SYSTEM LEVEL THERMAL HYDRAULIC PERFORMANCE OF WATER-BASED AND PAO-BASED ALUMINA NANOFLUIDS. Thesis. Submitted to. The School of Engineering of the

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1 SYSTEM LEVEL THERMAL HYDRAULIC PERFORMANCE OF WATER-BASED AND PAO-BASED ALUMINA NANOFLUIDS Thesis Submitted to The School of Engineering of the UNIVERSITY OF DAYTON in Partial Fulfillment of the Requirements for The Degree Master of Science in Mechanical Engineering By Aaron Richard Veydt UNIVERSITY OF DAYTON Dayton, Ohio December, 2010

2 SYSTEM LEVEL THERMAL HYDRAULIC PERFORMANCE OF WATER-BASED AND PAO-BASED ALUMINA NANOFLUIDS Name: Veydt, Aaron Richard APPROVED BY: Robert J. Wilkens, Ph.D., P.E. Advisory Committee Chairman Associate Professor, Chemical and Materials Engineering Director, Chemical Engineering and Bioengineering Michael Elsass, Ph.D. Committee Member Lecturer, Chemical and Materials Engineering George R. Doyle, Jr, Ph.D. Committee Member Professor, Mechanical and Aerospace Engineering John G. Weber, Ph.D. Assistant Dean School of Engineering Tony E. Saliba, Ph.D. Dean School of Engineering ii

3 ABSTRACT SYSTEM LEVEL THERMAL HYDRAULIC PERFORMANCE OF WATER-BASED AND PAO-BASED ALUMINA NANOFLUIDS Name: Veydt, Aaron Richard University of Dayton Advisor: Dr. R.J. Wilkens Current and future military aircraft have a critical need for improved avionics heat removal. A drop-in replacement is sought for the existing heat transfer fluid, poly-alphaolefin (PAO). Nanofluids have been considered for this application due reports of increases in thermal conductivity higher than predicted by conventional theory. In this study, a coolant loop apparatus was designed and built to evaluate the laminar and turbulent heat transfer performance of water-based and PAO-based alumina nanofluids in a flowing system. Alumina/water solutions showed an increase in pressure drop with particle loading which caused the heat transfer coefficient (H) at equal pumping power to be lower than at equal flowrates. In turbulent heat transfer, the alumina/water nanofluids show a 1-5% increase in H at equal flowrates. At equal pumping power, the nanofluid H is lower than water. In laminar flow at equal flowrates the H is decreased, which is not predicted. iii

4 The alumina/pao nanofluid showed similar pressure drop performance to the pure PAO base fluid. In turbulent flow at equal flowrates and equal pumping power, the H increase is only 1-3%. In laminar flow, a similar increase of 1-3% was observed. This increase is too small to warrant further testing of these fluids. In addition, particle settling was observed after only a few hours, which leads to questions about the long term stability of these nanofluids in a continuously flowing system. Overall, the fluids tested showed only marginal enhancement to the heat transfer coefficient. There were no significant (i.e. order of magnitude) increases observed between the results and conventional theory as have been reported elsewhere. The results of this work show that coolant loop apparatus is a valuable system-level screening tool for the U.S. Air Force to evaluate new single-phase coolants for avionics cooling. iv

5 ACKNOWLEDGEMENTS I would first like to thank the U.S. Air Force for giving me the privilege of serving our great nation. To all of my professors and mentors at the U.S. Air Force Academy for teaching me invaluable lessons in leadership and officership and instilling in me the core values of the Air Force. Sometimes, the most lasting lessons are learned through our own failures. Next, I would like to thank the Thermal Sciences and Materials Branch for funding and supporting this research. Thank you to Tim Reid, Cheryl Castro, George Fultz, Art Safreit and Peter Deak for performing numerous property measurements, many times on short notice. Also, to Ed Snyder and Lois Gschwender for assigning this project to me and helping me to learn some of the fluid chemistry along the way. Thank you to Mike Green. You have a gift of knowing when to allow students to struggle and when to step in and help. I cannot possibly put a price on the practical knowledge that you have given me over the past two years. I would like to especially thank my research advisor, Dr. Robert Wilkens. I am deeply thankful for your wisdom and patience with me. Through your busy schedule, you were always able to find time for me even just to talk. Your investment in my learning and growth as a young engineer far outweighed any personal interest you may have had in this work. v

6 Thank you to my awesome family my parents and my sisters. You are blessing from the Lord and your love has helped me to become who I am today. To my beautiful wife, in two years, we ve been through so much. You ve stood by me through it all. You are my encouragement and my helper. You have been far more patient with me than I deserve. I want you to know that I am proud of you for taking on the burden of school and work in order for us to have more time together in the future. Finally, and most importantly, thank you to my Lord and Savior Jesus Christ. You have given me more than I deserve and I owe you my life. vi

7 TABLE OF CONTENTS ABSTRACT... iii ACKNOWLEDGEMENTS...v LIST OF ILLUSTRATIONS... ix LIST OF TABLES...xv LIST OF SYMBOLS... xviii SUBSCRIPTS...xx ABBREVIATIONS... xxi 1. INTRODUCTION Background Approach LITERATURE REVIEW AND MODELING BACKGROUND Literature Review Nanofluid Convective Heat Transfer Systems Nanofluid Thermal Fluidic Properties Modeling Background Convective Heat Transfer Coefficient Film Heat Transfer Coefficient for Fully Developed Pipe Flow Heat Transfer Coefficient for Thermally Developing (TD) Pipe Flow Overall Heat Transfer Coefficient and the LMTD Sensitivity Analysis of the H Pressure Drop and Pumping Power Laminar, Thermally Developing Flow Plate and Frame Heat Exchangers...24 vii

8 3. DESCRIPTION OF APPARATUS AND EXPERIMENTAL PROCEDURE Description of Apparatus Experimental Procedure Calibrations Viscosity Effects on Flowrate Measurement Uncertainty Information and Calculations TEST RESULTS AND DISCUSSION Water/Al 2 O 3 Results and Analysis PAO and PAO NF System Pressure Data Results and Analysis PAO and PAO NF Individual Pressure Data Results and Analysis CONCLUSIONS AND RECOMMENDATIONS Conclusions Recommendations...73 REFERENCES...75 APPENDIX A: PHOTOS OF FLOW LOOP COMPONENTS...82 APPENDIX B: OPERATING PROCEDURES...95 APPENDIX C: ADDITIONAL DATA PLOTS...99 APPENDIX D: FLOWMETER CALIBRATION PROCEDURE AND RESULTS APPENDIX E: HEATER CONTROL CALIBRATION APPENDIX F: THERMOCOUPLE CALIBRATION DATA AND RESULTS APPENDIX G: MEASURED FLUID PROPERTY INFORMATION APPENDIX H: RAW DATA APPENDIX I: METSS CORP. PERMISSION TO PUBLISH RESULTS viii

9 LIST OF ILLUSTRATIONS Figure 2.1 Schematic of microchannel heat exchanger (number of flow channels not to scale) Figure 3.1 Schematic of coolant loop apparatus Figure 3.2 Photo of completed coolant loop apparatus Figure 3.3 Figure 4.1 Amount of time required for the apparatus to reach steady operating conditions is approximately 2 hours from cold startup Water heat duty comparison. Heater input power is compared to microchannel heat exchanger, process, and cooling water calculated thermal energies Figure 4.2 Flowrate comparison of system pressure drop for water and Al 2 O Figure 4.3 Figure 4.4 Figure 4.5 Figure 4.6 Figure 4.7 Flowrate comparison of water and Al 2 O 3 Plate/Frame overall heat transfer coefficient Pump power comparison of water and Al 2 O 3 Plate/Frame overall heat transfer coefficient Flowrate comparison of water and Al 2 O 3 heat exchanger 1 film H Flowrate comparison of water and Al 2 O 3 heat exchanger 2 film H Flowrate comparison of system pumping power for PAO and PAO NF Figure 4.8 Flowrate comparison of the overall H for PAO and PAO NF Figure 4.9 Photo of particles settling in the PFA tubing after sitting for one night ix

10 Figure 4.10 Figure 4.11 Figure 4.12 Pump power comparison of PAO and PAO NF plate/frame overall H Flowrate comparison of PAO and PAO NF heat exchanger 1 Film H Flowrate comparison of PAO and PAO NF heat exchanger 2 Film H Figure 4.13 System power comparison of heat exchanger 1 Film H Figure 4.14 System power comparison of heat exchanger 2 Film H Figure 4.15 Component pressure drop for PAO Figure 4.16 Component pressure drop for PAO NF Figure 4.17 Plate/Frame pressure drop comparison of the overall H for PAO and PAO NF Figure 4.18 Flowrate comparison of HX1 film H for PAO and PAO NF Figure 4.19 Flowrate comparison of HX2 film H for PAO and PAO NF Figure 4.20 Figure 4.21 Microchannel pressure drop comparison of HX1 film H for PAO and PAO NF Microchannel pressure drop comparison of HX2 film H for PAO and PAO NF Figure A.1 Haight 6US GPM gerotor pump drawings Figure A.2 Haight gerotor pump and motor assembly Figure A.3 Figure A.4 Leeson speedmaster variable DC motor control for pump. Control includes a power switch, brake switch and a switch to reverse flow Lytron CP20 Microchannel Cold Plate Heat Exchanger was the smallest of its kind commercially available at the time of purchase. There are 42 flow channels H x 0.03 W x 2.0 L. The heat transfer surface area is 4in 2 per side x

11 Figure A.5 Figure A.6 Figure A.7 Cartridge heater control board from Proheat, Inc. Includes a variac (center) to set heater power setting, a voltmeter (top left) and ammeter (top right). The four switches control which heater blocks are operating at any time. A temperature alarm (bottom center) displays the hottest block temperature and shuts power off if temperature goes over the alarm setting Dimensions used for machining cartridge heaters hole in a 2 x2 x2 copper block Finite element analysis of 300W cartridge heaters in a copper block Figure A.8 Photo of heating test section Figure A.9 Figure A.10 Figure A.11 Figure A.12 Wiring diagram for omega PX2300 series differential pressure transducers SWEP Minex M10 chevron plates with gaskets were arranged in alternating fashion SWEP Minex M10 manufacturer drawing with front, side and cutaway views Manufacturer instructions for assembling Minex M10 gasketed plate/frame heat exchanger Figure A.13 Lytron Kodiak chiller control panel Figure A.14 Lytron Kodiak chiller rear panel components Figure A.15 Omega FL4505-V rotameter Figure A.16 Reservoir tank fitting locations on 6 OD Stainless Steel flange Figure A.17 Figure A.18 SWEP Minex M10 Plate/Frame heat exchanger compression plates Detailed design drawing of the coolant loop apparatus showing the main components, pipe dimensions and pressure and temperature port locations. The dashed line represents the breadboard top on which the loop was constructed. All other pipes/components were mounted vertically below the breadboard where the pump is the lowest component xi

12 Figure C.1 Figure C.2 Figure C.3 Figure C.4 Figure C.5 Figure C.6 Figure C.7 Figure C.8 Figure C.9 2.5%wt Alumina/water heat duty comparison. Heater input power is compared to microchannel heat exchanger, process, and cooling water calculated thermal energies % wt Alumina/water heat duty comparison. Heater input power is compared to microchannel heat exchanger, process, and cooling water calculated thermal energies %wt Alumina/water heat duty comparison. Heater input power is compared to microchannel heat exchanger, process, and cooling water calculated thermal energies Flowrate comparison of pump power required for water and Al 2 O Pressure drop comparison of water and Al 2 O 3 Plate/Frame heat transfer coefficient Pressure drop comparison of water and Al 2 O 3 microchannel heat exchanger 1 film H Pressure drop comparison of water and Al 2 O 3 microchannel heat exchanger 2 film H PAO heat duty comparison for system pressure. Heater input power is compared to microchannel heat exchanger, process, and cooling water calculated thermal energies PAO NF heat duty comparison for system pressure. Heater input power is compared to microchannel heat exchanger, process, and cooling water calculated thermal energies Figure C.10 System pressure drop comparison for PAO and PAO NF Figure C.11 Figure C.12 Figure C.13 System pressure drop comparison of plate/frame overall H for PAO and PAO NF PAO heat duty comparison for individual pressure. Heater input power is compared to microchannel heat exchanger, process, and cooling water calculated thermal energies PAO NF heat duty comparison for individual pressure. Heater input power is compared to microchannel heat exchanger, process, and cooling water calculated thermal energies xii

13 Figure C.14 Figure C.15 Total pressure for PAO and PAO NF calculated from individual pressure drop Total pump power for PAO and PAO NF calculated from individual pressure drops Figure C.16 Flowrate comparison of plate/frame overall H Figure C.17 Pump power comparison of plate/frame overall H Figure D.1 Setup for bucket/stopwatch calibration technique Figure D.2 Figure D.3 Figure D.4 Figure D.5 Figure D.6 Figure E.1 Figure F.1 Figure F.2 FL902-G rotameter calibration for water shows that the flowmeter is calibrated correctly FL902-G rotameter calibration for PAO shows that the measured flow is much lower than the indicated flow FL902-G rotameter calibration for PAO NF shows that the measured flow is much lower than the indicated flow FL902-G rotameter calibration for EG/Water mixtures with higher viscosities and densities shows that increases in viscosity significantly reduce the measured flowrate as compares to the observed flowrate FL4504-V rotameter calibration for water shows that the flowmeter is calibrated correctly Indicated and Measured values for heating power. The power was calculated using P=IV, and clearly shows that the voltmeter and ammeter readings are accurate as compared to the value measured with a Fluke 182 True RMS Miltimeter calibrated through AF PMEL Photo of panel meters stacked. This caused a significant error to the thermocouple reading due to the heat transfer from the bottom meter to the top meter Photo of panel meters separate. This test verified the hypothesis that stacking the panel meters was causing an erroneous reading xiii

14 Figure F.3 Figure F.4 Photo of the test setup to calibrate the process and chiller fluid thermocouples. Results showed that when the panel meters were stacked, the indicated temperature was significantly higher than when the panels were separate Plot of stacked and separate temperature difference. The stacked reading is consistently 1 higher than for separate operation xiv

