Applied Science, Faculty of
Materials Engineering, Department of
DSpace
UBCV
Chester, Noel Leslie
Master of Applied Science - MASc
University of British Columbia
Controlled cooling on the runout table is a crucial component in the production of advanced high strength steels since there is a direct correlation between the cooling path experienced by the steel and its final mechanical properties. Recent trends are towards enhanced cooling of the steel so that complex dual phase microstructures can be obtained. Cooling of the steel on the runout table is usually achieved via impinging water jets on both the top and bottom of the strip, which can lead to significant heat transfer enhancement. The purpose of this study is to develop a fundamental understanding of bottom jet impingement cooling of hot steel plates. Specifically, this research has focused on the role bottom jet nozzle inclination angle and water flow rate have on the effectiveness of heat extraction on a hot steel plate. The methodology undertaken for this research involved carrying out experiments using a pilot scale runout table with both stationary and moving plates. Volumetric flow rates, strip speeds, and inclination angles were in the range of 35- 55 l/min, 0-1 m/s, and 10-30°, respectively. Experimentally, each plate was instrumented with numerous sub-surface thermocouples installed approximately 1 mm from the impingement surface. Temperature measurements were taken at the impingement point and several streamwise distances from the impingement point in all directions. Using the above measurements in conjunction with an Inverse Heat Conduction (IHC) model allowed the calculation of boiling curves or heat fluxes as a function of plate surface temperature and time. From the above measurements, transient cooling data on the hot steel plate by bottom jet impingement has been analysed.
A STUDY OF BOILING HEAT TRANSFER ON A HOT STEEL PLATE COOLED BY AN INCLINED CIRCULAR BOTTOM WATER JET By Noel Leslie Chester B.Sc. (Mechanical Engineering), University of Manitoba, 2004 A THESIS SUBMITTED IN PARTIAL FULFILLMENT OF THE REQUIREMENTS FOR THE DEGREE OF MASTER OF APPLIED SCIENCE in THE FACULTY OF GRADUATE STUDIES (MATERIALS ENGINEERING) THE UNIVERSITY OF BRITISH COLUMBIA December 2006 © Noel Leslie Chester, 2006 Abstract ii Abstract Controlled cooling on the runout table is a crucial component in the production of advanced high strength steels since there is a direct correlation between the cooling path experienced by the steel and its final mechanical properties. Recent trends are towards enhanced cooling of the steel so that complex dual phase microstructures can be obtained. Cooling of the steel on the runout table is usually achieved via impinging water jets on both the top and bottom of the strip, which can lead to significant heat transfer enhancement. The purpose of this study is to develop a fundamental understanding of bottom jet impingement cooling of hot steel plates. Specifically, this research has focused on the role bottom jet nozzle inclination angle and water flow rate have on the effectiveness of heat extraction on a hot steel plate. The methodology undertaken for this research involved carrying out experiments using a pilot scale runout table with both stationary and moving plates. Volumetric flow rates, strip speeds, and inclination angles were in the range of 35-55 l/min, 0-1 m/s, and 10-30°, respectively. Experimentally, each plate was instrumented with numerous sub-surface thermocouples installed approximately 1 mm from the impingement surface. Temperature measurements were taken at the impingement point and several streamwise distances from the impingement point in all directions. Using the above measurements in conjunction with an Inverse Heat Conduction (IHC) model allowed the calculation of boiling curves or heat fluxes as a function of plate surface temperature and time. From the above measurements, transient cooling data on the hot steel plate by bottom jet impingement has been analysed. Table of Contents iii Table of Contents Abstract ii Table of Contents iii List of Tables vi List of Figures vii Nomenclature xi Acknowledgements xiv Dedication xv Chapter 1 - Introduction 1 Chapter 2 - Literature Review 2 2.1 Industrial Runout Table Layout 3 2.2 Hydrodynamics of Jet Impingement 5 2.2.1 Hydrodynamic Jet Configurations 5 2.2.2 Hydrodynamic Regions 7 2.3 Boiling Heat Transfer Mechanisms 11 2.3.1 Pool Boiling 11 2.3.2 Jet Impingement Boiling Regions 13 2.3.3 Wetting Front Analysis 16 2.4 Investigative Methods for Jet Impingement Experiments 19 2.4.1 Steady-State Method 20 2.4.2 Transient Cooling Method 23 2.4.3 Review of Bottom Jet Experiments 25 2.5 Effect of Inclination 28 Table of Contents iv 2.6 Effect of Strip Speed 31 Chapter 3 - Procedures and Methodology 34 3.1.1 ROT facility 34 3.1.2 Test Material 36 3.1.3 Thermocouple Instrumentation and Carrier Design 37 3.1.4 Data Acquisition 42 3.1.5 Test Procedure 42 3.2 Image analysis 45 3.3 Numerical Analysis 47 3.3.1 IHCP Model 47 3.3.2 Wetting Front Propagation 53 Chapter 4 - Results and Discussion 56 4.1 Typical Test Data 57 4.1.1 Cooling Curves 57 4.1.2 Boiling Curves 60 4.1.3 Visual Observations 67 4.1.4 The Wetting Front 71 4.1.5 Quantification of Heat Extraction 79 4.2 Effect of Inclination Angle 86 4.2.1 Boiling Curves 86 4.2.2 Overall Energy Dissipation 89 4.3 Effect of Water Flow Rate 91 4.3.1 Boiling Curves 91 Table of Contents v 4.3.2 Overall Energy Dissipation 93 4.4 Effect of Sample Movement 94 Chapter 5 - Conclusions and Future Recommendations 100 5.1 Summary 100 5.2 Conclusions 101 5.3 Recommendations for Future Work 104 References 105 Appendix A - Sensitivity Analysis 113 A-l Mesh Sensitivity 113 A-2 Conductivity Sensitivity 115 Appendix B - Validation of ED Technique 118 Appendix C - Validating Interpolation Technique with Off-Axis Locations 120 C-l Comparison of Wetting Front Times in Elliptical Model 120 C-2 Comparison of Surface Temperatures in Elliptical Model 121 List of Tab les vi List of Tables Table 3-1: Chemical composition of test material in wt% 37 Table 3-2: Thermo-physical properties of the plate material 52 Table 3-3: Measurement errors 55 Table 4-1: Experimental matrix for stationary plate 56 Table 4-2: Experimental matrix for moving plate 56 Table 4-3: Area of initial wetted region (mm ) as a function of inclination angle and water flow rate 77 List of Figures vii List of Figures Figure 2-1: Schematic of a runout table (modified from Liu, 2002) 3 Figure 2-2: Three cooling systems 4 Figure 2-3: Types of jet configuration (modified from Wolf et al., 1993) 6 Figure 2-4: Pressure and velocity profile, and flow regions for a jet with a uniform velocity profile (modified from Wolf et al., 1993) 8 Figure 2-5: Hydrodynamic boundary layer and velocity profile of a planar top jet impinging on a (a) stationary plate, and a (b) moving plate (modified from Zumbrunnen et al., 1992) 9 Figure 2-6: Hydrodynamic boundary layer and velocity profile of a planar bottom jet impinging (a) vertically and (b) obliquely on a stationary plate 10 Figure 2-7: Typical pool boiling curve 11 Figure 2-8: Different boiling regions present during top jet impingement (modified from Zumbrunnen et al., 1989) 15 Figure 2-9: Different boiling regions present during bottom jet impingement 15 Figure 3-1: Schematic of the bottom jet cooling system 36 Figure 3-2: Diagram of an internal thermocouple connected by an intrinsic junction 38 Figure 3-3: Diagram of a stationary plate thermocouple configuration and designation of different flow directions 38 Figure 3-4: Diagram of a moving plate thermocouple configuration and designation of strip direction 39 Figure 3-5: Diagram of flow directions : '. 40 Figure 3-6: Schematic of the upper carrier for stationary plate 41 Figure 3-7: Schematic of the upper carrier for moving plate 42 Figure 3-8: Images taken from stationary experiment #5 at 2s after jet impingement (i.e. water temperature = 30°C, flow rate = 45 1/min, and inclination angle = 20°) 46 Figure 3-8: Schematic of IHC boundary conditions in a) Forward and Reverse flow direction, and b) Lateral flow direction 49 Figure 3-9: Wetting front progression taken from stationary experiment #4 (i.e. flow rate = 35 1/min, and inclination angle = 20°) in the forward flow direction 54 List of Figures viii Figure 4-1: Measured thermal history for positions away from the centre of the jet for experiment #4 (35 l/min flow rate, and 20° inclination) showing: a) Forward flow direction, b) Reverse flow direction, and c) Lateral flow direction 59 Figure 4-2: Calculated heat flux versus time in the forward flow direction for experiment #9 (55 l/min flow rate and 30° inclination) 61 Figure 4-3: Calculated boiling curves for stationary experiment #9 (55 l/min flow rate and 30° inclination) showing: a) Forward flow direction, b) Reverse flow direction and c) Lateral flow direction 64 Figure 4-4: Boiling curve outside impingement zone at +32mm from jet centreline in the lateral flow direction from stationary experiment #9 (45 l/min and 20°) 65 Figure 4-5: Video images taken during experiment # 5 (45 l/min and 20° inclination) for: a) 0.0333s after jet impingement, b) 0.267s after jet impingement and c) 1.333s after jet impingement 69 Figure 4-6: Effect of inclination angle on the symmetry of the surface temperature profile at 2s after turning on the jet at 45 l/min and a) 0° inclination, b) 10° inclination, c) 20° inclination, and d) 30° inclination 70 Figure 4-7: Comparison of wetting front estimates between 2nd order derivative method and video images from stationary experiment #4 (35 l/min and 20°) in the lateral flow direction 71 Figure 4-8: Effect of inclination angle and flow rate on the wetting front progression for: a) Forward flow direction, b) Lateral flow direction and c) Reverse flow direction....74 Figure 4-9: 2D wetting front progression over time taken from stationary experiments at 45 l/min flow rate and showing a) 0°, b) 10°, c) 20° and d) 30°. 