific volume of saturated steam is 129.2 m3/kg. This large vapor volume leads to several challenges for absorber and evaporator (components falling within that low pressure) including the need of significantly large vessels. In addition, at the desorption step it is required a large contact area to separate water vapor from aqueous lithium bromide solution, which makes the generator to be bulky. To a large extent this is consequence of two related facts: The temperature range required for desorption process: 75°C to 120°C [64] and the utilization of indirect firing - which in fact allows the integration of low grade heat sources. These features are the main reasons that each heat exchanger of traditional lithium bromide ARSs use the typical structure of a horizontally arranged tubes bundle operating on falling film mode [65]. In contrast, smaller absorption chillers are direct–fired and utilize ammonia-water as working solution. Although this working pair is free from crystallization issues and is chemically stable for a wide range of temperature and pressure conditions [66], it requires high
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activation temperatures in the range of 125°C to 170°C [64]. The reason behind this has to do with the dramatic variation of temperature that ammonia-water mixture has in phase change. Namely, this pair is a zeotropic mixture [66]. In consequence, desorption of ammonia from the aqueous solution behaves quite differently from the evaporation of the pure substance itself. Accordingly, the thermodynamic behavior of properties of ammonia- water mixture completely differs from the independent properties of water and ammonia.
This atypical behavior is especially found at vapor quality values higher than 0.75 and as expected, it also on the mass fraction of ammonia in the mixture [66]. In this scenario, the boiling point temperature difference between ammonia and water is rather small and in consequence, the vapor generated in the desorber contains a certain amount of water.
Rectification of the generated vapor is then needed by a rectification column and a dephlegmator[67]. Although includes this additional step, ammonia–water chillers are of particular interest and also attractive due to the high vapor pressure and the low specific volume of the refrigerant, ammonia. At some extent, both parameters provide some room to optimize the internal heat and mass transfers at absorber and generator and to reduce the fluid charge, thus helping to get a compact machine design. Following this trend, strong development efforts have been undertaken mainly in Europe to evolve some small scale ammonia–water single–effect absorption chillers with driving temperatures of 85°C and up. They have been available in the market since the end of 2006 [68] and some of these examples are given in Table 2.2.
Table 2.2. Examples of commercial ammonia–water single–effect absorption chillers of cooling capacity up to 50 kW
Company: Pink Solarice AGO
Product Name: Chilii PSC 19 AAC25 Congelo50
Cooling Capacity (kW): 19 25 50
Driving Temperature
(°C): 85 95 115
Cooled Water
Temperature (°C): 6 -3 -10
COP: 0.62 0.5 0.54
Dimensions (L x W x H): 1.9 x 0.8 x 0.6 2.0 x 1.4 x 1.4 2.35 x 2.63 x 1.35
Weight (kg): 550 950 1600
In parallel with commercial developments, prototypes of ammonia-water absorption chillers have been built for research purposes. The Gazi University (Turkey) built a prototype operated by solar energy obtaining the best COP (0.705) at a driving
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temperature of 80 °C for hot water and an evaporating temperature of 3 °C [69]. The University of Madrid (Spain) constructed a 2 kW prototype designed for solar-powered refrigeration in small rural applications. This prototype used a transfer tank instead of a solution pump. The former was connected alternatively between zones of high and low pressure, assuming that the opening and closing of valves would discharge the respective solutions. It was proved that this substitution was in fact inefficient and the system returned a COP lower than 0.05 [70] . The University of Applied Sciences (Germany) built a 10 kW cooling capacity water-ammonia absorption chiller including a novel membrane solution pump. At a driving temperature of 90 °C, cold water temperatures of 15 °C could be achieved with a cooling capacity of 7.2 kW and a COP of 0.66 [71]. The French National Institute for Solar Energy developed a 4.2 kW prototype based on brazed plate heat exchangers operating at 80°C, 27°C and 18°C temperatures for the generator, absorber/condenser and evaporator, respectively. This design yielded a COP of 0.65 [72].
The configuration of single effect system (depicted in Figure 2.1) has its own thermodynamic restrictions, such as the impossibility to exceed COP values greater than 1 and the imposition of a minimum temperature for the driving heat. In the effort to overcome these technical challenges, some noticeable proposals of cycle arrangements have been emerged and they basically consist on multi-staging the single-effect cycle.
These systems are better known as “advanced absorption cycles”. The approach behind advanced cycles mainly includes [73]: condensation heat recovery, absorption heat recovery, generator absorber heat exchange (GAX) and some others process arrangements. The GAX system is based on the utilization of the temperature overlap between the generator and the absorber: When pressures and concentrations in the absorber and desorber are maintained properly, there is a temperature overlap such that some of the heat of absorption can be injected to the desorber. In consequence, recuperating the energy in this fashion reduces the amount of heat input needed into the desorber. Thus, for a given thermal load in the evaporator, less energy is required to drive the cycle, and the COP increases as consequence. It can be considered that the COP for a GAX system could be 20-30% higher than that of the single-effect system operating at the same working conditions [74]. Figure 2.2 shows an example of the GAX cycle, where also two additional steps of heat recovery are included: That in between condensation and evaporation and also the utilization of the heat released in rectification stage.