15 LIST OF TABLES Table 2.1 Summary of single-phase Nanofluid convective heat transfer... 6 Table 2.2 Plate/Frame Nusselt correlation constants for water and PAO Table 3.1 Table of uncertainty values for direct measurements Table F.1 Calibrations used for thermocouple calibrations Table G.1 Water property data from McCabe, Smith and Harriott (McCabe et al. 2005) Table G.2 Nyacol/Water thermal conductivity data Table G.3 Nyacol/Water viscosity data Table G.4 Nyacol/Water density data Table G.5 Nyacol/Water heat capacity data Table G.6 PAO and PAO NF thermal conductivity data Table G.7 PAO and PAO NF viscosity data Table G.8 PAO and PAO NF density data Table G.9 PAO and PAO NF specific heat data Table H.1 Water data Table H.2 2.5% wt Al 2 O 3 data Table H.3 5.0% wt Al 2 O 3 data Table H % wt Al 2 O 3 data Table H.5 PAO run #2 data xv

16 Table H.6 PAO run #3 data Table H.7 PAO run #4 data Table H.8 PAO run #5 data Table H.9 PAO run #6 data Table H.10 PAO run #7 data Table H.11 PAO run #8 data Table H.12 PAO run #9 data Table H.13 PAO run #10 data Table H.14 PAO run #11 data Table H.15 PAO NF run #1 data Table H.16 PAO NF run #2 data Table H.17 PAO NF run #3 data Table H.18 PAO NF run #4 data Table H.19 PAO NF run #5 data Table H.20 PAO NF run #6 data Table H.21 PAO NF run #7 data Table H.22 PAO run #12 data Table H.23 PAO run #13 data Table H.24 PAO run #14 data Table H.25 PAO run #15 data Table H.26 PAO run #16 data Table H.27 PAO run #17 data Table H.28 PAO NF run #8 data Table H.29 PAO NF run #9 data xvi

17 Table H.30 PAO NF run #10 data Table H.31 PAO NF run #11 data Table H.32 PAO NF run #12 data Table H.33 PAO NF run #13 data Table H.34 PAO NF run #14 data Table H.35 PAO NF run #15 data xvii

18 LIST OF SYMBOLS Cross sectional area, Heat capacity, Diameter, f Friction factor, dimensionless Gz Graetz Number, dimensionless h Heat transfer coefficient, Average heat transfer coefficient, H Height, I Current, k Thermal conductivity, L Duct/channel length, Mass flowrate, n Number of channels xviii

19 Nu Nusselt Number, dimensionless P Perimeter, Pressure difference, Pr Prandtl Number, dimensionless q Heat flux, Q Heat transfer, Re Reynolds Number, dimensionless T Temperature, t Plate thickness, Velocity, Average velocity, U Overall heat transfer coefficient, V Voltage, Flowrate, W Width, Work, w i Uncertainty, units to match i Entrance length for fully developed flow, xix

20 Chevron angle, degrees Density, Viscosity, SUBSCRIPTS c Cold stream C Cross-sectional ch Channel eff Effective f fluid h Hot stream H Hydraulic ht Heat transfer lm Log Mean m mean p Pump s surface xx

21 w Wall ABBREVIATIONS AF PMEL Air Force Precision Measurement Equipment Laboratory EG Ethylene Glycol H Heat Transfer Coefficient HX Heat Exchanger HX1 Microchannel Heat Exchanger 1 HX2 Microchannel Heat Exchancer 2 LMTD Log Mean Temperature Difference NF Nanofluid PAO Poly-alpha-olefin, synthetic oil used for avionics cooling Plate/Frame RMS Root Mean Square SBIR Small Business Innovative Research SG Specific gravity SS Stainless Steel xxi

22 Thermocouple WPAFB Wright Patterson Air Force Base xxii

23 1. INTRODUCTION 1.1 Background Thermal management is one of the largest issues facing today s Air Force. Improved radar systems, more electric aircraft and directed energy weapons with megawatt heat loads are driving the thermal demands further than ever before. Unfortunately, cooling systems have not maintained pace. Previously, engineers could design a larger cooling system to meet the cooling needs. Now, system miniaturization complicates the heat transfer problem in that heat exchangers can no longer simply be made larger to increase surface area available for heat transfer. The result is that many of today s newest aircraft overheat while sitting on the runway because the cooling systems are unable to maintain the sensitive electronics at the proper operating temperature. While the current heat transfer fluid, poly-alpha-olefin (PAO), is the only fluid that meets the Air Force avionics coolant specification (MIL-PRF-87252), its heat transfer characteristics are very poor. The thermal conductivity of PAO is only about 25% that of water and the heat capacity just over 50% that of water. In addition to poor heat transfer properties, the viscosity of PAO is 4-5 times higher than water, which means more energy is required to pump PAO than water at equal flowrates. A drop-in replacement 1

24 coolant for PAO with increased heat transfer ability, and no penalty to the pumping power, is essential to the thermal management of current and future weapons platforms. Choi et al proposed a new class of heat transfer fluids that suspended nanosized metallic particles, commonly known as nanoparticles, into traditional heat transfer fluids in an attempt to increase their heat transfer performance (Choi 1995). These new heat transfer fluids were termed nanofluids. In general, nanofluids consist of <100nm particles suspended in solution with thermal conductivity at least 20 times greater than that of their base fluids, such as water, EG/water or PAO (Wen et al. 2009; Keblinski et al. 2005; Nguyen et al. 2007). Early reports showed that nanofluids enhanced the thermal conductivity greater than predicted by conventional theory. For this reason, nanofluids were expected to result in novel heat transfer fluids with a greater ability to transfer heat over that of their respective base fluids. A successful nanofluid, which is a passive heat transfer solution, has the ability to revolutionize heat transfer systems through size, weight and cost reduction (Wang and Mujumdar 2010) (Wen et al. 2009). Potential applications include the defense, automotive, and energy industries to name a few, all of which have current programs dedicated to the formulation and testing of nanofluids (Zhou et al. 2010; Kim et al. 2009; Ho et al. 2010). While, in certain instances, nanofluids have shown an unusually high thermal conductivity, results vary significantly between research groups. This could be due to the fact that there are many techniques available to measure the thermal conductivity. Also, very little work has been done to determine the potential effects of increased viscosity in a flowing system. Increased viscosity could offset the thermal conductivity and heat capacity increases by also increasing the pumping power. Nearly every research group to date has concluded that 2

25 there is a significant lack of experimental data concerning real-world validation testing of nanofluids (Lee et al. 2008; Wang and Mujumdar 2010; Zhou et al. 2010; Keblinski et al. 2005; Daungthongsuk and Wongwises 2007; Wen et al. 2009; Kim et al. 2009; Nguyen et al. 2007; Ho et al. 2010). Both of these concerns can be addressed though a system-level thermal hydraulic heat transfer evaluation. 1.2 Approach The goal of this work is to design and build a system capable of evaluating the convective heat transfer coefficient for various fluids. Water will be used to baseline the system due to its well-known thermal fluidic properties. Once a water baseline is established various concentrations of Al 2 O 3 particles in water will be tested. Heat transfer coefficient results will be compared to the base fluid. Similarly, PAO will be run followed by a PAO-based alumina nanofluid. As previously stated, a drop-in replacement coolant for PAO that increases the convective H without a penalty to pumping power is desired. There is a great deal of uncertainty surrounding the topic of nanofluids. The results of this work will give researchers insight into their potential use in current/future defense systems, specifically, advanced avionics. The coolant loop apparatus will be used in the future to evaluate the system level heat transfer performance of new coolants. 3

26 2. LITERATURE REVIEW AND MODELING BACKGROUND 2.1 Literature Review Nanofluid Convective Heat Transfer Systems A thorough review of the literature showed that minimal work has been done on single-phase forced convection heat transfer for nanofluids. Of the work done, most used Al 2 O 3 /water nanofluids, and no work was found using PAO-based alumina nanofluids, which will be used in this study (Yu and Liu 2009; Nguyen et al. 2007; Ho et al. 2010; Lee and Mudawar 2007; Wen and Ding 2004). Table 2.1 summarizes these results. Yu and Liu found that for volume concentrations of 2% alumina in water, the maximum convective heat transfer coefficient increase was 10.6%, but reported that overall the enhancement was small (Yu and Liu 2009). Similar results were seen three years earlier by Lee and Mudawar (Lee and Mudawar 2007). Additionally, Yu and Liu showed that the traditional pressure drop and heat transfer equations were useful in predicting their results. On the other hand, for a 1.6% particle volume fraction Al 2 O 3 /water, Wen and Ding showed a up to a 47% increase in convective heat transfer coefficient over the base fluid and concluded that the traditional pressure drop and heat transfer models substantially underpredicted their results (Wen and Ding 2004). Similarly, Nguyen et.al. reported a 40% increase in convective H using a 6.8% volume fraction of Al 2 O 3 in 4

27 water (Nguyen et al. 2007). Although they did not show the data, Yu and Liu were the only group to definitively conclude that nanofluids can enhance the convective H even with the penalty of increased pressure drop. The H as a function of pressure drop will be reported in this study. Others simply showed results as a function of Reynolds number. This is misleading as increased viscosity may have a significant effect on the pumping power required to obtain equal nanofluid and basefluid Reynolds numbers. The results of these four studies show the need for an increase in experimental work to be done in this area, which is the motivation for this research. Recent work done at the Air Force Research Labs (AFRL) by Sommers and Yerkes on the effects of Al 2 O 3 /propanol showed a slight increase in the convective H, but revealed the pressure drop penalty associated with nanofluids as they saw a 400% to 600% increase in pressure drop. However, they observed a decrease in the pressure drop penalty at higher flowrates, which they attributed to the shear-thinning nature of nanofluids (2010). Stability is the most important factor when considering nanofluids for use in a real-world application. Therefore, care should be taken in fluid preparation to ensure that the particles are well dispersed. Without a stable suspension there are many possible adverse effects; such as agglomeration, sedimentation in the system, increased part wear and erosion on metal surfaces (Lee and Mudawar 2007). Another drawback is the breakdown of traditional pressure drop and heat transfer predictions for poorly dispersed 5

28 Table 2.1: Summary of single-phase nanofluid convection heat transfer Reference Base Fluid Particles volume percent size Flow Regime Results (Lee and Mudawar 2007) Water Al 2O 3 1.0, nm 332<Re<1641 Increased h and ΔP for increasing wt% at similar Re (Sommers and Yerkes 2010) Propanol Al 2O 3 1% 10nm 1800<Re<2800 (Ho et al. 2010) Water Al 2O 3 1.0, nm For Re<2300, h of propanol ranged from , for 1% values of ; ΔP increase 400%-600% For 1%vol at Re = 332 and 1641; a 40% and 53% increase respectively in Nusselt number (Nguyen et al. 2007) Water Al 2O 3 1.0, 3.1, nm 200<Re<1000 For 6.8% vol, h increased up to 40% (Fotukian and Nasr Esfahany 2010) Water Al 2O , 0.054, 0.067, nm 6000<Re<31,000 48% increse in h for 0.054% vol at Re=10,000; no effect on h with increasing vol%; 30% increase in ΔP at Re=20,000 (Asirvatham et al. 2009) Water Cu Re=1,350; Re=1,700 8% increase in h (Chandrasekar et al. 2010) Water Al 2O nm Re= % increase in h at Re=2275; no pressure drop penalty (Wen and Ding 2004) Water Al 2O 3 0.6, 1.0, nm 500<Re<2100 (Hwang et al. 2009) (Zeinali Heris et al. 2007) Water Al 2O to 0.3 Water Al 2O 3 0.2, 0.5, 1.0, 1.5, 2.0, nm Fully developed laminar regime 700<Re<2050 8% increase in h for 0.3%; enhancement increases with vol%; friction factor in good agreement with theory h increases with Pe number and vol% and exceeds single phase laminar predictions (Li and Xuan 2002) Water Cu 0.3,0.5,0.8,1,1.2, 1.5, 2.0 <100nm 800 < Re < Increased h with particle vol %; 60% increase in h at 2% as compared to water (Jung et al. 2009) Water Al 2O nm 5<Re<300 32% Increase in convective H (Kim et al. 2009) Water Al 2O 3 3% 20-50nm 800<Re<2200; 3000<Re< % and 20% increase in h for laminar and turbulent flow, respectively 6

29 particles. Most report that a 30 minute sonication time is sufficient to break up agglomerations and ensure proper particle size/dispersion before use.(yu and Liu 2009) Therefore, it will be necessary to apply this procedure to each nanofluid prior to running an experiment. Nanofluid heat transfer is a widely debated topic for which multiple predictive methods exist for thermal properties alone. This complicates the procedure for predicting the thermal hydraulic performance. The goal of this study was not to predict thermal properties, rather it was to measure the convective heat transfer coefficient and pressure drop for flowing nanofluids, and compare these results to pure fluid Nusselt correlations using measured thermal properties of the nanofluids. Therefore, four critical thermophysical properties were measured. These properties included the thermal conductivity (k), viscosity (μ), density (ρ), and heat capacity ( ). For similar work done on a flowing system with single-phase nanofluid heat transfer, Liu and Yu concluded that if the effective thermophysical properties of a nanofluid were measured accurately, then the conventional theory for heat transfer and pressure drop applied and will predicted the experimental results (Yu and Liu 2009). The next section will briefly cover the techniques that will be used to measure the thermal fluidic properties of nanofluids. Based on literature, the expected trends in nanofluid properties as compared to their base fluids will be discussed Nanofluid Thermal Fluidic Properties Thermal conductivity is the most controversial nanofluid property based on reported values that are higher than predicted by classical models such as Maxwell or 7