76 Figure 4-10: Size of wetted zone with respect to time for stationary experiments with 451/min flow rate 78 Figure 4-11: Schematic showing heat flux curves overlaid on nodal grid based on wetting front contours in forward flow section. Data from stationary experiment #4 (35 l/min, 20° inclination) 80 Figure 4-12: Schematic indicating technique used to estimate the heat flux (QDP) at a discrete nodal point between two heat flux curves at a given time, t 81 List of Figures j x Figure 4-13: Heat flux contours [W/m ] taken in forward flow section at 2, 5, and 7.5s into stationary experiment #4 (35 1/min flow rate and 20° inclination angle) 82 Figure 4-14: Schematic showing heat flux contours with position of wetting front at 20s into cooling in forward flow section. Data from stationary experiment #4 (35 1/min, 20° inclination) 83 Figure 4-15: Weighted average heat fluxes with wetting front time from stationary experiment#4 (35 1/min, 20° inclination) 83 Figure 4-16: Characteristic spatial distribution of heat fluxes on the wetting front 84 Figure 4-17: Variation in overall calculated heat flux across an area of the plate as a function of time. Data illustrated was taken from stationary experiment #3 (55 1/min, 10° inclination) 86 Figure 4-18: Effect of inclination angle in forward flow direction showing a) initial wetted zone (16mm from the centre) and b) outside initial wetted zone (64mm from the centre) 87 Figure 4-19: Effect of inclination angle in the lateral flow direction showing a) within the initial wetted zone (16mm from the centre) and b) outside the initial wetted zone (32mm from the centre) 89 Figure 4-20: Overall energy dissipation as a function of inclination angle and water flow rate 90 Figure 4-21: Effect of flow rate for a jet with a 10° inclination angle in the forward flow direction showing a) initial wetted zone (16mm from the centre) and b) outside initial wetted zone (112mm from the centre) 92 Figure 4-22: Effect of flow rate for a jet with a 30° inclination angle in the forward flow direction in a region close to the initial wetted region (48mm from centre) 92 Figure 4-23: Effect of flow rate at 20° inclination in reverse flow direction outside the initial wetted zone (32mm from the centre) 93 Figure 4-24: Overall energy dissipation versus flow rate 94 Figure 4-25: Characteristic cooling curve for moving strip for 10 passes from moving plate experiment (1 m/s strip speed, 0° inclination angle, 45 1/min flow rate) 95 Figure 4-26: Effect of strip speed (pass #5, flow rate = 45 1/min, inclination angle = 0°) 96 List of Figures x Figure 4-27: Effect of strip speed (pass #6, flow rate = 45 l/min, inclination angle = 0°) 97 Figure 4-28: Effect of strip speed on heat flux versus surface temperature (0° inclination angle and 45 l/min flow rate) 98 Figure 4-29: Effect of strip speed on heat flux versus time 99 Figure A- l : Change in heat flux for mesh variance above thermocouple tip 114 Figure A-2: Change in heat flux for mesh variance below thermocouple tip 114 Figure A-3: Effect of thermally dependent conductivity on heat flux 116 Figure A-4: Relative error in heat flux with respect to surface temperature 117 Figure A-5: Relative error in heat flux with respect to time 117 Figure B - l : Equivalent depth technique validation of TCI for experimental (0° inclination and 30 l/min flow rate) data taken from (Zhang, 2004) 119 Figure B-2: Equivalent depth technique validation 119 Figure C-l: Measured internal temperature curve with inclination angle of 10° and 35 l/min flow rate in forward flow direction 121 Figure C-2: Comparison between internal measured and estimated off-axis internal temperature curves from experiment #8 (45 l/min flow rate and 30° inclination angle) 122 Nomenclature Nomenclature cp Specific heat (J7kg°C) dj . Circular jet diameter at exit of the nozzle diameter (nozzle diameter) (mm dy Jet diameter at impingement (mm) D * Diameter of test cell (mm) DP Discrete Point F Liquid solid contact area fraction (%) g Gravitational acceleration (m/s2) h Heat transfer coefficient (W/m2°C) H Distance between the nozzle and the heated plate (mm) hig Latent heat of evaporation from liquid to vapour (J/kg) Ja Jakob number ((p,/pg)(cp • ATsuh I \ ) ) k Thermal conductivity (W/m°C) L Characteristic length (mm) P Pressure (Pa) Pr Prandtl number (pcp/k) q" Heat flux (W/m2) r Radial coordinate (mm) Re Reynolds number (vL/v) t Time (second) T Temperature (°C) Tsat Saturation temperature of fluid (°C) v Velocity (m/s) Vj Jet velocity at exit of the nozzle (m/s) vy Jet velocity at the impingement point (m/s) W Width (mm) Z Distance between the nozzle and the heated plate (mm) Nomenclature xii Greek symbols a Regularization parameter AT s a t Wall superheat, the difference between surface temperature and fluid saturation temperature (°C) ATSUb Subcooling, the difference between fluid saturation temperature and fluid temperature (°C) u. Dynamic or absolute viscosity (Pa s) v Kinematic viscosity (p/p) (m2/s) 6 Inclination angle (°) 6DP Incident angle with respect to forward flow axis (°) p Density (kg/m3) CT Surface tension (N/m) CTSB Stefan-Boltzmann radiation constant (5.67x 10"8 W/(m2oC4)) Subscripts B Black region CHF Critical heat flux FB Film boiling FNB Fully developed nucleate boiling g Vapour i Impingement ini Initial j Jet 1 Liquid MHF Minimum heat flux n Nozzle NB Nucleate boiling ONB Onset nucleate boiling Nomenclature xiii sat Saturation sub Subcooling TB Transition boiling w Wall wet Wetted Zone Acknowledgements Acknowledgements xiv I would like to express my sincere thanks to Dr. Mary Wells, Dr. Vladan Prodanovic, and Dr. Matthias Militzer for their invaluable guidance and support throughout my study and for granting me the opportunity to work on this project. Thanks to Ross Mcleod, Carl Ng, David Torok, Serge Milaire, and Gary Lockhart for help with the experiments. I am thankful to Nancy Oikawa and Mary Jansepar for all their help. I also wish to express my appreciation to all my colleagues, especially to Kailas Jondhale, for their friendship. Dedication Vincit qui Patitur (He conquers who endures) Aulus Persius Flaccus Chester Family Motto Introduction 1 Chapter 1 - Introduction The transient cooling of hot surfaces by liquid jet impingement is commonly utilized in industrial processes, especially in hot rolling of strip steel where it plays an essential role. For hot rolled steel, after passing through the finishing mill, the steel continues on to the runout table where arrays of top and bottom water jets are employed to cool and control the temperature profiles in the strip. When the boiling phenomenon is used effectively in a technological application such as runout table cooling, it is able to remove enormous quantities of heat while ensuring that the end product has reached the desired mechanical properties. Studying heat transfer on a full scale runout table is difficult, and very expensive to do. As an alternative method, researchers often perform small laboratory experiments, which include either one jet or an array of jets with round or rectangular nozzles, impinging on a stationary or slowly moving test specimen. The effects of various parameters such as flow rate, subcooling, wall superheat, and inclination on the cooling have been investigated to some or very limited extent previously. Many of these investigations to date have primarily dealt with top jet impingement, and limited information is available for bottom jet cooling. Thus, the general purpose of the present study is to obtain fundamental knowledge on the heat transfer during bottom jet impingement cooling. As well, inclination of the nozzle will also be considered since bottom jets are often set obliquely to the plate in industrial runout table processing. Literature Review 2 Chapter 2 - Literature Review The demand for highly tailored steels has placed considerable emphasis on the type of cooling method employed during steel strip production. In particular, controlled cooling during runout table processing has attracted significant research interest since the microstructural evolution, thus the final mechanical properties of the steel are strongly influenced by the cooling path (Militzer et al., 2000; Liu, 2001; Prodanovic et al., 2004). The usual cooling method employed in hot strip mills is by water jet impingement. To date, extensive theoretical and experimental work on jet impingement cooling has been described in various reviews (Kalinin et al. 1975, 1987; Wolf et al., 1993; Webb et al., 1995). Much research has been carried out for steady-state cooling experiments in which the specimen is heated and maintained at a set energetic level during jet impingement. However, very limited research has been carried out for transient cooling processes in which the specimen is allowed to cool down to room temperature. Transient cooling experiments in a laboratory are much closer to industrial conditions seen in runout table processing. The purpose of this chapter is to