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Figure 2.2. Generator absorber heat exchange (GAX) cycle. Figure taken from Herold, Radermacher, and Klein [67]
Ng et al. [75] performed experimentation with a 7 kW direct–gas fired GAX ammonia-water absorption air-cooled chiller. The nominal COP of the system was about 0.8 at an operating condition of a generator temperature of 200°C, condenser temperature of 44°C, absorber temperature of 41°C and evaporator temperature of 5°C. From the thermodynamic standpoint, a proper basis for refrigerant process-average temperature was derived. For instance, process-average temperature was realized as a non- measurable parameter which needed to be computed from the properly-weighted piecewise assimilation of measured temperatures along non-isothermal paths. Therefore the non-isothermal processes occurring in the absorption chiller were identified and quantified the inaccuracies that derived from an incorrect process of obtaining average temperature. The study revealed the significance of the refrigerant process-average temperature in the analysis of the chiller. The work of Ng et al. [75] is nowadays valuable due to the fact of the accuracy on predicting chiller performance by estimating optimal operating conditions which lead to chiller diagnostics and optimization. In a similar scenario, Priedeman et al. [76] validated numerical approximations against experimental results. They tested a direct-gas fired 17.5 kW ammonia-water GAX system under steady state, achieving temperatures of 8.3°C for chill water and 35°C at condenser. The COP of the system was found to be 0.68 at a full load condition. Simulation was also carried out, and the results of the experimentation were found to be in close proximity with the simulation results which showed an initial target COP value of 0.70. It was identified that lower burner generator efficiency, a pressure drop between the evaporator and pump inlet and less heat recovery in the GAX resulted in a lower performance for the system.
Velázquez and Best [77] reported the thermodynamic analysis for a 10 kW air-cooled GAX
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system operated with hybrid natural gas and solar energy. The COP was found to be 0.86 and also it was proved that as the temperature lift increases, the efficiency of the system decreases. Thus, that system was found to be an excellent option for air conditioning purposes where the temperature lift was small. That GAX ammonia-water prototype has been designed and constructed in the Centro de Investigación en Energía in Morelos, México. Following that research line, Gómez et al. [78] investigated the effect of utilize thermal oil as a heat source in the GAX prototype previously described. The system showed a theoretical COP of 0.58 at a generator temperature of 192 °C. It was found that the experimental values were lower than the design value, due to the lower operating temperature range of the heat source and the low performance of the absorber. An extensive review of simple GAX cycle and other advanced configurations also based on internal heat coupling are presented by Jawahar and Saravanan [74]. For example, the GAX cycle with an additional solution pump is known as “branched GAX cycle”. The reader is referred to that work for details on these latter adavanced cycles.
In small scale (<50 kW) ARS, the typical desorber is a falling film type with horizontal tubes where the heating working fluid flows inside the tube bank. Also, the tray type distiller coupled with a gas burner is widely utilized in 5–ton (17.58 kW). This type of desorber (i.e.gas-fired) requires the availability of a fuel source capable of providing the flame that allows the ammonia-water separation to be carried out.
PHEs are increasingly used as absorbers, desorbers, and solution heat exchangers in absorption chillers and heat pumps using NH3-based mixtures as they are compact and possess high thermal effectiveness. This latter advantage may allow to lower the driving temperature if compared with the tray type distiller desorber. PHEs can provide satisfactory heat and mass transfer performance in absorption refrigeration systems, but high solution pressure drops are also introduced accordingly because of the narrow flow channels and turbulent flow. However, since the operating pressures of NH3-based mixtures are relatively high, the pressure drop caused by PHE has a negligible effect on the system performance [79]. As mentioned in subsection 2.1.1, Mandal et al. [27] compared the hydrodynamics and thermal performance of conventional PHE and STHE with results reported for CFI heat exchanger by Kumar et al. [7]. It was found that the enhancement of heat transfer in CFI heat exchanger as compared to PHE and STHE is higher than the increase in pressure drop.
Previous research using coiled flow inverters as heat exchangers has mainly focused only on the heat transfer enhancement occurring at the tube side under a single-phase flow, while the mass transfer has not received attention yet. The closest work available in the literature on characterizing the transport of a fluid in two phases in a CFI is limited to observing and quantifying the volume of air occupied in water without there being heat transfer. Hence, Vashisth and Nigam [6] tested sixteen CFI(s) of different design configurations and evaluated the flow patterns encountered in two-phase gas liquid flow and void fraction. That two-phase was given by pumping water and driving air into a static
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mixer and then feeding the gas-liquid mixture into the CFI. They observed that as the liquid flow rate was increased, the gas blockage decreased for a given gas flow rate. Empirical correlations were reported for void fraction and two-phase friction factor. It should be noted that the working fluid fed to the exchanger was not a homogeneous mixture and, furthermore, its separation does not requires the injection of thermal energy.
A literature survey found that there are no studies on the mass and heat transfer effects for an ammonia–water desorption in a CFI. As already mentioned, the chaotic flow developed in a CFI produces the removal or weakening of thermal and hydrodynamic boundary layers. In this scenario, as the velocity increases as a consequence of the decrease in the hydrodynamic boundary layer, the mass transfer coefficient increases because of “fresh” fluid parcels coming into the mass transfer interface. If “fresh”
fluid elements come to the interface, then mass transfer driving force is higher and consequently molar flux is also higher. Analogous to this happens with the convective heat transfer coefficient when decreasing the thermal boundary layer (see previous section 2.1.1.). As a consequence of the heat/mass transfer intensification in a CFI and if it is used as desorber in a VARS, it would enable any low-temperature heat supply to drive the system. For example, if there is a hot water source at 95°C with a flow rate of 1000 gpm, this heat soruce may be utilized in an enhanced desorber that, if it were not, would require the heat source at 150°C and 1000 gpm to obtain the same refrigerant capacity.