30 Hamilton & Crosser (Keblinski et al. 2005). Additionally, there are great discrepancies in the literature for reported thermal conductivity enhancement ranging from 1% up to 150% (Warnakulasuriya and Worek 2008). Brownian motion of particles, particle size, particle loading, particle type, temperature, thermophoresis, aspect ratio and fluid stability have all been cited as possible reasons for the inconsistent results but will not be discussed here. In this research, the nanofluids were assumed to be well-dispersed solutions with spherical particles. Recently, a study done by 30+ organizations worldwide on the thermal conductivity of nanofluids found Maxwell s effective medium theory to be in good agreement with the experimental data (Buongiorno et al. 2009). In a 2008 review of nanofluids, Wang et.al. compared the results from nearly 20 groups studying the thermal conductivity of nanofluids. The reported values for thermal conductivity improvement varied from 12% - 300%. The greatest enhancement was reported for Carbon Nanotube (CNT) suspensions, which will not be studied here because there were no stable PAO-based nanofluids readily available. Typically, water is used as a base fluid and fewer studies have been conducted with either ethylene glycol (EG) or oil as a base fluid. The thermal conductivity of a pure fluid can be measured in many different ways. The transient hot wire (THW) method of measuring thermal conductivity is the most widely accepted and will therefore be used here (Buongiorno et al. 2009; Wang and Mujumdar 2008; Kim et al. 2009). Narvaez did a comprehensive study of available techniques to measure the thermal conductivity of PAO-based nanofluids and found that the THW method was the most accurate and precise (Narvaez 2010). Although there are discrepancies in the magnitude of thermal conductivity increase, it is important to note 8

31 that no one has reported a reduction in thermal conductivity when comparing the nanofluid to its base fluid. Therefore, it is expected that the thermal conductivity of the nanofluids in this study will be greater than that of their respective base fluids and will increase with particle loading (Narvaez 2010). Tables G.2 and G.6 show the thermal conductivity values obtained for the water and PAO-based nanofluids, respectively. Heat capacity and density predictions for nanofluids are straight-forward and less debated than thermal conductivity and viscosity. Ho et.al. use the following equation to predict density: where,,, and are the nanofluid density, fluid density, particle density and particle volume fraction, respectively (Ho et al. 2010). This is a linear prediction based on the particle loading. Since the density of solids is higher than for liquids, it is expected that the density of the nanofluids will be greater than the base fluids. Table G.4 and G.8 show the density values obtained for the water and PAO-based nanofluids, respectively. For water/al 2 O 3 fluids, no technique or reported data was available to measure the densities at various temperatures. However, the density of water changes very little within the experimental temperature range. Therefore, only one density measurement was taken using a 10mL pycnometer at room temperature. The heat capacity of pure alumina is lower than water and PAO, so it is expected that the nanofluid heat capacity will be slightly lower than for the pure fluid. The most common and reliable method of measuring the constant pressure heat capacity,, is 9

32 using a Differential Scanning Calorimeter (DSC). Table G.5 and G.9 show the heat capacity values obtained for the water and PAO-based nanofluids, respectively. The property of most interest in this study was the viscosity of the nanofluids and its effect on the convective heat transfer coefficient at equal pressure drop. For the most part, reported values for nanofluid viscosity are rare and no reliable model is available for prediction (Murshed et al. 2008). Perhaps that is because there are an increasing number of reported results showing that classical models such as the Dittus-Boelter and Einstein equations substantially underpredict the effective viscosity (Wang and Mujumdar 2010; Wen et al. 2009). Although, the values for viscosity cannot be predicted, there are some predominant trends found in literature. First, the relative viscosity of nanofluids is higher than traditional predictive models. Next, the nanofluid viscosity increases with increased particle size and loading. Finally, relative viscosity of nanofluids is independent of temperature (Zhou et al. 2010). Measured values from a Brookfield controlled rate viscometer will be used to obtain the base fluid viscosities (Narvaez, 2010). Tables G.3 and G.7 show the viscosity values obtained for the water and PAO-based nanofluids, respectively. 2.2 Modeling Background Convective Heat Transfer Coefficient The overall heat transfer coefficient of a fluid is a function of four thermal fluidic properties: thermal conductivity, viscosity, heat capacity and density. Assuming steady state conditions for an incompressible fluid with no elevation change, no shaft work, no 10

33 irreversibilities, a small pressure change, and a constant heat capacity, the first law for a flowing system becomes: Eqn 2.1 where is the temperature difference is between the fluid bulk inlet and outlet temperatures. The mass flowrate,, is equal to the volumetric flowrate times the density. Convective heat transfer occurs where a fluid is moving past a solid surface. At that solid-fluid interface, the heat flux is related to the difference in the solid wall temperature,, and the bulk fluid temperature,. This relation is known as Newton s law of cooling. It is a defining equation for the h, which is called the film heat transfer coefficient (H): Eqn 2.2 where is the total heat transfer. For a heat exchanger, the bulk fluid temperature is typically taken as the average of the inlet and outlet temperatures of the fluid. The film H is a function of the fluid properties, flow conditions and geometries. As a fluid flows over a surface with a constant wall heat flux a thermal boundary layer grows in the direction perpendicular to flow. The local convective heat transfer coefficient decreases in the direction of flow because the fluid temperature increases in the direction of flow and decreases the driving force for heat transfer. The average film H,, can be calculated using the entire surface area, in Eqn 2.2. The average H is an indication of how well a fluid will transfer heat given a particular driving force, or temperature gradient, flow conditions and geometry. The heat transfer,, 11

34 calculated from Eqn 2.1 and Eqn 2.2 are equal when an average film H is used. It is common practice to relate the two equations in order to obtain the measured film H. The convective H will vary between laminar and turbulent flow. Laminar boundary layers are characterized by orderly flow, whereas fluid motion in turbulent flows is more random due to eddies in the flow. The irregular motion of fluid particles in turbulent flow greatly increases fluid mixing and enhances the energy transfer. For this reason, the convective H for turbulent flows is greater than for laminar flows. This section will present the equations and assumptions necessary to perform a proper H analysis for laminar and turbulent flow Film Heat Transfer Coefficient for Fully Developed Pipe Flow When designing an experimental apparatus, it is typical to predict the performance in order to validate the system results. The traditional dimensionless form of the heat transfer coefficient is the Nusselt number. The Nusselt number is the ratio of convective to conductive heat transfer: Eqn 2.3 where is the characteristic length and k is the thermal conductivity of the fluid, evaluated at the bulk fluid temperature. Inspecting Eqn 2.3 shows that for pure conduction,. If that means that the heat transfer is enhanced by convection. Higher Nusselt numbers indicate a higher heat transfer at the surface. The Nusselt number is commonly predicted from empirical correlations. Therefore, experiments must be done to obtain correlations based on flow conditions (i.e. 12

35 laminar or turbulent) and geometry. The most general correlation for forced convective flow is given in Eqn 2.4. The empirical constants C, m, and n are independent of the fluid, but dependent on the geometry and flow condition: Eqn 2.4 Where the Reynolds number, Re, is ratio of momentum to viscous forces: Eqn 2.5 and the Prandtl number, Pr, is the ratio of momentum and thermal diffusitivies: Eqn 2.6 In the case of constant heat flux conditions for fully developed laminar flow in a circular tube, Nu can be approximated as 4.36 and the film H can be solved for using Eqn 2.3. When the surface condition is constant temperature, Nu is For applications that involve convective heat transfer in noncircular ducts, it is necessary to calculate the hydraulic diameter. The hydraulic diameter,, is defined as four times the ratio of the flow area to the wetted perimeter: Eqn 2.7 where A is the cross sectional area of flow and P is the wetted perimeter. The hydraulic diameter is important for characterizing flows in noncircular ducts where a tube diameter does not exist. For the case of a circular duct, the equation simplifies to the diameter. The Reynolds number is important in all flowing systems as it allows one to quickly predict the flow regime. It is valid for all fluids and geometries given the correct characteristic length is used. For pipe flow, the diameter is used as the characteristic 13

36 length. Laminar pipe flow is characterized by and turbulent pipe flow occurs when. The Reynolds number is also used to determine the entrance length,, required for hydraulically fully developed flow. Hydraulically fully developed flow occurs at the point where the fluids velocity changes are based only on the distance from the centerline, r, and is independent of the axial length. This velocity profile will remain unchanged until the character of the pipe or fluid changes. Fully developed flow allows the assumption of steady flow to be made and simplifies the Nusselt correlations. For laminar flows: Eqn 2.8 If, where L is the duct length, then the flow is fully developed because the duct is long enough for the flow to reach its fully developed state. Conversely, if, the flow is not fully developed and additional correlations are necessary to predict the heat transfer coefficient. For turbulent flow, Eqn 2.9 where D is the pipe diameter. The Prandtl number (Eqn 2.6) is a fluid transport property dependent on the fluid and its current state, i.e. temperature and pressure (Michael J Moran et al. 2003). Typical values are around 7 for water at room temperature and between 100 and 40,000 for engine oils. For small Pr numbers, heat diffuses very quickly compared to the velocity (momentum), as in gases. For liquids, higher Pr numbers indicate that momentum transport has a greater effect on the heat transfer. 14

37 If an average convective H is desired for fully developed turbulent flow, the Reynolds and Prandtl number must be known to predict the Nusselt number using an appropriate Nusselt correlation. Perhaps the most well-known H correlation is the Dittus-Boelter equation: Eqn 2.10 where the exponent n is 0.4 in the case of surface cooling and 0.3 for a fluid heating a surface. The Dittus-Boelter equation is valid for and. In situations where there is more than one flow channel in parallel, it is valid to assume an average velocity based on the flowrate ( ), number of channels (n) and the channel area ( ): Eqn 2.11 It is worthwhile to note that using the velocity shown in Eqn 2.11 results in an average Reynolds number when used in Eqn 2.5. For a single flow channel, n =1, so that the tube velocity, u, reduces to Heat Transfer Coefficient for Thermally Developing (TD) Pipe Flow The Nusselt number for laminar flow in a channel remains constant only for fully developed flows. There are two types of fully developed flow hydraulically and thermally. For laminar pipe flow, the Reynolds number is used to determine the length required for hydraulically fully developed flow using Eqn 2.8. A flow can be fully developed hydraulically and remain thermally developing (TD). In thermally fully 15

38 developed flow, the dimensionless temperature ratio remains constant in the direction of flow: Eqn 2.12 where is the surface temperature, is the mean fluid temperature and is the local fluid temperature. In this case, the film H is constant and independent of x. When the flow is thermally developing, the average Nusselt number can vary with Reynolds number. The Graetz number is used to determine if the entrance effects of velocity or temperature are negligible. Eqn 2.13 For, the flow is characterized as thermally developed. For, the flow is characterized as thermally developing and a different Nusselt correlation is required to account for entrance effects (McCabe et al. 2005). For TD flow ( ), a Nusselt correlation based on the Gz number is used so that: Eqn 2.14 where The term is the ratio of the bulk fluid viscosity and the viscosity at the wall. In most cases, this term is equal to unity due to negligible changes in temperature between the wall and the bulk fluid. For laminar, thermally undeveloped duct flow, Eqn 2.14 is used in place of a correlation in the form of Eqn 2.4 to determine the film H. 16

39 Overall Heat Transfer Coefficient and the LMTD Predicting the heat transfer coefficient of heat exchangers is slightly different than for pipe flow. Heat exchangers are devices that exchange heat between two fluids separated by a solid wall. For a heat exchanger operating at steady state with negligible potential and kinetic energy and negligible heat loss, the thermal energy transfer between the two fluid streams must balance, i.e. the amount of heat removed from the hot fluid must be the same as the amount gained by the cold fluid as calculated by Eqn 2.1. For heat exchangers, an overall heat transfer coefficient, U, is used in place of the local or average film H. The overall H is necessary because the difference between the two fluid streams changes within the heat exchanger. Therefore Eqn 2.2 for heat exchangers becomes: Eqn 2.15 where is the mean temperature difference between the streams and A is the heat exchange surface area. Similarly, as for the film H, Eqn 2.1 and Eqn 2.15 must balance, and the measured overall H can be obtained. The mean temperature difference is based on the heat exchanger configuration (i.e. counterflow, parallel flow, etc). For this work, a counterflow arrangement was used due to its high thermal efficiency. The appropriate takes the form of the log mean temperature difference (LMTD), Eqn

40 where outlet and is the temperature difference between the hot stream inlet and cold stream is the temperature difference between the hot stream outlet and cold stream inlet. The overall H is predicted using a representative thermal circuit based on the hot and cold convection coefficients and the wall thermal conductivity so that for cartesian coordinates: Eqn 2.17 where and are determined as outlined in section using the appropriate Nusselt correlation. The wall thickness is represented by t. As previously discussed, Nusselt correlations are experimentally determined based on geometry and flow conditions. This work involved the use of a gasketed plate frame heat exchanger which consist of thin, pressed sheet metal plates (made of stainless steel) that separate the hot and cold fluid streams. Gaskets separate the two fluids and set the channel spacing between the plates. Compression plates on either end hold the inlet and outlet ports and provide the force required on the gaskets to prevent fluid mixing. Figure A.10 provides a picture of a plate with gasket. Corrugations in the plates increase the heat transfer area and induce turbulent flow at low Reynolds numbers. Thus, they significantly increase the thermal-hydraulic performance of the heat exchanger. Their modular design allows for quick changes to cooling capacity simply by adding or removing plates. The geometry of the plates (i.e. corrugations, chevron angle) plays a major role in the heat transfer performance and in turn affect the associated Nusselt correlations. 18

41 The appropriate plate/frame Nusselt correlations for and are those that result in the best approximation of the measured overall H from Eqn For chevron-type plates, the Nusselt correlations are generally developed based on flow conditions and chevron angle (β). A literature review done by Ayub found that over 30 practical Nusselt correlations exist for heat exchangers (Ayub 2003). Therefore, the correlation constants for and are not necessarily the same and will be different for hot stream and cold stream fluids with dissimilar thermal fluidic properties and flowrates Sensitivity Analysis of the H From section 1.1, the H is a function of four thermal-fluidic properties,. The goal of this work was to determine the relative difference in H,, between a nanofluid and its base fluid. The notation represents the nanofluid. If, then the H was increased by the presence of nanoparticles, if, the H was lowered. The thermal fluidic properties of interest are a function of the temperature. Therefore, for each sensitivity analysis, the fluid properties must be evaluated at the same bulk fluid temperature. To validate experimental results, a sensitivity analysis was performed to determine the degree to which each measured fluid property will affect the H. In this sensitivity analysis, similar properties were grouped so that the nanofluid property was in the numerator. The resulting exponent on each property group determined the sensitivity of the calculated value to that property. Section 1.1 outlined the necessary equations and assumptions used to determine the film and overall H. This section builds upon that work to develop the equations used to predict for laminar, thermally developing flow in a non-circular duct as well 19