present theoretical and experimental work done to date on runout table cooling with focus on the modes of heat transfer and jet hydrodynamic characteristics that occur during runout table cooling. Literature Review 3 2.1 Industrial Runout Table Layout The runout table operation is one of the last processes during hot rolling of steel in which water is applied to the hot steel strip to cool it down to a desired coiling temperature. As shown in Figure 2-1, the strip is subjected to jet cooling as it exits the finishing mill and passes through the runout table before entering the coiler for storage and perhaps further processing. Finishing Mill Exit Temperature C o i l i n g Temperature Finishing Mill ^ j ^ ° n k H e a d e r = ll : 'IV =11- '[V -IV •»' : | | : : | | : : | | : M l 'll' 'll' 'll' 'ill l|| ||l l||l ||l lMl u u u u u u u u u u u u u u u u u ri ri ri ri ri Bottom Jets Down Co i l e r ri ri ri ri ri ri Figure 2-1: Schematic of a runout table (modified from Liu, 2002). The temperature of the hot rolled strip usually lies in the range of 800 - 950°C at the last stand of the finishing mill and has a temperature in the range of 500 - 750°C before entering the coiler (Liu, 2001). Along the runout table a number of water banks are employed to supply water to the headers. There are several headers per water bank and each header consists of one or two rows of jet lines. Typically, the water banks are mounted at the top and bottom of the strip. Top jet banks are normally gravity-fed by water towers. Bottom jet banks utilise pumps to supply water from the storage tank to the steel strip. The water is mostly recycled throughout the day such that a rise in water temperature is observed as well as an increased presence in impurities (e.g. oil in the Literature Review 4 water), which can affect the effectiveness of runout table cooling. To ensure that there is stable water flow in the jets the level of water in the storage tanks are maintained at a relatively constant level. There are three types of nozzles employed in runout table cooling; round tube nozzles for laminar flow, slot-type nozzles for water curtain cooling, and spray nozzles for spray cooling systems, as illustrated in Figure 2-2. Laminar Bottom Jet Cooling Water curtain Spray cooling Figure 2-2: Three cooling systems. A study was performed by Tacke et al. (1985) to gauge the cooling capacities of the three aforementioned cooling systems. It was found that water curtain and laminar flow cooling had higher cooling capacities than spray cooling thus greatly increasing overall heat transfer rates. Results from Kohring et al. (1985) had similar findings which Literature Review 5 showed laminar flow cooling to be about 10-30% more efficient than spray. Water curtain and laminar flow are desirable because it allows for close contact with the cooled surface with negligible splash and thus a long liquid-solid contact time (Han et al., 1991). Further Zumbrunnen et al. (1989) and Chen et al. (1992) were able to show that water curtain cooling provided a uniform cooling rate across the width of strip since the water jet spanned the entire width of the strip. They also showed that between planar and circular jets, planar jets were 48% more efficient at removing heat in the case of a test strip heated to 900°C and traveling at lOm/s. However, laminar flow cooling systems can provide more efficient cooling per unit volume of water. The disadvantage for this type of cooling system is in a non-uniform cooling rate that occurs across the width of the strip. 2.2 Hydrodynamics of Jet Impingement 2.2.1 Hydrodynamic Jet Configurations Jet impingement hydrodynamics describes the interaction between water flowing in contact with an impingement surface, which in this instance is a flat steel plate. The importance of knowing this interaction is that the mode of heat transfer is strongly dependent on the hydrodynamics of flow. Five different hydrodynamic jet configurations used for jet cooling have been described by Wolf et al. (1993), which are shown in Figure 2-3. They are; free surface jets, plunging jets, submerged jets, confined jets, and wall jets. Literature Review 6 Nozzle Free Surface Stagnation Point Centreline of Jet Liquid A. Free-surface Gas Liquid \ \ \ \ \ \ \ \ \ \ \ \ \ \ \ \ \ \ \ \ \ \ V \ . \ \ \ \ \ C. Submerged Nozzle Plate Liquid Liquid B. Plunging Nozzle Plate Gas "* Liquid ** D. Confined Gas \ \ \ \ \ \ \ \ \ \ \ \ \ V A \ \ V A \ V A V \ A \ E. Wall (free-surface) Figure 2-3: Types of jet configuration (modified from Wolf et al., 1993). In rolling mills, however, the conventional configuration is for either free surface and/or plunging jets depending on the length of the runout table and the position of a particular jet along the runout table. A free surface jet configuration is typically observed at the start of runout table operations when liquid travels relatively unobstructed from the nozzle to the impingement surface of the strip. As the strip travels through the runout table a water layer builds up on the strip surface. In this instance a plunging jet configuration is observed since the water jet has to pierce a pooled water layer on the Literature Review 7 impingement surface. It has been observed that as the thickness of the water layer increases the effectiveness of the jet at removing heat decreases in which case the maximum heat transfer occurs for the free surface configuration (Summerfields, 2004). Jets are further sub-classified by the nozzle shape (i.e. rectangular, cylindrical, etc.), and by the orientation of the nozzle; for instance, the jet can be set normal or obliquely to the impingement surface. During runout table processing, bottom jets are often set obliquely to the impingement surface and angled against the direction of strip velocity. 2.2.2 Hydrodynamic Regions Three hydrodynamic flow regions have been defined for a free surface jet with uniform velocity profile impinging on a stationary plate. These flow regions, depicted in Figure 2-4, are demarcated by the pressure and velocity distributions from the stagnation point to points further along the surface of the plate. For a stationary plate being impinged by a jet normal to the impingement surface, the stagnation region is found immediately underneath the jet and is at least the size of the jet diameter (r/djj > 0.5). Where r is the radial distance along the impingement surface from the stagnation point and dy is the diameter of the jet at impingement. Literature Review 8 Nozzle Free Surface Stagnation Point Centreline of Jet \ V \ \ \ \ \ \ \ \ \ \ ^ \ \ V \ \ \ \ \ \ \ \ \ V \ * X A. Stagnation Region B. Acceleration Region C. Parallel Flow Region Figure 2-4: Pressure and velocity profile, and flow regions for a jet with a uniform velocity profile (modified from Wolf et al., 1993). The maximum pressure and the lowest streamwise velocity occur at the stagnation point, which corresponds to a point along the centreline of the jet at the impingement surface. The flow is deflected after impingement and starts to travel along the impingement surface. As the flow travels radially outwards from the centreline of the jet, the maximum pressure decreases and the streamwise velocity accelerates. This region is denoted as the acceleration region, which normally ranges from 0.5 < r/dy < 2 (Hauksson, 2001). In the parallel flow region (r/dy > 2), which develops after the acceleration region, the pressure gradient approaches zero (i.e. ambient conditions) and the streamwise velocity approaches that of the jet velocity (Hauksson, 2001; Meng, 2002). In the case of Literature Review 9 bottom jet impingement, due to the gravitational forces acting on the water no parallel flow region exists. The impingement region, or impingement zone, refers to the cross-sectional area of the jet at the impingement surface. It has been observed that as the liquid travels further from the stagnation point the thickness of the liquid layer decreases. Further downstream the transition to turbulent flow can take place. In runout table cooling the strip moves through the runout table at a high velocity. A moving surface can have a pronounced effect on the hydrodynamics and the velocity profile of a free surface jet impinging on a flat surface (Zumbrunnen et al., 1992). These differences in velocity profile are depicted in Figure 2-5, showing a free surface top jet impinging normally on a flat surface in a (a) stationary surface condition, and (b) moving plate condition. The reference frame origin on Figure 2-5 corresponds to the impingement point on the plate. Figure 2-5: Hydrodynamic boundary layer and velocity profile of a planar top jet impinging on a (a) stationary plate, and a (b) moving plate (modified from Zumbrunnen etal., 1992) In the case of a stationary surface, a symmetric velocity condition is shown on either side of the jet. However, in a moving surface condition, the downward side, where Literature Review 10 the flow and plate run concurrently, an enhancement of velocity is shown due to jet flow and surface traveling in the same direction. On the upward side, where flow runs counter to the moving plate, the momentum of the fluid is greatly reduced because jet flow and moving surface run in opposite directions. The differences in velocity profile shown for bottom jet impingement are depicted in Figure 2-6, showing a free surface jet impinging (a) normally and (b) obliquely to the stationary surface. In the case of a jet impinging normally on a