42 as for a plate and frame heat exchanger. First, a sensitivity analysis on the pressure drop,, and resulting pumping power,, will be presented Pressure Drop and Pumping Power In order for a fluid to perform well as a drop-in replacement, it must achieve a higher heat transfer coefficient at the same system pressure drop. The pressure drop is a function of the geometry, velocity, viscosity and density. Eqn 2.18 Where f is the Fanning friction factor. The general form for the friction factor is shown below. Eqn 2.19 The friction factor is dependent on the flow regime so that p is 1.0 for laminar pipe flow, 0.3 for turbulent pipe flow. The constant also varies for laminar and turbulent flow, but is not important in the sensitivity analysis as it is cancelled out. The sensitivity analysis for Eqn 2.18 shows that at equal flowrates: Eqn 2.20 For equal flowrates, the term goes away. For laminar flow, the density term goes away because and the increase in pressure drop is directly proportional to the increase in viscosity. For turbulent flow, the pressure drop increase is more dependent on the increase in density, although viscosity does play a role. With an expected increase in 20

43 viscosity and density, it should follow that the pressure drop of the nanofluids will be higher than for the base fluids. Therefore, it is expected that. It is also appropriate to solve to the flowrate ratio at equal pressure drops for the nanofluid and the base fluid so that: Eqn 2.21 For laminar flow, the flowrate ratio is then proportional to the change in viscosity. However, for turbulent flow, where p is 0.3, the flowrate ratio is affected by both viscosity and density changes. Eqn 2.22 A sensitivity analysis should now be performed on the system pumping power so that total system power required for each fluid can be compared. From the mechanical energy balance for flowing systems with no shaft work, no elevation change, equal velocity, no frictional losses and an incompressible fluid, the pump power,, is calculated by: Eqn 2.23 The assumption of no shaft work is valid because the control volume evaluated is not the pump itself. Rather, the control volume evaluated here is an experimental test section for which the power required to obtain a certain flowrate and pressure drop through that section is sought. The sensitivity analysis of Eqn 2.23 at equal flowrates shows that the relative increase in pumping power is equal to pressure drop: 21

44 Eqn 2.24 Therefore, Eqn 2.24 shows the same findings as Eqn 2.20 in terms of overall effect of viscosity and density and it is expected that Laminar, Thermally Developing Flow From section 1.1.3, the Gz number is required to predict the film H in thermally developing flow. The Nu number of Eqn 2.14 is used to solve for the film H. An analysis was first done with water to ensure that Eqn 2.14 accurately predicts the measured film H and is valid for the current heat exchanger. Due to the small channels and aluminum construction, the fin heat transfer area,, must be used in Eqn 2.2, rather than only the top and bottom surface area. (Park and Punch 2008; Rosa et al. 2009; Hegab et al. 2001; Wang and Peng 1994) Referencing Figure 2.1, the heat transfer area becomes: Eqn 2.25 Where n is the number of channels. Using the fin heat transfer area, Eqn 2.14 predicted well the heat transfer coefficient for water. Therefore, the Graetz number approximation for thermally developing flow applies. Therefore, it was used to perform the sensitivity analysis. 22

45 Figure 2.1: Schematic of microchannel heat exchanger (number of flow channels not to scale) For this application, the difference in flowrates at equal pressure drop is sought. Therefore, the flowrate is included in the calculation to determine its effect on the film H. Eqn 2.26 For equal flowrates, the term goes to unity and is only a function of the measured property values for density, heat capacity and thermal conductivity. If the nanofluid flowrate ( ) required for a given pressure drop is less than the base fluid, it will decrease the value of as compared to equal flowrates. Eqn 2.26 predicts that the thermal conductivity will have the largest affect on the film H. The flowrate, density, and heat capacity all equally affect the film H. If Eqn 2.26 is evaluated at equal pressure drops, Eqn 2.21 can be used so that: 23

46 Eqn 2.27 It is evident now that at equal pressure drops, any increase in viscosity of the nanofluid will reduce the overall heat transfer. For laminar fully developed flow,. Thus the sensitivity analysis at equal flowrates reveals that: Eqn 2.28 Therefore, the relative increase in heat transfer coefficient is directly proportional to the thermal conductivity increase and no other fluid properties as in TD flow. For fully developed flow, then it is expected that Plate and Frame Heat Exchangers Relevant correlations (based on Re and β) from Ayub s literature review were compared to the data collected for the water and PAO experiments and the measured overall H of Eqn 2.17 to determine the C, m, and n for Eqn 2.4 that provided the closest prediction of the measured overall H of Eqn It was found that the constants given in given by Vaie (Vaie 1975) best fit the data of these experiments for water. Kumar s constants for low Re number fit the data for PAO better and are used to predict its performance (Kumar 1984). Table 2.2 summarizes these results. 24

47 Table 2.2: Plate/Frame Nusselt correlation constants for water and PAO C 1 m n Re β Water < Re < 3000 n/a PAO < Re < The sensitivity analysis for the plate/frame HX from Eqn 2.17 shows that: Eqn 2.29 where Eqn 2.30 and the remaining property ratios come from the hot stream convection coefficient,. Using the constants from Table 2.2 and fluid properties found in Appendix G, one can predict the overall enhancement of adding nanoparticles to the base fluid. Eqn 2.29 shows that the exponent on the relative viscosity is negative. Thus increasing the viscosity will decrease the overall H. All other exponents are positive and any increase in the relative values will increase the H. Of these, the heat capacity is expected to play the smallest role and the thermal conductivity again has the greatest effect. The ratio accounts for the fact that the overall H is a combination of both the hot and cold stream convection coefficients. As written, it assumes that the wall conduction term in Eqn 2.17 is small. It is important to note that Eqn 2.17 is a sum of resistances. Therefore, if, then the cold stream resistance is much higher than 25

48 the hot stream resistance. This means that the overall H is dominated by the cold stream. Conversely, if, then the overall H will be dominated by the hot stream coefficient,. This means that changes in the cold stream H will be negligible compared to the hot stream. In this case, and the overall H coefficient is proportional to the ratio of hot stream convection coefficients,. For comparisons at equal pressure drop, Eqn 2.22 can be substituted in for the flowrate ratio to obtain a prediction based solely on property values so that: Eqn 2.31 Eqn 2.31 reveals that the heat transfer for equal pressure drops will be lower than at equal flowrates as the exponents on both the density and viscosity ratios are reduced from Eqn

49 3. DESCRIPTION OF APPARATUS AND EXPERIMENTAL PROCEDURE 3.1 Description of Apparatus The coolant loop was designed to model a typical aircraft avionics cooling loop. Exact sizes, flowrates, pressure drops were not followed. Rather, the objective was to construct a system that would subject fluids to similar conditions that the current heat transfer fluid (PAO) sees. Several design considerations were used. Since experimental fluids are costly and often difficult to obtain in large quantities, the total system volume was minimized to approx 0.75L. The heat exchangers were chosen based on the ability to have both laminar and turbulent test sections. Lastly, prior experience with nanofluids testing at WPAFB has revealed that nanoparticles cannot be removed easily from surfaces. Therefore, only components that were expendable or could be easily broken down and rag cleaned were considered. The 3/8 stainless steel tubes that make up the loop were connected with Swagelok fittings. The resulting apparatus is the continuously flowing system shown in Figure 3.1. It consists of a Haight gerotor pump, heating test section, cooling test section, fluid reservoir tank, and a recirculating chiller. The apparatus allows the user to measure the overall heat transfer coefficient for any fluid as a function of the system pressure drop at various flow conditions. 27

50 Figure 3.1: Schematic of coolant loop apparatus *Note: for this section all figures referenced are found in Appendix A except for Figure 3.1. A detailed drawing with dimensions is included at the end of Appendix A in Figure A.18. The fluid was stored in the reservoir tank, which also acted to remove air after filling. It was then drawn into a gerotor pump, which is a gear pump consisting of an inner and an outer ring, having N and N+1 teeth, respectively. During part of the rotation cycle, the mismatch in the number of teeth creates a partial vacuum and thus fluid is taken into the pump housing. As the gears continue to turn, another opening follows, allowing the fluid to be pumped out of the cavity. Figure A.1 shows a diagram of a 28

51 gerotor pump. For this project, a Haight HGT 6GPM stainless steel gerotor pump (Figure A.2) was used in conjunction with a 1-1/2 HP DC motor and a Leeson DC controller (Figure A.3) to precisely control the flow between 0-2 GPM at a maximum of 1750 rpm. A gerotor pump was chosen because this is the style of pump typically seen in aircraft avionics cooling systems. As the fluid continued out of the pump, it flowed through an Omega FL902G GPM heavy duty stainless steel rotameter which displayed the flowrate. A by-pass section added around the flowmeter provided the option of being able to reverse the pump flow when needed. The ability to reverse the flow direction was helpful when draining the apparatus, as the forward direction did not completely flush the system. Reversing the flow allowed for the last small amount of fluid to be removed before the system was broken down to clean. Additionally, a pressure relief valve was added as a safety measure in case of a partial or full blockage. After the going through the flowmeter, the fluid entered the heated test section. The heated test section used two Lytron CP20 Aluminum cold plate heat exchangers in parallel. They are labeled HX1 and HX2 in Figure 3.1. Chosen for their small size and low cost, they were treated as expendable items and were replaced for each new fluid. A detailed photo of the CP20 is shown in Figure A.4. The small flow passages forced the flow to be laminar, with Reynolds numbers typically below 100. One drawback to the CP20 s was that they were only suitable for a maximum of 1GPM water flow. Therefore, two CP20 s were used in parallel to allow the system flow at 2GPM. It was assumed that the flow was equally split between them, and this can be verified by the fluid temperature rise of each side. They each had 2 heat transfer sides at 2 x 2 for a total of 8in 2 heat transfer area per heat exchanger. The small area required a high heat flux to obtain an 29

52 adequate temperature rise. A specially designed heater control was used to provide up to 6 kw of heating power (see Figure A.5). It consisted of twenty 300W cartridge heaters equally divided and placed into 4 copper blocks (see Figure A.6) which sit on the top and bottom of each heat transfer surface of the heat exchangers. Finite element analysis was conducted on several cartridge heater arrangements in order to determine the best arrangement based surface temperature, minimization of conductive heat losses and ease of machining. It was determined that the arrangement shown in Figure A.7 was the best option. Thermally conductive silicone grease was used to ensure good contact between the block and the heat exchanger, which were also polished to reduce interfacial losses. Small Macor blocks were placed on the copper blocks to minimize heat loss through the top of the blocks. All components were then clamped together and insulated with fiberglass wrap. A photo of the assembled heating test section (without fiberglass insulation) is shown in Figure A.8. The heated section required a great deal of instrumentation. Each heat exchanger was instrumented at the inlet and outlet with an Omega sheathed T-type thermocouple located approximately in the middle of the flow to measure the bulk fluid temperature. Additional Omega.020 T-type thermocouples were strategically placed into machined channels on the heat exchanger side of the copper block to measure the wall temperature as well as to determine temperature distribution at the wall. Omega 6- Channel DP462 panel meters were used to display the temperatures. The pressure drop across this section was measured between pressure ports and as shown in Figure 3.1. An Omega PX2300-5DI, 0-5psi differential pressure transducer was used in conjunction with an Omega U24Y101 24V DC power supply, and displayed with an 30

53 Omega DPS3104-C 4-Channel scanner. A wiring diagram is included in Figure A.9. Cole Parmer ¼ OD clear perfluoroalkoxy (PFA) tubing combined with Swagelok fittings connected the transducer to the loop. The clear PFA tubing allows one to visibly observe if the lines are completely bled of air. The heater control unit has a voltmeter and an ammeter to determine the total heating power being delivered to the copper block. Due to the continuously flowing conditions of the coolant loop apparatus, it was necessary to cool the process fluid upon leaving the heating test section. Therefore, a cooling test section was used. The cooling test section used a SWEP Minex M10 gasketed plate/frame (GPF) heat exchanger to remove heat from the process fluid. The GPF was assembled in a counterflow arrangement with seven plates and gaskets (Figure A.10) so that there were three hot and three cold stream channels. Detailed design and assembly instructions for the GPF are shown in Figure A.11 and A.12, respectively. Four ports on the top compression plate provided the inlet and outlet for the heated process fluid and the chiller water. A 6kW Lytron Kodiak recirculating chiller provided the chiller water to the GPF. The water temperature was controlled from the front control panel (see Figure A.13). Garden hose fittings were attached on the rear of the chiller as shown in Figure A.14 to deliver the water to and from the heat exchanger. The main advantage to a GPF was its ability to be easily broken down and cleaned. It was also much smaller than other heat exchangers because of its high efficiency. This is because the corrugated plates induce turbulent flow for most flow rates. Instrumentation for the cooling section was not as involved as it was for the heated section. Again, sheathed T-type thermocouples were placed into the flow to measure the bulk fluid temperature into and out of the heat exchanger for both the 31

54 process ( Inlet, Outlet on Figure 3.1) and chiller water (Supply and Return on Figure 3.1). An Omega FL4505-V, 0-5GPM acrylic rotameter with valve was used to measure and control the chiller water flowrate (See Figure A.15). Additionally, an Omega PX DI, 0-25psi differential pressure transducer was used to measure the pressure drop between pressure ports and as shown on Figure 3.1. The output was displayed on the same 4-channel scanner as the 0-5psi transducer. The 0-25pdi differential pressure transducer could be re-configured to measure the system pressure drop by connecting to pressure ports and. This allows a system pressure drop and pump power comparison of fluids. After being cooled, the fluid returned to the reservoir tank, which served several purposes. First, it acted as an air separator. Numerous attempts were made to minimize volume by bleeding the system with a bleed valve. However, the air could not be completely removed thus a separator was installed. Second, the tank made filling and draining much simpler. Finally, if a fluid sample was taken during operation, the fluid volume in the tank was able to replace the lost fluid. Also, if so many samples were taken such that the total system volume was depleted, fluid could be added into the tank while the apparatus was in operation. The tank was made from a MDC Vacuum 6 OD 316 SS nipple with Viton elastomer gaskets. Two blank 316 SS flanges closed the ends of the tank and several ports were drilled and tapped into the flanges for an inlet and outlet port, a fluid level indicator and a pressure relief valve as shown in Figure A.16. The fluid level indicator connected into ports on the bottom and top flanges with Swagelok fittings and clear PFA tubing. The pressure relief valve as shown in Figure 3.1 was connected back to the reservoir so that the process fluid would flow back into the 32