stationary surface, a symmetric velocity condition is shown on either side of the jet. When the jet is inclined, an enhancement of velocity is shown due to an increase in flow rate in the downward side (i.e. in the direction where the jet is inclined). On the upward side, which is in the opposite direction where the jet is inclined, the momentum of the fluid is greatly reduced because jet flow rates are reduced. Figure 2-6: Hydrodynamic boundary layer and velocity profile of a planar bottom jet impinging (a) vertically and (b) obliquely on a stationary plate. Literature Review 11 2.3 Boiling Heat Transfer Mechanisms 2.3.1 Pool Boiling As mentioned before, limited research has been performed in gaining fundamental knowledge of boiling heat transfer during jet impingement in which the specimen is allowed to cool down to room temperature under transient conditions. As a first step towards understanding boiling heat transfer, it is important to be aware of the different heat transfer mechanisms that occur during pool boiling before an analysis can be performed for jet impingement boiling. Boiling heat transfer behaviour is best described by a plot of heat flux as a function of wall superheat. Wall superheat is defined as the thermal differential between the surface temperature of the heated solid and the saturation temperature of the coolant (AT s a t = Tw aii - T s a t). Single-Phase Forced Convection Regime Nucleate Boiling Regime Transition Boiling Regime Film Boiling Regime D B' x LL Onset of Nucleate Boiling Minimum Heat Flux Wall Superheat logAT, Figure 2-7: Typical pool boiling curve. Literature Review 12 As shown in Figure 2-7, which is a schematic of a typical pool boiling curve, there are 4 different heat transfer regions that exist on the boiling curve. Arranged in increasing wall superheat, they are single-phase convection, nucleate boiling, transition boiling, and film boiling. Point A on the pool boiling curve denotes the start of boiling on a heated surface. This point demarcates a transition from single-phase convection to the incipience of nucleate boiling, also known as onset of nucleate boiling (ONB). At this stage small bubbles begin to form in tiny crevices on the surface. Heat is transferred from the heated surface via conduction from the plate through the liquid, evaporation of the liquid, and micro-convection within the fluid caused by bubble growth and detachment. As the surface temperature is increased to point A' on the pool boiling curve the heat transfer mechanism changes from partially developed nucleate boiling to fully developed nucleate boiling. During this region maintenance of bubble growth and subsequent detachment into the colder fluid above is sustained. A continued increase in the surface temperature causes a subsequent increase in the heat flux throughout the nucleate boiling regions up to a maximum point known as the critical heat flux (point C). This is due to an enhancement in mixing associated with micro-convection during bubble growth and departure, and with intimate water-solid contact on the heater surface. With increasing wall superheat, the point at which the critical heat flux occurs will change into the transition boiling regime, found between points C and D on the boiling curve. With it comes a sharp decrease in heat flux with an increase in wall superheat. Conditions are highly dynamic such that short periods of nucleate boiling and film boiling occur on the same heater surface during wet and dry periods, respectively. Literature Review 13 The variation in heat flux with wall superheat is a result of change in the fraction of time each boiling mode is present on a given area (Dhir et al., 1998). With continued temperature increases, this would result in a decrease in the area fraction covered by the liquid-heat surface interface. This drop in heat flux is mainly a result of significant bubble generation and coalescence of bubbles on the heater surface since water-vapour has a significantly lower capacity to remove heat than that of liquid water. With a continued rise in wall superheat, the surface temperature drops to a minimum heat flux, which is denoted as the Leidenfrost point (point D). Beyond point D, marks the final transition where the heated surface is able to supply enough heat energy to support a continuous vapour barrier. The mode of heat transfer is primarily by convection of the vapour, with radiation becoming dominant at higher surface temperatures (Wolf et al., 1993). From this point towards continuously increasing wall superheats, a stable film boiling regime is said to exist. This effectively separates the liquid contact from the heated surface. Increases in surface temperature is a result of poor heat transfer capacity of water-vapour in removing heat. 2.3.2 Jet Impingement Boiling Regions Similar boiling regions are observed in jet impingement boiling, although the shape of the boiling curve may become distorted as the effect of forced convection becomes important. During transient cooling of a hot steel plate by a liquid jet, different heat transfer rates are observed at different distances from the stagnation point. This results in large variations of the surface temperature and consequently different boiling regimes occur along the plate at the same time. It is expected that all different boiling Literature Review 14 mechanisms that take place during pool boiling occur for jet cooling on the runout table since the strip surface is subjected to a range temperatures, usually from 900°C down to 300°C. Figure 2-8 illustrates the heat transfer regions during top jet impingement. Very shortly after impingement a highly wetted zone appears underneath the jet, shown as Region I. The cooling effectiveness is thought to be very high; however, the surface temperature is below levels required for boiling to occur (Zumbrunnen et al., 1989). Thus, in this region, single-phase forced convection is the primary mode of heat transfer. Region II takes up a very narrow band on the impingement surface, where both nucleate and transition boiling regimes are thought to exist. Region III occurs as the distance from the stagnation point is increased further where the bubbles begin to coalesce into the formation of a stable vapour film barrier. In Region IV, the liquid layer begins to thin down as the distance from the jet increases, isolating water into droplets on the heater surface. Film boiling still occurs but much less effectively since the water supply is removed. Beyond this radiation and air convection are said to be the primary method of heat removal. Figure 2-9 illustrates the heat transfer regions during bottom jet impingement. It can be seen that the boiling regions are similar to top jet except for the absence of Region IV (i.e. agglomerated pooled zone). This is due to the stable vapour film barrier along with gravitational forces deflecting water from contacting the impingement surface. Thus, water is unavailable to pool, and consequently a reduction of the heat transfer rate is observed. Literature Review 15 Liquid Jet Regions I - Single Phase Forced Convection II - Nucleate / Transition Boiling III - Forced Convection Film Boiling IV - Agglomerated Pools V - Radiation and Convection to Surroundings Steel Plate Figure 2-8: Different boiling regions present during top jet impingement [modified from Zumbrunnen et al. (1989)]. >* ii * H — M I — H « - v i ^ 4 C Regions I - Single Phase Forced Convection II - Nucleate / Transition Boiling III - Forced Convection Film Boiling IV - Radiation and Convection to Surroundings Liquid Jet Figure 2-9: Different boiling regions present during bottom jet impingement. Literature Review 16 Colas et al. (1987, 1994), Evans et al. (1993), and Prieto et al. (2001) studied the modes of heat transfer present in each hydrodynamic region during impingement by a circular, free surface water jet. It was found that the hydrodynamic regions can be used as a rough classification for the different cooling zones present. In the impingement zone, which incorporates both the stagnation and acceleration zones, the predominant cooling mechanism is by forced convection. In the parallel flow zone, the primary mode of heat transfer is film boiling since a stable vapour-film barrier develops in that region. After the parallel flow zone the primary cooling method is by radiation and occurs under no liquid-solid contact. For bottom jet cooling, however, there is essentially little or no parallel flow zone since gravity forces the liquid away from the surface of the strip. As well, a build up of a stable vapour-film barrier on the bottom surface would only enhance jet deflection. It should be noted that the heat transfer mechanisms are highly dependent on a number of factors, which can alter the time or temperature at which certain boiling regimes occur. Such factors include surface temperature, the degree of subcooling, strip velocity, water jet flow rate, thermal material properties of the heated solid and the cooling fluid such as thermal conductivity, and system geometries such as nozzle type and configuration, etc. 2.3.3 Wetting Front Analysis During transient cooling of a hot plate by an impinging water jet a pronounced temperature profile develops close to the stagnation point and becomes reduced as the distance from the stagnation point increases. Often along the time-temperature profile Literature Review 17 curve an abrupt and sharp change in surface temperature will occur, denoting the existence of two distinct regions on the surface of the plate during jet impingement; the wetted and the non-wetted zone. This demarcation between the highly wetted and non-wetted zones is known as the wetting front. Despite a significant number of experimental studies carried out recently by a number of researchers (Woodfield, 2005; Mozumber, 2005; Prodanovic, 2006) there is still limited knowledge in the wetting front phenomenon. For jets perpendicular to the impingement surface, the wetted zone forms into a circular region about the stagnation point shortly after impingement. Within the wetted zone, a very high degree of liquid-solid contact occurs whereby the liquid progresses outward unobstructed up until it becomes deflected from the surface by the presence of a stable vapour film barrier. Consequently, the boiling modes are typically transition or nucleate boiling followed by single-phase convection in the wetted zone and film boiling in the non-wetted zone. The most rapid cooling occurs when a region of the cooling surface is in the presence of the wetting front. This would suggest that the maximum heat flux can also be placed in the vicinity of the wetting front, thus confining the whole transition boiling region into a very narrow area about the border of the wetting zone (Filipovic et al., 1995c). The wetting front progression is delayed just after impingement at an initial radius, which has been found to depend on impact velocity and degree of subcooling. Essentially, a highly wetted zone remains at a constant size immediately underneath the jet for a certain period of time before the wetting front begins to grow. Liu (2001) Literature Review 18 observed that the initial radius of the wetted zone to be about 1.6-2.1 jet diameters whereas Hauksson (2001) reported smaller radii ranging from 0.9 - 1.7 jet diameters. A study of the transient boiling heat transfer during cooling by a bottom water jet was carried out by Mitsutake et al. (2001) to observe the behaviour of the wetting front. They were able to demonstrate that the speed at which the wetting front moves is very important in predicting the total amount of heat removed. Their data were fitted to a power function with respect to time as shown in equation 2-1. r^=a-tn (2-1) where a and n were experimentally determined, t is the time, and r w e t is the radius of the wetted zone from the centreline of the jet. An important finding relevant to the current study was that the wetting front spread faster with increased subcooling and jet velocity. Mozumder et al. (2005) conducted transient heat transfer experiments using an upward facing subcooled water jet to simulate quench cooling processes. Results showed that at a time just after the jet strikes the hot surface the maximum heat flux is always attained, which is at the same time the wetting front starts moving. Prodanovic et al. (2006) analysed the behaviour of the wetting front with respect to the flow rate and subcooling, based on a number of top jet impingement boiling experiments carried out using an industrial scale circular water jet (dj=19mm) impinging on a horizontal steel plate under transient conditions. The effect of subcooling and water flow rate in the delay and progression of the wetting front, and the initial size of this wetting zone, was investigated in the range from 5-70°C, and 15-45 l/min, respectively. The temperature at which the wetting front passed over a thermocouple location tended to fall between 400-500°C, and occurred near the start of the sharp transition on the time-Literature Review 19 temperature curve. Flow rate was found to have a significant effect on the progression of the wetting front, which is especially pertinent to the current study. At a AT S U D = 30°C, it took the wetting front approximately 65 s to reach a position 95mm from the stagnation point in the case of 15 1/min, but only approximately 45s in the case of 30 1/min. A recent study was carried out by Hammad et al. (2004) to observe the progression of the wetting front for a bottom jet impinging on a hot copper block. Some interesting results revealed that the wetting front position does not occur at the maximum heat flux. It was also found that not to occur in the transition boiling regime, as previously thought (Filipovic, 1995c), but in a region that is highly wetted. Another interesting finding the authors revealed was that the maximum heat flux does not start at the same time when the jet water strikes the heated copper block, but starts when the surface temperature decreases to a value less than 170°C. 2.4 Investigative Methods for Jet Impingement Experiments There are two types of investigative methods employed in jet impingement cooling experiments. These methods include steady-state cooling experiments carried out in either temperature-controlled or heat-flux controlled conditions, and transient cooling experiments in which the heated specimen is allowed to cool down from an elevated temperature. In the instance of a steady-state condition, the specimen is heated and maintained at a set energetic level by either a temperature controlled or heat flux controlled system while being cooled by a regulated jet stream run concurrently until a steady-state condition is achieved (Robidou et al., 2002, 2003; Fry et al., 1997). In this form of Literature Review 20 measurement, the heat flux removed by the cooling system is continually replenished, usually by an induction heater being supplied electrical energy input. Thus the effectiveness of a jet configuration can be estimated directly from the amount electrical input that is being supplied, which greatly facilitates obtaining the heat transfer coefficients. In contrast, transient cooling experiments are performed under cooling conditions that have no fixed surface temperature or constant surface heat fluxes. With this approach, the test plate is heated up to the test temperature and then cooled under jet impingement. During this process, a temperature-time history of the cooled plate is recorded with strategically placed thermocouples. Mathematical methods are then employed to determine the heat fluxes and/or heat transfer coefficients by using the measured thermocouple data (Zhang, 2004). Differences between the two forms of experimental study were shown in recent work by Heas et al. (2003) who examined boiling incipience obtained on wires and ribbons. Their findings showed that under transient conditions, boiling is delayed leading to much higher temperatures at which boiling incipience occurs. This phenomenon was thought to be related to a decrease in nucleation sites promoting bubble growth. 2.4.1 Steady-State Method Miyasaka et al. (1980) carried out steady-state experiments to observe the heat transfer mechanisms during the transition of the nucleate boiling regime up to the CHF. Data were collected at the stagnation point while being cooled by an upward facing water jet. The high heat flux was generated on a heat transfer surface fixed to a copper heat conductor. The test surface was made of a 0.05mm thickness platinum foil which was Literature Review 21 attached to a 1.5 mm diameter copper cylinder of 3 mm length. The heat was generated by an AC power supply controlled by a 10-KVA transformer. Two Copper-constantan thermocouples were attached to the copper cylinder 1.0mm and 2.0mm above the platinum foil/copper cylinder interface. The measured temperature differential provided a means of calculating for heat flux by applying Fourier's Law for steady-state conduction. Superheat ranged from 10-800°C, and the effects of subcooling and jet velocity were varied 0-85°C and 1.5-15.3m/s, respectively. The maximum heat flux attained was 46.7 MW/m . The authors were able to find that the nucleate boiling curve appeared as a linear extension to the ordinary pool boiling curve. As well, their data seems to show the existence of heat transfer transition regions from the end of the nucleate boiling regime to the CHF. Wang et al. (1997) conducted a steady-state boiling heat transfer study with a single plane jet fixed parallel to a horizontally set surface. The purpose of this study was to investigate the heat transfer mechanisms that occur in the hydrodynamic parallel flow zone. The heater surface was rectangular and variable at 40, 60, and 80mm in length and 10 and 20 mm in width. The slot jet could be varied from l-2mm in height. Jet velocity was determined by measuring the pressure difference between the inlet and outlet of the nozzle by means of a differential pressure transducer, and was varied from 3 to 15 m/s. Subcooling was controlled through a combination heater / cooling coil, and was varied from 0 to 60K. The specimen was well insulated such that the heat generated in the foil was properly transferred to the test liquid. Therefore, the heat flux across the surface was determined by assuming the heat was transferred completely into the liquid across it. The electric power to the heated surface was increased incrementally until a point was Literature Review 22 reached where the surface temperature ran away to burnout. The researchers developed an improved correlation for CHF data over existing correlations in the case of a highly subcooled liquid under forced convective boiling. Under steady-state heating conditions Robidou et al. (2002, 2003) conducted experiments to study the boiling heat transfer that occurred during planar jet impingement on a copper block. The effects of subcooling, jet