55 reservoir in case of a blockage. Including a partially filled reservoir, the total charged system volume was roughly 0.75L. As designed, the coolant loop is a tremendously versatile apparatus. The frame was made of steel Unistrut framing material that is strong, adjustable and requires no welding. Caster wheels enable the loop to be easily moved. The coolant loop was constructed on a Vere, Inc 3 x5 x2 optical breadboard with 1/8 ferromagnetic stainless steel top and back skins and honeycomb core. A standard ¼ -20, 1 hole pattern permits system modifications to be made without affecting the overall integrity of the apparatus. A photo of the final assembled coolant loop test apparatus is shown in Figure 3.2. Figure 3.2: Photo of completed coolant loop apparatus 33

56 3.2 Experimental Procedure A great deal of thought and trial and error went into the experimental procedures. Only details of the data collection procedure will be given in this section. A complete list of procedures can be found in Appendix B and will be referenced in this section. To determine the H at various flow conditions, the following parameters must be measured: the process and chiller flowrate, the system and/or individual pressure drop, fluid temperature into and out of both heat exchangers, the wall temperature of the microchannel heat exchangers, and the heater voltage and amperage. This section outlines the data collection process. The data collection took a considerable amount of time due to the manual operation of all equipment. Manual operation was chosen due the sensitivity of coolant loop systems and the need for human oversight during operation. All data collected was first handwritten and then copied into Excel so that a hard copy of the data was kept for record. This reduced the risk of errors associated with simply typing raw data into Excel and allowed a hard copy of the data to be preserved. A complete log of all raw data can be found in Appendix G. Proper safety precautions were taken (see Safety Procedure, App B) with every experiment. Before collecting data, the coolant loop working surface was cleared of any unnecessary items and absorbent matting laid down in case of a spill. Depending on the status of the loop, any number of procedures may be required before collecting data. If starting with a new fluid and a clean apparatus, the filling procedure must be run to charge the system (see Filling Procedure, App B). Then the pressure transducer bleeding procedure must be run to remove the air from the 34

57 connecting lines and ports. After that, the system was ready for data collection. If the system was already charged at the beginning of a test, then only the start-up procedure was run. Prior to testing, the microchannel HX s were wrapped in fiberglass insulation to minimize heat loss to the ambient. Once these preliminary steps had been accomplished, the pump and heaters were started to begin collecting data. Each complete run consisted of the temperature, flow, pressure and heat input data for a minimum of 5 individual process flowrates. It was decided early on that at least five flowrates were necessary for each fluid in order to reasonably determine any trends. For all tests, the heating power was held constant and the process flowrate was varied. For each test condition, the cooling water flowrate was adjusted such that a common 1 was obtained throughout the test. 1 measured the temperature into microchannel HX1 as shown on Figure 3.1, and was typically held between For water and water-based fluids, a flowrate range of 0.5GPM 2.0GPM was tested with 0.25GPM increments. For PAO and PAO-based fluids, the flowrates varied from 1.0GPM 2.0GPM, in increments of 0.2GPM. Given its poor heat transfer properties, PAO could not be run at flowrates below 1.0GPM due to overheating the copper heating blocks at the desired heat input rate. For convective heat transfer coefficient calculations steady state operation was essential. Figure 3.3 shows how the overall H increased with time. It took approximately 2 hours for the system to reach steady operating conditions. During the initial 2 hours of operation, the overall H in the increased from 650 W/m 2 K to 770 W/m 2 K. At that point, the overall H reached its maximum. Thus the start-up procedure was modified to allow the system to warm up for 2 hours before collecting 35

58 data. Unfortunately, this was not discovered until part way through the PAO data collection. At that point, approximately one-third of the experiments had been completed. Therefore, those experiments were not included in the data analysis. However, the 2 hour startup time was used on every subsequent experimental run. Additionally, experiments run with and without insulation on the HX showed that insulation made little difference in the time required to reach steady conditions. It was believed that the long start up time is related to time required for the large mass of the GPF compression plate (See Figure A.17) to reach steady state Measured H (W/m 2 K ) Uninsulated Insulated Time (min) Figure 3.3: Amount of time required for the apparatus to reach steady operating conditions is approximately 2 hours from cold startup. Beginning with the lowest flowrate, data was collected at steady state conditions. The flowrate was then increased by a small amount and the system was allowed to return to steady state. In order to maintain a constant inlet temperature (1), the cooling water 36

59 on the GPF was adjusted using the rotameter control valve (V2 from Figure 1) to provide the amount of cooling necessary to maintain a constant microchannel HX1 inlet temperature as measured by 1. Typically, the system took 20-30min to return to steady state after the process flowrate was increased. Once the inlet temperature set point was reached, the data was once again collected. This procedure was repeated up to the maximum flowrate. At that point, the test was concluded. Provided enough fluid volume is available, it is suggested that this process be performed a minimum of five times in order to calculate average values for the measured data. 3.4 Calibrations Verification calibrations were done on all instrumentation except for the pressure transducers. However, a no-flow differential pressure was recorded prior to each run to eliminate any bias error. These calibrations and results can be found in Appendix D-F. Liquid Rotameters are variable area flowmeters that are calibrated for one liquid. For any other liquids, the observed reading must be adjusted to account for property differences. To perform this calibration, the liquid is allowed to flow into a bucket and the time recorded. Knowing the weight difference in the bucket before and after, the time, and the density, allows one to calculate the measured volumetric flowrate which is then compared to the observed flowrate from the rotameter. This technique was accurate and repeatable for any liquid and should be applied for liquids with a specific gravity different than the calibration liquid. The rotameters used in this work were calibrated by Omega for water. To verify that the rotameters were properly calibrated for water, the 37

60 bucket and stopwatch technique was performed. From Figure D.1 and D.6, both rotameters used were properly calibrated for water. The thermocouples were calibrated using a constant temperature recirculating viscosity bath, a high accuracy T-type thermocouple and displayed on an Omega HH509 probe as a reference as shown in Figure F.3. All of the thermocouples for the apparatus that measured a fluid temperature were placed in the bath and a temperature difference from the reference was determined. These calibration experiments revealed a significantly higher temperature reading when the two panel meters were stacked on top of one another as opposed to when they were separated during operation (see Figures F.1 and F.2). This affected the Inlet, outlet, Supply and Return thermocouple readings. Specifically, the Inlet, Supply and Return thermocouple displayed a higher reading because they were displayed on the top meter (see Figure F.1). It is believed that heat from the bottom meter was causing the junction temperature of the top meter to raise, thus causing erroneously high temperature readings. Given that all data had been collected with the meters stacked, it was necessary to go back and correct the temperature data using the individual correction equations developed in Table F.1. More detail and results of this calibration process are given in Appendix F. Since all data analysis was done using a delta temperature, using these calibrations eliminated any bias error in the thermocouple readings and the uncertainty in the measurement was reduced to the precision of the panel meter. Finally, a Fluke 189 True RMS Multimeter, calibrated by the AF PMEL was used to confirm the voltage and amperage readings on the heater control panel (Figure E.1). There was less than a 0.5% error between the indicated and 38

61 measured power input. No further calibration was necessary on the reading from the heater control Viscosity Effects on Flowrate Measurement After the data collection was complete, it was observed that the thermal energy balances for fluids other than pure water were flawed, with the error in the measurement increasing with the particle loading. According to the first law for flowing systems (Eqn 2.1), there were only three possibilities that could cause this issue the flowrate, the heat capacity or the temperature difference. From the water heat balance, where the flowmeter was for a liquid with SG = 1, the heat balance was in good agreement. The pure water flowmeter calibration as shown in Figure D.2, revealed that the indicated value was correct. Further, the heat capacity of water is well-known and if there were an issue with the thermocouples or heat capacity data, the error should be seen in the water heat balance. The change to nanofluids increased both the density and the viscosity. For variable area flowmeters, the force balance on the float must be corrected for changes in density. However, viscosity does not generally affect the force balance on the float because the Re is sufficiently large to maintain a constant drag coefficient. However, if large changes in viscosity drop the Re below 100, the drag coefficient increases sharply. Calibrations were not able to be done on the water based nanofluids due to the limited volume of fluid available. The calibrations done for PAO and PAO NF are shown in Figure D.3 and D.4. These showed an unexpected decrease in measured flowrate if density was the only factor. The traditional rotameter flow correction for changes in density is: 39

62 Eqn 3.1 According to this equation, a rotameter is expected to underpredict the flowrate for a fluid with a lower SG than the calibration fluid. Data was collected under the assumption that the density correction of Eqn 3.1 was adequate to correct the flowrate. However, the heat balance data indicated otherwise. The viscosity of pure PAO is nearly 5 times higher than water, which has a viscosity of 0.001Pa*s at room temperature. The bucket and stopwatch calibration of PAO revealed a lower than expected measured flowrate. To further investigate the role of viscosity, bucket and stopwatch experiments were run with 100, 75, 50 and 25% EG/Water mixtures. EG was chosen because its viscosity at room temperature is about 15 times higher than water, so any viscous effects would be magnified. In addition, the density of pure EG is only 10% higher than water (1000 kg/m 3 ) and decreases as the amount of EG is decreased in the mixture. The results are shown in Figure D.5. Two key observations can be made. First, as the viscosity increases, the measured flowrate decreases when compared to the indicated flow. Second, as viscosity increases, the nonlinear region at low flow rates grows. If the flow reached a Re < 100, then the drag coefficient,, is proportional to for pipe flow. Further, the analysis shows that if the drag coefficient changes (as is does for Re<100), then: Eqn 3.2 The experiments with EG substantiated the predictions from Eqn 3.2. Based on fluid properties at 20 C, the measured flow for pure EG and PAO was predicted to be about 40

63 7% and 45% of the observed flow, respectively. The measured values for EG and PAO are 8% and 32% respectively, confirming that viscosity has an effect. In order to correct the flowrates, three assumptions were made based on the water heat balance results: 1. The copper blocks were well insulated with minimal heat loss, <5% 2. The reported heat capacities were accurate 3. The temperature readings and thermocouple calibrations were accurate By operating under these assumptions, and estimating conduction losses from the copper blocks, an effective flowrate was calculated by balancing the heater input power with the calculated heat transfer to the fluid: Eqn 3.3 where I and V are the measured heater voltage and current. Eqn 3.3 was used to calculate the effective coolant flowrate,, for all fluids other than water. This was unnecessary for water because the flowmeter was calibrated for water and a bucket/stopwatch calibration showed that it was accurate as shown in Figure D.2. Further, the energy balance matched energy gained by the chiller water. 3.5 Uncertainty Information and Calculations Manufacturer uncertainty information can be found in Table 3.1. The uncertainty of each calculated value were determined using the Method of Kline and McClintock 41

64 (Kline and McClintock 1953). They proposed that for values calculated from n- independent variables, x i, using a known functional relationship and if the uncertainties of the independent variables,, are all of the same odds, then the uncertainty of the calculated value, w y, can be determined. Eqn 3.4 This method was carried through all calculation steps necessary to compute a value. For example, the equations for calculating heat transfer, Q, to a fluid uses a measured temperature difference. To find the uncertainty of Q, the uncertainty in ΔT,, is first calculated using the Kline and McClintock Method and then used to calculate the uncertainty of Q. The final uncertainty factor that must be considered is human response time. When using a stopwatch, error exists both in the measured time and in the human response time. Boff et.al., determined that the human response time is between ±0.15 and ±0.50 seconds, depending on the individual. (Boff et al. 1986) 42

65 Table 3.1: Table of uncertainty values for direct measurements. Measured Instrument Variable uncertainty units odds Reference Stopwatch time ± seconds /2 Least count n/a response time ± 0.35 seconds Boff et.al., 1986 Scale weight ± lb m /2 least count NETSCH specific heat ± 0.10 kj/kgk 10% measurement Type-T Omega Temperature temperature ± 1.0 C Thermocouple Handbook Thermocouple Type-T Calibration and Thermocouple (bias temperature ± C assumed panel meter error eliminated) precision (± 0.1) FL902G ( GPM) FL4504-V (0-5 GPM) flow rate ± GPM flow rate ± 0.15 kg/s PX2300-5DI pressure (0-5) ± psi PX DI pressure (0-25) ± psi Omega Flow & Level Handbook (2% full scale) Omega Flow & Level Handbook (3% full scale) Omega Pressure, Strain & Force Handbook (0.25% of full scale) Omega Pressure, Strain & Force Handbook (0.25% of full scale) 43

66 4. TEST RESULTS AND DISCUSSION 4.1 Water/Al 2 O 3 Results and Analysis Heat Transfer (W) Input Power 2900 Microchannels Process 2700 Chiller Process Flowrate (GPM) Figure 4.1: Water heat duty comparison. Heater input power is compared to microchannel heat exchanger, process, and cooling water calculated thermal energies. The first fluid to be tested in the flow loop apparatus was water. The properties of water are well known and it does not require intensive cleaning at the conclusion of a test, making it the ideal fluid to baseline the system. Figure 4.1 shows the heat balance obtained for water. All thermal energies balance well with respect to the heater input power calculated from the voltage and current. The remaining values were calculated 44