velocity, and the nozzle to plate spacing were investigated. They found that at higher subcooling (i.e. at 17°C) the transition boiling regime existed in a larger range of wall superheat. The effect of jet velocity on the boiling curve at stagnation was found not have an influence on the transition boiling regime, as well as, the nozzle to plate spacing. Liu et al. (2004) investigated the effects from a subcooled circular water jet on the nucleate boiling regime and the CHF for a test surface that was upward facing. The tests were performed under a steady-state condition where wall superheat was varied from 10-600°C. During experimentations standoff distance fixed at 5mm, however, impact velocity varied from 0.5-6.0m/s, nozzle diameter varied from 3-12mm, and subcooling was changed from 15-80°C. A solid copper cylinder of 50mm diameter and 120mm length was heated by a ribbon electric heater wrapped around the cylinder. Three thermocouple wires were attached parallel to the impingement starting 2mm below the impingement surface in 3mm increments. The heat fluxes obtained were in the range of approximately 1-10 MW/m 2 The authors' findings were that subcooling, impact velocity, and nozzle diameter had a strong effect on the CHF in which the highest CHF was achieved with higher jet velocities, lower jet diameter, and higher subcooling. The Literature Review 23 authors were also able to show that the relationships from saturated liquid conditions were relevant to subcooled conditions. 2.4.2 Transient Cooling Method Ishigai et al. (1978) conducted both transient and steady-state boiling at the stagnation point with a planar water jet. The effects of jet velocities and subcooling were investigated. With higher jet velocity the boiling curve shifted to higher wall superheat and heat fluxes. In the case of higher subcooling the boiling curve increased monotonically to higher heat fluxes. Ochi et al. (1984), who investigated the boiling heat transfer on a hot stainless steel plate impinged by a circular water jet, reported similar trends to Ishigai et al. (1978). They varied nozzle diameters, subcooling, and jet velocity from 5-20 mm, 5-80°C, and 2-7 m/s, respectively. Kumagai et al. (1995a, 1995b) performed transitory experiments on a two-dimensional 1x28mm impinging water jet cooling an upward facing surface. The test plate was made of copper and had a rectangular shape with dimensions of 20mm width, 150mm length, and 120mm in thickness. The jet velocity was maintained at a constant 3.5m/s. The effect of subcooling was the primary focus and was varied from 14-50°C. The impingement surface was heated up to an experimental start temperature of 400°C before experimentation began. From the measured data, the researchers developed transient profiles of the surface temperature and heat flux across the impingement surface. One of their findings showed that for cases above 14°C subcooling, a solid-liquid contact occurred almost instantaneously at stagnation, however, the liquid film was violently splashed further down resulting in a small amount of liquid supplied to the Literature Review 24 wetting front. The temperature at which the initial wetting zone began increased with increasing subcooling and that the initial wetted zone was narrowed closer to the stagnation point. Another important finding was that the temperature differential at the wetting front was maintained constant along the impingement surface even at regions very far from stagnation. This finding could suggest that the wetting front is progressing at a constant rate. Filipovic et al. (1995a, 1995b, 1995c) conducted a number of transient boiling experiments on a large, preheated test specimen exposed to a water wall jet on its top surface to gain insights concerning nucleate, transition, and film boiling. Parameters of interest included the initial plate temperature, the cooling water temperature, and the flow velocity. Temperatures were measured at discrete locations on the surface and were used with a solution of the transient heat conduction equation to obtain the temperature field throughout the specimen, as well as the heat flux distribution at the surface. The test apparatus used an oxygen-free copper block with thermal conductivity (k = 350 W/mk) and dimensions (252mm thick, 38.1 mm wide, and 0.508m long). The initial test cell temperature ranged from 700-810°C, and water temperature ranged from 25-55°C. Jet impingement velocities were considered in the range 2-4 m/s. Results indicated that the wetting front velocity increased with higher subcooling and jet velocity. The wetting front surface temperature was found to decrease as distance from the impingement point increased due to decreases in local subcooling. As well, the wetting front velocity was found to increase near the edge of the test specimen since the stored energy available in the vapour film barrier was decreased. Literature Review 25 In order to study industrial runout table cooling, Liu (2001) designed and constructed a pilot scale runout table facility at UBC, which was equipped with industrial sized nozzles. A steel plate with a size of 280x280x7.6 mm is heated in the furnace to a desired initial temperature and then cooled- with water jets. Liu (2001) carried out stationary plate experiments to gauge the effect of subcooling and flow rate. Hauksson (2001) conducted stationary, plate experiments with a vertical downward facing jet to observe the effect of varying water temperature and flow rate, which ranged from 30-50°C and 15-45 l/min, respectively. Meng (2002) carried out similar experiments although at a higher water temperature range from 60-95°C. It was found that having higher subcooling and higher flow rates increased heat fluxes, and that the maximum heat flux was always obtained at stagnation. 2.4.3 Review of Bottom Jet Experiments Monde et al. (1994) performed steady-state experiments on a downward facing surface to observe the effect forced convective subcooling had on the CHF at stagnation. A circular bottom jet of 2mm diameter was used on a stainless steel foil heater of 0.3mm thickness, 7mm width, and a variable length of 40 to 60mm. Heat was supplied to the test sample by two copper blocks connected to a DC power supply, and the heat input was regulated by a computer. The test specimen was heated under steady-state and subcooling, jet velocity and pressure were varied from 0-115°C, 5-16m/s, and 1-3 bar, respectively. The heat fluxes obtained at the CHF were in the range of 2-10 MW/m before burnout occurred. The temperature change of the heated surface was obtained by the measured change in the resistance of the foil heater. The authors were able to show Literature Review 26 that a modified equation developed for saturated liquid conditions could be used to determine the CHF for subcooled convective boiling (i.e. factoring in the effect of subcooling, jet velocity and stagnation pressure). This is shown as Equation 2-2. C = 0-221 where f ^ 0.645 ^ Pi f 2CJ Ptv:{D-d\ 1 + VT+4C • Ja (X + Dld)"™ (2-2) 0.95 C = - I (1 + — )-°"364 (2-3) y- N. 0.645 s 2a p,v;'{D-d\ where D is the diameter of the heater surface, and dj is the diameter of the cylindrical jet. Mozumder et al. (2005) and Woodfield et al. (2005) conducted transient heat transfer experiments using an upward facing subcooled water jet to simulate quench cooling processes. The liquid jet issued from a nozzle 2mm in diameter and centrally located 45 mm from the test surface. Velocity and subcooling were varied 3-15 m/s and 5 to 80 K, respectively. The research findings by Mozumder et al. (2005) showed that at a time just after the jet strikes the hot surface and the wetting front starts moving, the maximum heat flux is always attained. Findings from experiments carried out by Woodfield et al. (2005) showed that the Leidenfrost point was not clearly defined at the initial solid temperature of 350°C, and it was thought that the Leidenfrost point shifted to higher surfaces temperatures. These findings were consistent with previous results (Ishigai, 1978) which places it at AT s a t = 700K (@ AT S U D = 45K) for impinging circular water jets. Literature Review 27 Mitsutake et al. (2001) conducted a transient boiling heat transfer study with a single bottom water jet impinging upward on a hot cylindrical copper block. The block was heated by slot heaters to an initial temperature of 250°C. The experimental data were taken for the following conditions: a degree of subcooling of AT s a t = 20-80K, a jet velocity of u = 5-15m/s, a nozzle diameter of 2mm, and three copper materials of copper, brass, and carbon steel. The water with three different subcooling was fed to the nozzle by a pump. Sixteen thermocouples of 0.1mm in diameter were used to measure the temperatures at eight radial locations. Each location included a pair of thermocouples, which were located at 1mm and 5 mm below the wall, respectively. The wall heat flux was evaluated with the measured temperatures at two different depths using a numerical model of two-dimensional heat conduction. The authors were able to demonstrate that the advancement of the wetting front radius was well correlated with a power function, and that the nucleate boiling regime seems to occur where the wetting front is, which is where the peak heat fluxes occur. Zhang (2004) studied the effect of subcooling with a single circular free-surface jet on a downward facing hot surface. The steel plate was heated in an electric furnace up to