67 using Eqn 2.1. The energy balance for the lowest flowrate showed a large error as compared to data collected at subsequent flowrates. This may be explained by the fact that the system had not yet reached steady state. At that time, the 2hr start up time had not yet been employed. Another observation was the increasingly higher chiller heat transfer at higher flowrates. Viscous heating could occur in such a high shear gear pump, and would be more prominent at higher flowrates. This is evident in the data which shows a temperature rise across the pump. In this case, the fluid would have to be cooled below the desired microchannel inlet temperature so that the desired microchannel inlet temperature (from 1) was maintained. The large error bars on the chiller were due to the fact that there was a minimal temperature rise relative to the precision of the thermocouples. Overall, the heat balance was acceptable and gave confidence that the apparatus was reliable. Heat duty comparisons were plotted for all water-based nanofluids, and can be found in Appendix C. The water based nanofluid used for these tests was Nyacol, a commercially available product consisting of about 40nm Al 2 O 3 particles dispersed in water. As received, it was a 20%wt dilution of Al 2 O 3. Dilutions of 20%, 10%, 5% and 2.5% by weight were run for this thesis. The first test of 20%wt failed when a sudden spike in temperature was seen in one microchannel heat exchanger. It is thought that the fluid was moving so slowly that it began to boil and deposit nanoparticles on the surface of the heat exchanger. This caused further clogging and eventually led to nearly total blockage and thus the temperature spiked. Lee and Mudawar reported similar results in their flow boiling experiments with alumina/water nanofluids (2007). From this experience, three new protocols were developed. The first was that tests should be done in order of 45

68 increasing particle loading. The second was that the initial flowrate should be the highest, not lowest. That meant boiling at the surface was much less likely due to greater fluid momentum. Finally, a bypass section was placed around the rotameter to allow reverse flow through the system in the case of a clog. After this experiment, the microchannel heat exchangers were replaced and 2.5%, 5% and 10%wt dilutions were run in that order. The remainder of this section presents those results. Pressure drop is a critical measurement when considering nanofluids as drop-in replacement coolants. Pressure drop has the potential to offset any gains in heat transfer at equal Reynolds number because more pumping power is required to obtain equal Re. Figure 4.2 shows the system level pressure drop data from water and the three dilutions of Nyacol. It was clear that the presence of nanoparticles increased the pressure drop at equal flowrates. Further, the pressure drop increased as the particle loading increased. This was expected based on the prediction of Eqn 2.20 using fluid properties at 20 C. At 10%wt and 1.5GPM, the pressure drop was 1.65 times that of water. At 0.75GPM, the pressure drop for 10%wt was 1.79 times that of water. For laminar flow it is known that the increase in pressure drop is directly proportional the ratio of the viscosities, which was 1.80 for water and 10%wt Nyacol at 20 C. From Eqn 2.20, the turbulent flow pressure drop increase is a ratio of viscosities to the 0.3 power and the densities to the 0.7 power. For these conditions, that translated to a pressure drop increase of 1.24 over water. Since the system had a combination of laminar and turbulent flow, the pressure drop increase was expected to be in the range of 1.24 to Further, from Eqn 2.24, calculations of system pump power are directly proportional to the pressure drop and are 46

69 plotted in Appendix C. Subsequent heat transfer coefficients will be compared to the system pumping power System Pressure Drop (psi) % wt 5% wt 2.5% wt Water Flowrate (GPM) Figure 4.2: Flowrate comparison of system pressure drop for water and Al 2 O 3 Figure 4.3 shows the overall heat transfer coefficients as a function of the process flowrate for the various dilutions of alumina/water. The observed shape of the H curve was normal for a plate/frame heat exchanger, which typically have a power law fit with an exponent less than one. With few exceptions, the data show that the addition of nanoparticles did enhance the overall heat transfer coefficient at equal flowrates. From Eqn 2.29, it was expected that the oveall heat transfer coefficient would increase for each nanofluid, and also increase for higher particle loadings. The 2.5% fluid showed the greatest enhancement in overall H, and 5% showed the least increase (<1%) and in some cases showed up to a 6% decrease. It is unclear why this occurred as only a 47

70 minimal increase of 1% was expected for 2.5%, where as a 6% enhancement was predicted for the 10% dilution. It is possible that the increased viscosity effects of higher particle loadings had a larger than expected effect on the overall H as it was the only property that changed significantly with particle loading Overall Heat Transfer Coefficient (W/m 2 K) Water 2.5% wt 5% wt 10% wt Process Flowrate (GPM) Figure 4.3: Flowrate comparison of water and Al 2 O 3 Plate/Frame overall heat transfer coefficient 48

71 4000 Overall Heat Transfer Coefficient (W/m 2 K) Water 2.5% wt 5% wt 10% wt Pump Power (W) Figure 4.4: Pump power comparison of water and Al 2 O 3 Plate/Frame overall heat transfer coefficient A flowrate comparison does not tell the entire story of the heat transfer coefficient. Fig 4.4 shows the overall H as compared to the system pumping power. A significant amount of the enhancement seen in Fig 4.3 was cancelled out by the effects of increased viscosity and density on the pumping power required to achieve those flowrates. Fig 4.2 showed that for equal flowrates, there was a significant increase in pressure drop. At equal pumping power, Eqn 2.31 predicted that the H would decrease for all dilutions of Nyacol. Fig 4.4 showed that as the particle loading increased, the overall H decreased (because. An 11% decrease for the 10%wt dilution at a pumping power of 3W was predicted, which is similar to the observed decrease of 10%. There was a slight increase in measured H for the 2.5%wt nyacol, which was not predicted. It is possible that experimental error could be the reason for this as the error bars overlap the water data. In addition, this was a comparison 49

72 using the system pump power. It was recommended after the water tests that the individual heat exchanger pressure drops be measured for PAO and PAO NF to give a more accurate estimation for each of the components of the system pressure drop. The microchannel heat exchangers were used to evaluate the film H in laminar flow. Figure 4.5 indicates that for HX1 the film H decreased for each particle loading. For laminar flow, using Eqn 2.25 at equal volumetric flowrates, the relative increase in H was predicted to do the opposite of what is seen in Figure 4.5. At the temperatures seen in the heat exchanger, the film H for 2.5%, 5% and 10% should increase 5%, 6% and 7%, respectively. Figure 4.5 shows that the 5% sample performed better than the 2.5% sample. It was unclear why the measured values are so far removed from the predictions, and also why the order was not the same as predicted. Experimental error is always a factor, but cannot explain a more than 50% decrease in H for the 10% dilution and the fact that the shape of the H curve was much flatter than expected. For the 10%wt Nyacol, clogging could have led to higher than expected wall temperatures and would prevent the assumption of equal flow in each heat exchanger. Also, if the particles began to agglomerate then the thermal properties may differ from those measured and reported. However, properties measured after being run showed little change as compared to measurements taken prior to being run. It was difficult to conclude why 5% would outperform the 2.5% fluid and multiple runs with each fluid were not possible due to only small amounts of Nyacol on hand at the time of testing. After running subsequent fluids (PAO and PAO NF) many times, it was found the data sets can vary by up to 10% from run to run and enough runs should be done to average the data. Therefore, the potential exists that the true H for the 5% solution 50

73 was up to 10% higher than its measured value and the 2.5% solution is up to 10% lower than its measured value Heat Exchanger 1 Film H (W/m 2 K) Water % wt 5% wt % wt Microchannel Flowrate (GPM) Figure 4.5: Flowrate comparison of water and Al 2 O 3 heat exchanger 1 film H The Microchannel HX2 data were similar to the results from HX1 explained previously. However, there were some differences. Figure 4.6 displays the slope of the 10% is much more reasonable, although the film H values were still only slightly higher than 50% of water. The data for the 2.5% and 5% data were much closer together and in some cases showed the 5-6% increase that was predicted by Eqn The difference in the data for HX1 and HX2 seemed to indicate that clogging was an issue at higher concentrations. The pure water data were nearly identical for both heat exchangers, but film H values for the nanofluids in HX2 were 10-20% higher than the values from HX1 in almost every case. If agglomeration was affecting the fluid 51

74 properties, both heat exchangers should have been affected, as well as the plate/frame. Further inspection of the raw data for the 10%wt revealed that the wall temperatures for HX1 reached upwards of 150 C, which is more than twice the 65 C wall temperature of HX2. This was the only data that showed temperatures that high. It is reasonable to believe that an aqueous solution would boil at those temperatures and likely deposited alumina on the channel wall. With such small channels, minimal clogging in one heat exchanger could have caused a significant disparity in flowrates due to an increased pressure drop. Another explanation could be that two-phase gas/liquid flow would choke the flow in one or more of the microchannels. In either case, film heat transfer coefficients calculated under the assumption that the total flow was split evenly between the two heat exchangers were erroneous. The heat transfer coefficient data compared to the system pressure drop can be found in Appendix C. As in turbulent flow, the H data was worse when compared at equal pumping power/pressure drop. 52

75 4500 Heat Exchanger 2 Film HeH (W/m 2 K) Water % wt 500 5% wt 10% wt Microchannel Flowrate (GPM) Figure 4.6: Flowrate comparison of water and Al 2 O 3 heat exchanger 2 film H Experimental results for water and Nyacol were reasonably successful. The apparatus was able to evaluate the heat transfer coefficient and reflect any changes when different dilutions of nanoparticles were added. Conventional theory explained the increase in pumping power and overall H coefficient in the plate/frame, but did not consistently predict the performance of the microchannel heat exchangers. This was most likely due to clogging. If clogging was due to boiling, then it should not be an issue with PAO because of its high temperature capability. Another observation was that the nanofluids all displayed a severe gray-blue discoloration after being run even for minimal time. For this reason, only one experimental run was completed on each fluid. A potential cause could be a graphite bushing on the pump. From these first sets of 53

76 experiments, one major change was made to the apparatus. That was, the pressure data was collected for both the system and the individual heat exchangers to allow for a more detailed pump power comparison for each test section. Another recommendation for the next test was to perform multiple runs for each fluid. 4.2 PAO and PAO NF System Pressure Data Results and Analysis The PAO and PAO nanofluid (PAO NF) performance was compared to the system pressure and individual component pressure. Results from the system pressure comparison are detailed in this section. As with water, the heat balance for the system was evaluated for both fluids to ensure that the apparatus was operating correctly. The effective flowrates were calculated as outlined in the calibration section, using Eqn 3.3. Figures C.8 and C.9 show the heat duty comparisons for PAO and the PAO NF. From lessons learned with water, the data shown in all figures for PAO and the PAO NF are an average of at least five data runs. The charts do not display every data run in order to prevent clutter. The raw data from each individual run can be found in Appendix G. The PAO NF was prepared by METSS Corp. under a contract run by the U.S. government as part of the small business innovative research (SBIR) program. The nanofluid was made via the two-step process where dried nanoparticles were added to the basestock. The basestock was Synfluid 2cSt PAO, which was also the PAO run to compare to the nanofluid. A proprietary additive package was developed under the contract to improve particle dispersion and stability. Finally, 40nm Al 2 O 3 nanoparticles were added at 5%wt and the mixture was sonicated to disperse the particles. All SBIR data rights are reserved to the contractor and no data is published without their expressed 54

77 written consent. A letter from METSS Corp. is included in Appendix I granting permission to publish the results. The system pumping power as a function of flowrate is shown in Figure 4.7. It is directly proportional to the pressure drop which can be seen in Appendix C, Figure C.10. The pumping power required for the PAO NF was slightly less than PAO at high flowrates. There was no difference until about 1.4GPM, when the two began to diverge. From the reported viscosity data at 40 C, it was predicted from Eqn 2.24 that the power for PAO NF should be 96% of the PAO power. At high flowrates, was 0.95, which was very close to the predicted relative difference. The pressure drop results were much different than the water data, where the pressure drop increased for the nanofluids. The reason for this may be that the fluid was formulated by METSS Corp. to meet or exceed MIL-PRF which governs the properties of avionics coolants, and that the additive package used to maintain suspension, may act as a drag reducer in turbulent flow. 55

78 System Power (W) PAO Average PAO NF Average Process Flowrate (GPM) Figure 4.7: Flowrate comparison of system pumping power for PAO and PAO NF 1200 Overall Heat Transfer Coefficient (W/m 2 K) PAO Average PAO NF Average Process Flowrate (GPM) Figure 4.8: Flowrate comparison of the overall H for PAO and PAO NF 56

79 Figure 4.8 shows that for equal flowrates, the PAO NF slightly enhances the overall H as compared to PAO. However, this enhancement diminished as the flowrate increased. The relative increase in overall H was predicted by Eqn 2.29 and the appropriate coefficients from Table 2.2 for PAO. At 40 C and 0.75GPM, the overall H for PAO NF was expected to be 9% higher than PAO. However, the measured increase was 3%. A potential source of error was particle settling as seen in Figure 4.9. Particle settling was a large issue throughout the PAO NF data collection as large amounts of particles would settle out overnight. If enough particles settled out of the fluid and collected in the reservoir tank, it is possible that the actual thermal conductivity was lower than reported before the fluid was run. It has been shown previously that the thermal conductivity has the greatest effect on the H enhancement. Particle settling also would have lowered the density and increased the heat capacity, both of which would decrease the predicted enhancement. Figure 4.9: Photo of particles settling in the PFA tubing after sitting for one night The pump power required to obtain the overall H for PAO and PAO NF are shown in Figure The measured values showed a slight increase in overall H for 57

80 PAO NF. It was predicted using the exponents in Table 2.2, and property data Appendix G that the overall H will increase 7.4% and 9.3% for 1W and 8W, respectively. In reality, the measured H increase was 3.3% and 1.3% at 1W and 8W, respectively. The differences between the predicted and observed values were close and could be attributed to the difficulty in reading exact values from a chart or from estimations of system power at equal flowrates used to obtain the H. Another source of error was the property data. Nanofluild heat capacity measurements typically had a ±10% error on the reading. If the measured properties were incorrectly high, then the predicted H increase would be greater than the actual. Although the predicted and measured increases were not equal, they were close and both agree with an increased overall H. The same analysis was valid for the pressure drop data and can be seen in Figure C Overall Heat Transfer Coefficient (W/m 2 K) PAO Average PAO NF Average System Power (W) Figure 4.10: Pump power comparison of PAO and PAO NF plate/frame overall H 58