a temperature of approximately 900°C. The experimental conditions involved varying the degree of subcooling of ATSUb = 50-80K and a flow rate ranging from 30-60 l/min from a fixed nozzle diameter of 18mm. A total of eight internal thermocouples were instrumented to the test plate 1mm below the impingement surface at increments of 16mm from the stagnation point at different radial locations. The wall heat fluxes were generated from a numerical two-dimensional inverse heat conduction (IHC) model using the measured internal temperatures. The author was able to show that increasing Literature Review 28 subcooling had the effect of increasing heat fluxes, however, the maximum heat flux shifted to lower surface temperatures at the stagnation point. The effect of increasing flow rate was also studied and was found to increase heat fluxes, although, no influence within the nucleate boiling regime was observed. Another interesting finding by the author was that the wetting front velocity increased with decreasing water temperature and increasing flow rate. 2.5 Effect of Inclination In this section, a discussion will be presented of recent papers published on the effects of nozzle inclination on the boiling heat transfer during jet impingement. It should be mentioned that the experimental data collected in literature reviewed for boiling heat transfer has usually been correlated according to factors such as shape and size of the heating surface and fluid jet. However, orientation between the jet and the heating surface has been treated as a secondary factor and its effects have been eliminated from even.the most detailed considerations (Nishikawa et al., 1983). It is partially due to the fact that there is very limited available information concerning those effects on boiling. Orientation may give rise to significant differences in heat transfer, and consequently becomes very important to investigate. Often in industrial steel production bottom jet nozzles are inclined at a certain angle relative to the strip surface. The effects of surface orientation have been investigated. For cases involving pool-boiling, or in the cases of jet impingement, air is employed as the coolant and the jet is usually fixed as downward or upward facing and the heater surface is rotated. Some early work on the effect of relative heater/jet orientation was carried out by Githinji et al. Literature Review 29 (1963) who studied subcooled nucleate pool-boiling with isopropyl alcohol and found that the boiling curve shifted upwards as the surface was rotated from a horizontal facing surface to a vertical position. It was also found that higher heat fluxes could be obtained for a horizontal surface which is upward facing compared with a heating surface that is downward facing. Marcus (1963) reported that the heat transfer coefficients in nucleate boiling, where water was the cooling fluid, increased as the surface orientation was changed from a horizontal surface facing upwards to a vertical position. Nishikawa et al. (1983) investigated in detail the effect of surface orientation on nucleate pool-boiling heat transfer on a flat copper plate. The orientation was varied from 0° to 175° from the horizontal plane in which 0° denotes an upward facing horizontal surface. They were able to find that at low heat fluxes the boiling curves shifted upwards with increase of inclination angle from upward facing to downward facing. Ma et al. (1997) investigated the local convective heat transfer from a vertical heated surface to an obliquely impinging free-surface jet of transformer oil under single phase convection. The jet was rotated about the horizontal plane creating an orientation with respect to the impingement surface. The Reynolds number and inclination angle were ranged from 235-1745 and 90°-45°, respectively. A major finding was that as the obliqueness of impingement decreased an increasing asymmetry was observed in the Nusselt number profiles. The heat transfer rate on the side downstream of the maximum heat transfer point was found to be higher than that on the upstream. It clearly showed that with increasing jet incidence angles the heat transfer rate on the upstream direction side drops off more rapidly while that on the downstream tends to decline more slowly, which they explain is the result of higher flow rates on the upstream side. Literature Review 30 Tong (2002) studied the effect an oblique liquid air planar jet had on the associated heat transfer when in contact with an upward facing heated horizontal surface. The authors evaluated the Nusselt number across the entire heater surface at inclination angles from 45° to 90° in which the case of 90° has the jet positioned perpendicular to the impingement surface. Observations showed that for a uniform jet the peak Nusselt number increased as the angle of attack was shifted to 45°. The distribution of the Nusselt number and maximum pressure location along the impingement zone became more asymmetric and shifted upstream from the impingement point as the inclination decreased. They discovered that the Nusselt number was the greatest at the center of the jet. Roy et al. (2003) conducted a similar study with an air jet set obliquely to the impingement surface; however, the flat impingement surface was inclined rather than the planar nozzle. In this case, they were able to find a similar occurrence to Tong (2002) in the case of the local Nusselt number increasing with increased inclination angle, regardless of the Reynolds number. Beitelmal et al. (1998, 2000) performed experiments to determine the effect of the inclination of an impinging two-dimensional air jet on the heat transfer from a uniformly heated flat plate. The heater surface was made of stainless steel and the width of the planar jet corresponded to the width of the heater element. Inclination angle of the jet relative to the plate was in the range from normal to the surface at 90° to 40°, and the Reynolds number was in the range of 4,000-12,000, which corresponds to an exit jet velocity of 6.3 to 18.7m/s. Some important findings were that the maximum Nusselt number decreases from its value for normal impingement as the inclination angle is Literature Review 31 decreased. Another finding was that the higher percentage of heat transfer took place on the downhill side (i.e. in the direction the nozzle is aligned). Also, that a decrease in local heat transfer from the centreline of the jet takes place, which they attributed to the decrease in momentum of the wall jet, and that the maximum heat transfer point occurred at a position 1.5 jet diameters from the centreline of the jet. 2.6 Effect of Strip Speed Cooling of moving surfaces with liquid jets has been widely used in manufacturing processes such as metal rolling, continuous casting, and others since the mechanical properties of these materials are strongly affected by the rate at which the product is cooled. An understanding of relevant heat transfer mechanisms and their relationship to the moving surface as a parameter needs to be well understood to optimise specifications for runout table cooling system design. For instance, recent cooling technologies such as thermo-mechanical treatment (TMT), delivers shorter distances on. the runout table that the steel strip has to pass through, increasing rolling speeds, decreasing water consumption and adding to the ability to manufacture heavier strips (Hernandez, 1999; Chen; 1988, 1991; Han, 1991). The moving speed of the work piece can vary from 0.5 - 15m/s, while the jet velocity is in the range of 1.0 - 5.0m/s (Han, 1991). In the current study, a strip speed was varied from 0-1 m/s to simulate industrial conditions. The following section will discuss research that has been carried out for boiling heat transfer on moving surfaces. Han et al. (1991), and Chen et al. (1988, 1991) performed a number of cooling experiments on moving and stationary plates to observe the influence moving surfaces Literature Review 32 have on the heat transfer. A rectangular test plate moved at speeds from 0-1.5m/s while being impinged by a 4.76mm diameter circular nozzle where water jet temperatures ranged from 22-27°. The authors were able to show that for the moving case, the water film stretched in the moving direction which enhanced heat transfer downstream from the stagnation point. The cooling zone, found within the stagnation zone, formed into an elliptical shape. As the fluid flowed upstream of the stagnation point it quickly lost its momentum to the opposing viscous forces, while the flow downstream gained momentum due to the surface motion and propagated longer than in the stationary case. Except that the flow downstream had a ragged boundary due to increased turbulence. It was observed that as the plate moving speed increased the cooling length became shorter since the duration of the plate underneath the jet decreased. They showed some interesting findings in that the peak heat flux for the stationary plate was found to be higher than that in the moving plate because of the highly initial transient behaviour of cooling for the stationary case. As the plate speed increased the peak heat transfer coefficient decreased from 110 kW/m2K (u = Om/s) to 65 kW/m2K (u = 0.32m/s) to 50 kW/m2K(u= lm/s). In the case of peak heat fluxes, a value of 24x106 W/m2 was found for stationary plate which was higher than the maximum heat flux of 17xl06 W/m2 in the moving case. The effect of surface motion on forced