81 Figure 4.11 shows the film H for HX1 as compared to the flowrate. A slight decrease of less than 1% was observed for the nanofluid over the entire flow range. The relative increase was expected to be 1.1% for all flowrates. The differences were well within the calculated error and the experimental error was greater for the microchannel data than the plate/frame. Previous work with the microchannels showed that difference in flowrates could lead to error between the data of the microchannels. Therefore, it was not surprising that the data in Figure 4.12 for HX2 differed from the data for HX Heat Exchanger 1 Film H (W/m 2 K) PAO Average PAO NF Average Microchannel Flowrate (GPM) Figure 4.11: Flowrate comparison of heat exchanger 1 Film H Figure 4.12 shows that the HX2 film H increases 4-5% for the PAO NF and is consistent throughout the flow range. This was greater than the predicted increase of 1.1%. The primary cause of the observed difference could be the breakdown of the equal flowrate assumption for one or both of the fluids. It is doubtful that PAO or the PAO NF 59

82 would clog the heat exchanger because of their high boiling point and lubricating characteristics. Therefore, small differences in pressure drop through fittings could cause unequal flowrates. However, great care was taken in fabricating the test sections to ensure similar pipe lengths and fittings. The relative flowrate, is directly proportional to the relative pressure drop,, between the two heat exchangers. Earlier pressure drop experiments showed that each CP20 had a slightly different pressure drop. If the pressure drop in HX1 and HX2 was not identical, then the assumption of equally split flow is incorrect. If the pressure drop in HX2 was higher than HX1, then less than 50% of the flow was passing through. Thus, analysis using the split flow assumption will calculate a higher than actual H and vice versa for HX Heat Exchanger 2 Film H (W/m 2 K) PAO Average 650 PAO NF Average Microchannel Flowrate (GPM) Figure 4.12: Flowrate comparison of heat exchanger 2 Film H 60

83 The heat transfer is directly proportional to the flowrate for similar fluid. If one heat exchanger showed an increase in H and the other a decrease, it was reasonable to average them to estimate the overall enhancement. In this case, the average increase was 3-4%. Again, the expected increase was 1%. By inspecting the raw data, it became evident that something else was happening in the PAO data. For HX2, there were three PAO runs that showed substantially lower H values than the average data for PAO and four that matched the results from HX1. All of the below average runs were done on the same day, leading to some suspicion that the apparatus could have been performing poorly that day, i.e. the system may not have reached steady state. It is unclear why that would happen. The 2hr start up time procedure had not been discovered at that point which would have led to flawed data. However, those three runs were still done after the system had been running for greater than 2hrs as the first run on that day was excluded from data analysis. Eliminating all data from that day would cause the PAO to regain the slight advantage seen in HX1. Figure 4.13 and Figure 4.14 show the film H for microchannel HX1 and HX2 at equal system power. Given the nearly equal pumping power for the two fluids, it was expected that the results would be very similar to the flowrate comparison. Here the PAO NF decreased the film H by about 1.5% in HX1, whereas there was a 5% increase in HX2. The predicted increase was between 0.76% and 1.78%. If an average was to be taken between the two heat exchangers, the increase would be 3-4%. From this data, it remained unclear why there was such a large discrepancy in data between the two heat exchangers. However, there was not a significant difference in the average film H enhancement as compared to the predicted. 61

84 1200 Heat Exchanger 1 Film H (W/m 2 K) PAO Average PAO NF Average System Power (W) Figure 4.13: System power comparison of heat exchanger 1 Film H 1400 Heat Exchanger 2 Film H (W/m 2 K) PAO Average PAO NF Average System Power (W) Figure 4.14: System power comparison of heat exchanger 2 Film H 62

85 The tests done with PAO system pressure showed good agreement with theory and no major problems occurred during testing. Overall, PAO proved to be a much worse heat transfer fluid than water and was a more difficult fluid with which to work. However, it provided consistent results. It was during these tests that the 2hr start-up procedure was developed. Close data inspection revealed that the overall H for the plate/frame increased over time until reaching steady state after 2hrs of operation. This was a major breakthrough at the time and led to much more consistent results. 4.3 PAO and PAO NF Individual Pressure Data Results and Analysis Based on the results from the system pressure test data, it was recommended that the fluids should be compared using component pressure drops. Previously, it was difficult to estimate the actual effect of the laminar or turbulent flow on the overall pressure drop. Based on the Re number and heat exchanger type, it was assumed that for the microchannel heat exchangers, the flow was laminar and for the plate/frame the flow was turbulent. Therefore, it was expected that the measured component pressure drop will provide a better basis for comparing the H. The heat balances were again calculated using the effective flowrate method. No new reported trends were seen in the heat duty comparisons, which can be found in Appendix C, Figures C.12 and C.13. The individual pressure drop was then evaluated. A plot of the total pressure drops can be found in Appendix C Figure C.14, along with a plot of total system power as a function of flowrate Figure C.15. The total system power was calculated by adding the two component pressure drops. Nothing new was found in this data except that the total calculated system pressure drop was lower, presumably because some amount of 63

86 pressure drop from the fittings was missing. Figure 4.15 and Figure 4.16 are bar charts of the individual component pressure drops. The pipe/fittings pressure drop was calculated using the data from the system pressure and subtracting the laminar (microchannel) and turbulent (plate/frame) pressure drop data. The turbulent component accounts for 50% of both PAO and PAO NF pressure drop. As in the system pressure drop, there was a slight decrease in the pressure drop with PAO NF that could be attributed to a slight decrease in viscosity. The data labels illustrate the measured pressure drop values for each component. Comparing the two fluids, the data from the microchannel heat exchangers was very similar. However, the fittings and plate/frame pressures were higher for PAO. Of the two, the fittings decreased more in numerical value and also had a larger percent decrease. 12 Pressure Drop (psi) ΔP Fittings (psi) 2.41 ΔP MC (psi) ΔP (psi) Process Flowrate (GPM) Figure 4.15: Component pressure drop for PAO 64

87 12 Pressure Drop (psi) ΔP Fittings (psi) ΔP MC (psi) 1.81 ΔP (psi) Process Flowrate (GPM) Figure 4.16: Component pressure drop for PAO NF The benefit of measuring individual pressure drop was that the overall heat transfer could be compared to the actual pressure drop and pumping power required for each test section. Previously, the overall H was compared to the pump power. Because pump power and pressure drop are directly proportional to one another, the pressure drop comparison will be shown here. Since there were no new conclusions to report regarding the pump power or flowrate comparisons, they can be found in Appendix C. Figure 4.17 shows the pressure drop comparison of the overall H. The overall H was consistently 4-5% higher for PAO NF. This is similar to the predicted gain of 5-6%. However, there was a greater increase than when compared to the system power results from Figure 4.10 where there was a 0-3% enhancement. The trend of a constant increase was predicted, only at a slightly higher percentage than what was 65

88 measured. One explanation for the greater increase could be the aforementioned particle settling. The individual pressure runs for PAO NF were carried out nearly 2 weeks prior to switching over to system pressure runs. Large amounts of particle settling were noticed in clear tubing part way into the individual pressure experiments. If particles had been settling in various locations in the loop, mainly the reservoir, it could result in the decrease in measured overall H for the system pressure drop. If this is true, it does not help the cause this nanofluid as a drop-in replacement for PAO. It could also mean that if a more stable nanofluid can be made with similar properties, then the increase of 4-5% would be substantial enough to warrant long term testing Overall Heat Transfer Coefficient (W/m 2 K) PAO Avg PAO NF Avg Plate/ Frame Pressure Drop (psi) Figure 4.17: Plate/Frame pressure drop comparison of the overall H for PAO and PAO NF 66

89 The microchannel HX1 film H data as a function of flowrate is shown in Figure It shows the same trend as previously with a system pressure measurement, only PAO is 4% higher than the PAO NF rather than only 1-2%. At equal flow, the predicted increase in film H for PAO NF was 1.1%. The discrepancy could again be due to small differences in flowrates. Figure 4.19 shows the same data for microchannel HX2. Again, a 4-5% increase is seen in the film H. The data agrees with the system pressure data from before, but is still much higher than predicted. The average film H increase for both microchannel heat exchangers is about 1%, which almost exactly the predicted value. Therefore, the predicted increase was valid for this data Heat Exchanger 1 Film H (W/m 2 K) PAO Avg Microchannel Flowrate (GPM) Figure 4.18: Flowrate comparison of HX1 film H for PAO and PAO NF 67

90 Heat Exchanger 2 Film H (W/m 2 K) PAO Avg PAO NF Avg Microchannel Flowrate (GPM) Figure 4.19: Flowrate comparison of HX2 film H for PAO and PAO NF The pressure drop data is reported here, with the pump power comparison found in Appendix C. Figure 4.20 and Figure 4.21 show the measured microchannel pressure drop comparison of the film H for HX1 and HX2, respectively. At equal pressure drop, the maximum expected increase was 1.2%. Due to very similar pressure drops, there were no significant changes to the results found at equal flowrates. At 1.3psi, there was a 3% decrease in performance for HX1 and a 4.2% enhancement for HX2. At 3.3psi, there was a 3.5% decrease for HX1 and a 3% increase for HX2. The predicted enhancement for the PAO NF at 3.3 psi was 1.3%. Averaging the values indicated that there was never more than a 1.2% enhancement in the film H. The error associated with these values was high due mainly to the large uncertainty in the heat capacity values. The individual pressure data gave no further insight into why there was a large 68

91 discrepancy in the calculated film H between the two heat exchangers. However, when an average film H increase was taken, it matched the predicted increase Heat Exchanger 1 Film H (W/m 2 K) PAO Avg PAO NF Avg Pressure Drop (psi) Figure 4.20: Microchannel pressure drop comparison of HX1 film H for PAO and PAO NF 69

92 1400 Heat Exchanger 2 Film H (W/m 2 K) PAO Avg PAO NF Avg Pressure Drop (psi) Figure 4.21: Microchannel pressure drop comparison of HX2 film H for PAO and PAO NF 70

93 5. CONCLUSIONS AND RECOMMENDATIONS A coolant loop apparatus was designed and built. Experiments were successfully completed to determine the convective heat transfer coefficient for water, Al 2 O 3 /water, PAO and a Al 2 O 3 /PAO nanofluid. Relative differences in the H were compared at equal flowrates and equal pumping power. No anomalously high results were observed as compared to the predicted performance from conventional theory. Overall, most results showed that the nanofluids provided a slight increase (<5%) in convective H over their base fluids. There was little difference between the laminar and turbulent results. Other issues such as fluid discoloration with the Al 2 O 3 /water solution and particle settling with the PAO NF suggest that these nanofluids are a questionable choice for a drop-in replacement for PAO. 5.1 Conclusions The following conclusions are made based on the results and analysis: Based on the water baseline test results as compared to values predicted by conventional theory, the coolant loop apparatus was able to reasonably measure the convective heat transfer coefficient. 71

94 There was no significant (i.e., order of magnitude) differences in H s measured by the coolant loop and those predicted by conventional theory. As expected, the pressure drop was significantly increased by the large increase in viscosity for the Al 2 O 3 /water nanofluids, which was found to increase with particle loading. Consequently, at equal pump power, the measured H for Al 2 O 3 /water was worse than the equal flowrate comparison. At each particle loading, there was severe gray-blue discoloration of the Al 2 O 3 /water fluids. However, the property data before and after showed no significant changes. The PAO NF showed a slight decrease in measured viscosity and pumping power as compared to PAO. Therefore the results at equal flowrates and equal pumping power were very similar. There was significant particle settling for the PAO NF after running in the loop. This could be due to agglomeration after long amounts of time in use or also due to the heating and cooling cycle or the high shear conditions in the gerotor pump. The coolant loop apparatus is useful for determining the convective H. The inherent system error will not allow the user to determine very small differences in H. However, minimal increases in H will not satisfy the need for increased heat transfer. In general, it is thought that a >5% increase in H over PAO is necessary to warrant further investigation of a fluid as a 72

95 drop-in replacement. The coolant loop can reliably distinguish those differences for pure fluids and stable nanofluids. The coolant loop apparatus is a valuable tool for determining system-level effects of nanofluids is a flowing system. For instance, effects of particle settling, fluid discoloration, part wear, and cleaning that may not otherwise be observed in static laboratory experiments can be evaluated in the coolant loop. The goal of this thesis was met to design and build an apparatus that could be used to screen candidate drop-in replacement heat transfer fluids for avionics cooling. 5.2 Recommendations As designed, the coolant loop apparatus showed repeatable H results for pure fluids and stable nanofluids. However, through the course of this work, it is evident that the following changes would further improve the coolant loop operation: By using two Lytron CP20 heat exchangers in parallel, it was evident that there were small differences in flowrate that affected the measured film H results. A laminar flow heat exchanger should be designed to allow thermally and hydraulically fully developed flow at 2GPM. It should be designed with temperature, pressure and flow instrumentation in mind. The rotameter used for these tests was significantly affected by the fluid viscosity. Therefore, a turbine flowmeter that is suitable for various viscosities and densities should be installed so that the flowrate needs no correction. 73

96 It is clear that the panel meters used to collect experimental data were greatly affected when used in stacked operation. They should be separated so that there is sufficient airflow around the meters to allow for an appropriate junction temperature. For this work, manual operation was the best choice. Future experiments should include a data acquisition system to accelerate testing and data analysis. The 2 hour start up time may be fixed by adding one plate to the HX and allowing chiller water to flow through it. This will create a total of four cold streams and three hot streams. The hot streams will be completely surrounded by cold streams, thus the hot stream will no longer be in contact with the end plate and adiabatic conditions can be assumed. This should reduce the amount of start-up time to reach steady state. Only stable nanofluid solutions should be considered for use in the coolant loop. Without a stable solution, the properties of the nanofluid could change over time, leading to reduced H data. 74

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104 APPENDIX A Photos of Flow Loop Components Figure A.1: Haight 6US GPM gerotor pump drawings 82

105 Figure A.2: Haight gerotor pump and motor assembly Figure A.3: Leeson speedmaster variable DC motor control for pump. Control includes a power switch, brake switch and a switch to reverse flow. 83

106 Figure A.4: Lytron CP20 Microchannel Cold Plate Heat Exchanger was the smallest of its kind commercially available at the time of purchase. There are 42 flow channels H x 0.03 W x 2.0 L. The heat transfer surface area is 4in 2 per side. Figure A.5: Cartridge heater control board from Proheat, Inc. Includes a variac (center) to set heater power setting, a voltmeter (top left) and ammeter (top right). The four switches control which heater block are operating at any time. A temperature alarm (bottom center) displays the hottest block temperature and shuts power off if temperature goes over the alarm setting. 84

107 Figure A.6: Dimensions used for machining cartridge heaters hole in a 2 x2 x2 copper block Figure A.7: Finite element analysis of 300W cartridge heaters in a copper block 85

108 Figure A.8: Photo of heating test section Figure A.9: Wiring diagram for omega PX2300 series differential pressure transducers 86

109 Figure A.10: SWEP Minex M10 chevron plates with gaskets were arranged in alternating fashion 87

110 Figure A.11: SWEP Minex M10 manufacturer drawing with front, side and cutaway views 88

111 89

112 Figure A.12: Manufacturer instructions for assembling Minex M10 gasketed plate/frame heat exchanger 90

113 Figure A.13: Lytron Kodiak chiller control panel 91

114 Figure A.14: Lytron Kodiak chiller rear panel components Figure A.15: Omega FL4505-V rotameter 92

115 Figure A.16: Reservoir tank fitting locations on 6 OD Stainless Steel flange Figure A.17: SWEP Minex M10 Plate/Frame heat exchanger compression plates. 93

116 Figure A.18: Detailed design drawing of the coolant loop apparatus showing the main components, pipe dimensions and pressure and temperature port locations. The dashed line represents the breadboard top on which the loop was constructed. All other pipes/components were mounted vertically below the breadboard where the pump is the lowest component. 94

117 APPENDIX B Operating Procedures Note: All test procedures refer to the labels from Fig 3.1 I. SAFETY 1. Prior to running any fluid other than water, check MSDS for: a. How to clean up spills b. What to do if fluid comes into contact with eyes, skin, ingested, etc. c. Material compatibility with all gaskets, tubing, and heat exchanger materials 2. Ensure proper ventilation (if necessary) 3. Always ensure proper PPE is used (goggles, gloves, mask, etc) II. FILLING PROCEDURE 1. Close V4,5,6 & 7 2. Tie fluid level indicator tube above top of the reservoir to prevent spills while filling 3. Pour fluid into reservoir 4. Open V5, 6 to allow fluid to gravity fill the loop 5. Turn on pump and slowly increase power until fluid begins to circulate 6. Close V3 on rotameter by-pass section 7. Continue to increase pump control to 75%, checking for leaks fix leaks as necessary. This may require stopping the pump. 95

118 8. Once no more air is bubbling out of the reservoir, reduce pump power to 0% and turn off 9. Replace reservoir cover and re-connect the fluid level indicator and pressure relief lines 10. Loop is ready for operation III. PRESSURE TRANSDUCER BLEEDING This procedure can be done at any time, but is required any time the loop is filled. 1. With pump running at 75%, place a cup under the transducer 2. Close V5 about half-way so that the line pressure increases 3. Loosen the screw on the high pressure side, being careful not to remove screw completely. Keep open until all air bubbles are removed from the clear lines. 4. Repeat step 3 for the low pressure side, starting with the bottom screw and then the top screw 5. Repeat steps 1-4 for all pressure transducers 6. Fully open V5 7. Loop is ready for operation IV. START-UP 1. Ensure all breakers on loop are in the off position 2. Ensure that pump and heater controls are in the off position 3. Turn breakers on 4. Ensure that V1, 2 on chiller are open and turn chiller on and set to desired temperature 96

119 *Never operate the heaters without first having the pump and cooling water turned on 5. Open V3, 5, 6; Close V4, 7 6. Turn on pump and slowly increase to desired flow checking for leaks 7. Close V3 and check for leaks again 8. Check that all instrumentation is functioning properly 9. Set heater power to 0% and turn on all 4 heaters 10. Increase heater control to desired power setting 11. Allow apparatus to run for 2 hours to allow loop to reach steady conditions 12. Adjust the chiller flowrate as necessary using V2 to gain desired inlet temperature (1) - Closing the valve will increase 1 - Opening the valve will decrease Apparatus is ready to begin data collection V. DRAINING PROCEDURE 1. Attach tygon tubing to valve V7 and place in bucket or drain (if using water) 2. Loosen the top cover of the reservoir tank 3. Open V7 and allow fluid to drain due to gravity 4. Ensure pump and heater controls are in the off position and plug in the apparatus 5. Turn on the breakers 6. Open V3 rotameter by-pass valve 7. Switch the pump direction to reverse on the pump control panel 97

120 * Never run the pump in reverse without the rotameter by-pass valve open as it could cause the pump to cavitate 8. Turn the pump on and slowly increase power to about 75% on the DC control and allow pump to drain the remaining fluid 9. Increase the pump to 100% for about 30s or until no more liquid comes out 10. Reduce pump control to 0% and turn off 11. The loop is now drained and ready to be broken down for cleaning or filled with the next fluid VI. CLEANING **Note that whenever using nanofluids, always wear eye protection and latex gloves** 1. Break loop down into small sections and lay neatly on table lined with absorbent cloth 2. Individually rag clean each component with solvent, i.e. hexane, acetone, etc. 3. Take apart and clean heat exchanger, pump, and rotameter 4. Rebuild the loop just as it was 98

121 APPENDIX C Additional Data Plots Heat Transfer (W) Input Power Microchannels Process Chiller Process Flowrate (GPM) Figure C.1: 2.5%wt Alumina/water heat duty comparison. Heater input power is compared to microchannel heat exchanger, process, and cooling water calculated thermal energies 99

122 Heat Transfer (W) Process Flowrate (GPM) Figure C.2: 5% wt Alumina/water heat duty comparison. Heater input power is compared to microchannel heat exchanger, process, and cooling water calculated thermal energies Input Power Microchannels Process Chiller Heat Transfer (W) Input Power Microchannels Process Chiller Process Flowrate (GPM) Figure C.3: 10%wt Alumina/water heat duty comparison. Heater input power is compared to microchannel heat exchanger, process, and cooling water calculated thermal energies. 100

123 12 10 Pump Power (W) % wt 5% wt 2.5% wt Water Flowrate (GPM) Figure C.4: Flowrate comparison of pump power required for water and Al 2 O Overall Heat Transfer Coefficient (W/m 2 K) System Pressure (psi) Water 2.5% wt 5% wt 10% wt Figure C.5: Pressure drop comparison of water and Al 2 O 3 Plate/Frame heat transfer coefficient 101

124 Heat Exchanger 1 Film H (W/m 2 K) Water 2.5% wt 5% wt 10% wt System Pressure Drop (psi) Figure C.6: Pressure drop comparison of water and Al 2 O 3 microchannel heat exchanger 1 film H Heat Exchanger 2 Film H (W/m 2 K) Water % wt 500 5% wt 10% wt System Pressure Drop (psi) Figure C.7: Pressure drop comparison of water and Al 2 O 3 microchannel heat exchanger 2 film H 102

125 Heat Transfer (W) Input Power Microchannels Process Chiller Process Flowrate (GPM) Figure C.8: PAO heat duty comparison for system pressure. Heater input power is compared to microchannel heat exchanger, process, and cooling water calculated thermal energies. Heat Transfer (W) Process Flowrate (GPM) Figure C.9: PAO NF heat duty comparison for system pressure. Heater input power is compared to microchannel heat exchanger, process, and cooling water calculated thermal energies. 103 Input Power Microchannels Process Chiller

126 12 10 System Pressure Drop (psi) PAO Average PAO NF Average Process Flowrate (GPM) Figure C.10: System pressure drop comparison for PAO and PAO NF 1200 Overall Heat Transfer Coefficient (W/m 2 K) PAO Average PAO NF Average System Pressure Drop (psi) Figure C.11: System pressure drop comparison of plate/frame overall H for PAO and PAO NF 104

127 Heat Transfer (W) Process Flowrate (GPM) Figure C.12: PAO heat duty comparison for individual pressure. Heater input power is compared to microchannel heat exchanger, process, and cooling water calculated thermal energies. Heat Transfer (W) Input Power Microchannels Process Chiller Input Power Microchannels Process Chiller Process Flowrate (GPM) Figure C.13: PAO NF heat duty comparison for individual pressure. Heater input power is compared to microchannel heat exchanger, process, and cooling water 105

128 calculated thermal energies. System Pressure Drop (psi) Process Flowrate (GPM) Figure C.14: Total pressure for PAO and PAO NF calculated from individual pressure drop 7 PAO Avg PAO NF Avg 6 5 System Power (W) Process Flowrate (GPM) Figure C.15: Total pump power for PAO and PAO NF calculated from individual pressure drops 106 PAO Avg PAO NF Avg

129 1200 Overall Heat Transfer Coefficient (W/m 2 K) PAO Avg 500 PAO NF Avg Process Flowrate (GPM) Figure C.16: Flowrate comparison of plate/frame overall H 1200 Overall Heat Transfer Coefficient (W/m 2 K) PAO Avg PAO NF Avg Pump Power (W) Figure C.17: Pump power comparison of plate/frame overall H 107

130 APPENDIX D Flowmeter Calibration Procedure and Results Flowmeter calibrations should performed on all fluids when there is sufficient volume of the fluid to do so. The bucket/stopwatch technique is the most accurate and repeatable way to calibrate a flowmeter. The bucket and stopwatch technique is detailed below: 1. Configure two buckets as shown below in Figure E.1 with the fluid inlet coming from the pump or an external source (if calibrating water) and flowing through the rotameter. Figure D.1: Setup for bucket/stopwatch calibration technique 2. Weigh empty bucket 2 3. Begin flow and set to desired flowrate as indicated on flowmeter 4. When ready, transfer fluid outlet tube to bucket 2 and begin timer 5. Allow fluid to fill bucket 2 for at least 60sec. Be sure tube outlet remains above liquid level so that the pressure drop stays constant 6. Transfer outlet tube back to bucket 1 and at the same time, stop the timer 108

131 7. Stop flow 8. Weight bucket 2 again 9. Determine weight of fluid 10. Record time and weight of fluid 11. Repeat steps 2-8 until a minimum of 5 measurements are taken for each flowrate. Usually 4-5 flowrates should be evaluated as in Figure D Use Eqn D.1 to calculate the measured flowrate Eqn D Make a plot of measured flow vs indicated flow and compare to as in Figure D Measured Flow (GPM) y = x Average Indicated Flow (GPM) Figure D.2: FL902-G rotameter calibration for water shows that the flowmeter is calibrated correctly 109

132 1.60 Measured Flow (GPM) y = x Measured Indicated Flow (GPM) Figure D.3: FL902-G rotameter calibration for PAO shows that the measured flow is much lower than the indicated flow. 2.0 Measured/Predicted Flow (GPM) y = x Measured Indicated Flow (GPM) Figure D.4: FL902-G rotameter calibration for PAO NF shows that the measured flow is much lower than the indicated flow. 110

133 Measured Flow (GPM) SG = 1 SG = SG = SG = SG = SG = y = x Indicated Flow (GPM) Figure D.5: FL902-G rotameter calibration for EG/Water mixtures with higher viscosities and densities shows that increases in viscosity significantly reduce the measured flowrate as compares to the observed flowrate. 6 Measured Flow (GPM) y = x Measured Indicated Flow (GPM) Figure D.6: FL4504-V rotameter calibration for water shows that the flowmeter is calibrated correctly 111

134 APPENDIX E Heater Control Calibration Calculated Input Power (W) Indicated Measured Variac Setting (%) Figure E.1: Indicated and Measured values for heating power. The power was calculated using P=IV, and clearly shows that the voltmeter and ammeter readings are accurate as compared to the value measured with a Fluke 182 True RMS Miltimeter calibrated through AF PMEL. 112

135 APPENDIX F Thermocouple Calibration Data and Results It was found during data analysis that the stacked panel meter operation (Figure F.1) was causing the top panel meter to display erroneously high temperatures. It is believed that heat from the bottom panel meter caused the junction temperature on the top meter to raise, resulting in erroneously high readings. The error was further amplified by a lower than normal reading on the Outlet thermocouple. Therefore a calibration was done for both stacked and separated (Figure F.2) operation to determine the calibration equations to use for each thermocouple reading. Thermocouples 1-4 from Figure 3.1 were also evaluated and corrections developed. The thermocouples were calibrated by placing them in a constant temperature recirculating bath along with a high accuracy T-type thermocouple and an Omega HH509 Probe as a reference (Figure F.3). Five temperatures over the normal range of operation were evaluated. This calibration allows the temperature difference from the reference to be determined for stacked and separate operation for each temperature. The experiment clearly showed that the temperatures displayed on the top meter ( Inlet, Supply and Return) were much higher when they were stacked than when they were separated (Figure F.4). Therefore, curve fit calibrations for stacked operation were developed and used to correct the temperature readings. The calibration constants developed in this experiment are shown in Table F.1. The constants are meant to fit Eqn F.1 Eqn F.1 where is the calibrated temperature and is the indicated temperature from the raw data found in Appendix H. In addition, Table F.1 shows the previously developed calibrations for 1-1 thru 4-1, which measure the wall temperatures. They were done after the loop had been sitting overnight so that the temperature was the same throughout the system. They are simply an offset from the actual temperature so that and. 113

136 Figure F.1: Photo of panel meters stacked. This caused a significant error to the thermocouple reading due to the heat transfer from the bottom meter to the top meter Figure F.2: Photo of panel meters separate. This test verified the hypothesis that stacking the panel meters was causing an erroneous reading 114

137 Figure F.3: Photo of the test setup to calibrate the process and chiller fluid thermocouples. Results showed that when the panel meters were stacked, the indicated temperature was significantly higher than when the panels were separate T From Reference () Inlet Stacked Separate Reference Temperature () Figure F.4: Plot of stacked and separate temperature difference. The stacked reading is consistently 1 higher than for separate operation 115