convection heat transfer due to an impinging, laminar planar jet was studied by Zumbrunnen et al. (1989, 1990, 1991). They carried out mostly a numerical analysis by using the Navier-Stokes equations to determine the heat and mass transfer distributions for a moving flat surface in which both the effects of surface temperature and surface motion were considered. In one Literature Review 33 experimental study, however, Zumbrunnen et al. (1990) developed an experimental method for measuring heat transfer distributions on moving plates cooled by a downward planar liquid jet. They found that the peak heat transfer coefficient occurred on the downward side of the stagnation point where nucleate boiling is more effective. Numerical solutions showed that surface motion affected local heat transfer coefficients near the stagnation line only when the surface temperature varied spatially. Procedures and Methodology 34 Chapter 3 - Procedures and Methodology In the present study, a pilot-scale runout table (ROT) facility was used in the Advanced Metallurgical Processing Engineering Laboratory (AMPEL) at the University of British Columbia (UBC). The facility design enables cooling experiments to be carried out under stationary and moving plate conditions to better simulate industrial conditions during runout table cooling of steel. Zhang (2004) carried out an extension to the ROT design by adding a single bottom nozzle to allow for bottom jet cooling experiments. The current study used a modified version of the bottom jet test rig designed by Zhang (2004). 3.1 Experimental 3.1.1 R O T facility The ROT facility is primarily composed of a large furnace to heat the test plates up to a desired temperature, a conveyor-chain system to transport the test specimens from the furnace to a position above the bottom jet, and a water tower, which includes an overhead tank and a containment tank. The overhead tank has a capacity of 1350 litres (i.e. 1.5 x 1.5 x lm) and is situated at the top of a 6.5m water tower. A 30kW electric heater fixed inside the overhead tank provides a means of controlling the temperature of the water up to 95°C. The containment tank, which is situated at the base of the water tower, has a capacity of 1550 litres (i.e. 0.689 x 0.711 x 2.64m). The containment tank functions as a basin to catch and recycle water being ejected from the bottom jet, as well as, increases the Procedures and Methodology 35 pumping capacity of the ROT facility since an added reservoir of treated water is available. An industrial sized water pump (11.2 kW, 3600RPM, 6.62 1/s) provides a means of transporting water throughout the ROT system. The conveyor-chain system is mechanically powered by a hydraulic drive system which uses oil under pressure to transmit power. The gear pump supplies pressurised oil to the hydraulic motor which then provides rotational speed directly to the conveyor. The hydraulic motor can provide up to 12.8kW of power and a maximum torque of 90Nm, which translates into a maximum speed of 1020rpm. Other essential components on the ROT facility include water flow valves, flow metres, and a single bottom nozzle that can be shifted to varying nozzle inclination angles and stand-off distances. A schematic of the bottom jet cooling setup is shown in Figure 3-1, in which subset of Figure 3-1 illustrates the primary components of the bottom jet experimental system. The bottom jet system diverts water from the main water line to the bottom nozzle. The bottom nozzle was constructed from a commercial stainless steel tube with an approximate inner diameter of 19mm and a pipe length of 40cm. The nozzle pipe was connected to an adjustable 90°-elbow. A Lindberg/Blue M furnace (5.8kW, 240V, 25A) was used for stationary plate experiments to heat up the test samples to the desired temperature. The furnace is able to achieve a maximum heating temperature of 1200±5°C. The furnace cavity has dimensions of 584x30x57mm. Two valves were used in the bottom jet system. Valve#2 is upstream from the flow meter and is used to vary flow rate to a pre-determined value Procedures and Methodology 36 and valve#3 downstream stopped water flow until experiments were about to commence. The bottom jet flow rate was measured by an OMEGA FTB-905 turbine flow meter in which the flow through the bottom nozzle was controlled by a valve upstream from the flow meter. Bottom Jet Flow Meter Top Tank Valve#3 Containment Tank Moving Plate Furnace A A / A /4S Moving Plate On Conveyor System Figure 3-1: Schematic of the bottom jet cooling system. 3.1.2 Test Mater ia l Drawing Quality Special Killed (DQSK) steel plates supplied by Stelco Inc. were used in this study for stationary plate experiments in the as-hot rolled condition. Table 3-1 lists the composition of the material used in this experiment taken from (Hauksson, 2001). High Strength Low Alloy (HSLA) steel plates supplied by Dofasco Inc. in the as rolled condition were used as test samples for the moving plate experiments. The 6.65 mm thick plates were cut to the dimensions of 1200 x 430 mm. Procedures and Methodology 37 Table 3-1: Chemical composition of test material in wt% Alloying Element C Al Cr Mn Mo Ni N P S Si DQSK 0.06 0.04 _ 0.24 _ _ 0.0035 0.005 0.011 0.006 HSLA 0.051 0.04 0.04 1.29 0.011 0.013 0.0045 0.012 0.004 0.102 3.1.3 Thermocouple Instrumentation and Car r ie r Design A number of Type-K thermocouples (Omega INC-K-Mo- 1.6mm) were used in this study because of their ability to operate under high temperature conditions of up to 1150°C, and for their good weldability with HSLA steels. The thermocouple wires are made from a Chromel-Alumel (i.e. Ni-Cr, Ni-Al) metallic pair. These wires had an approximate diameter of 0.05mm in which the wires were isolated by alumina filler. The metallic sheath is made from Inconel, and has an outside diameter of 1.6mm. Thermocouple wires are normally welded together at the wire tips to form a bead, also known as a junction. A characteristic voltage, which is dependent on the dissimilar metal pairing, can be used to determine temperature conditions at the thermocouple junction. However, an intrinsic type thermocouple junction was employed in this study for the purpose of eliminating thermal mass and increasing data responsiveness of the thermocouple (Hauksson, 2001). A diagram of an internal thermocouple connected by an intrinsic junction can be seen in Figure 3-2. Thermocouples were welded to each test plate at positions approximately 1mm below the impingement surface in a flat bottom hole. This placement to the impingement surface was done to reduce thermal lag between the impingement surface and flat bottom hole, thus ensuring high data responsiveness. Procedures and Methodology 38 Test material Thermocouple wires Ceramic Tube Intrinsic Junction Impingement Surface Figure 3-2: Diagram of an internal thermocouple connected by an intrinsic junction. Each thermocouple hole had a diameter of 1.6mm, which corresponds similarly to the outside sheath diameter of the thermocouple. Figure 3-3 and Figure 3-4 show schematics of the thermocouple arrangement for stationary plate and moving plate experiments, respectively. 430 mm 16 mm 215 mm -32 mm 265 mm Reverse Flow 32 mm Lateral Flow 16 mm Forward Flow 48 mm Figure 3-3: Diagram of a stationary plate thermocouple configuration and designation of different flow directions. Procedures and Methodology 39 The minimum spacing between each successive thermocouple hole, measured from centreline to centreline, was approximately 16mm. This distance corresponds to the minimum distance between thermocouple holes before thermal distortions in the temperature field caused by each thermocouple hole begins to overlap one another (Caron et al., 2006). 1220 mm 64 mm 32 mm 16 mm — 457 mm Direction of Moving Strip 412 mm Figure 3-4: Diagram of a moving plate thermocouple configuration and designation of strip direction. Due to the hydrodynamic asymmetry caused by the inclined jet thermocouples were arranged in a linear sequence along a line rather than radially. Three flow regions were defined on the plate due to the asymmetry of the jet hydrodynamics; forward flow, reverse flow, and lateral flow. The forward flow direction is aligned in the same direction as the inclined jet. The reverse flow direction is opposite to the forward flow, Procedures and Methodology 40 and lateral flow is assumed to be symmetric on either side of the jet. Figure 3-5 shows these flow regions as well as other experimental nomenclature. Reverse Flow Direction Lateral Flow Direction Jet Centreline Forward Flow Direction Direction of Moving Plate Figure 3-5: Diagram of flow directions. Two upper carrier designs were used in this study because there was a great difference in plate dimensions between stationary and moving plate experiments. In both types of experiments the test plates were mounted to a carrier for the purposes of delivering the test plate from the furnace to the test jig or lower carrier, to ensure that the test plates were positioned properly above the nozzle, and to protect the thermocouple connections from being dislocated. An upper carrier design was adapted from one used by Zhang (2004) in the case of stationary plate experiments. This carrier was built from angle iron and a steel pipe welded together, in which the test plate was attached to the angles and the steel pipe Procedures and Methodology 41 provided a means of handling the plate. The welded angle section had approximate dimensions of 29.2 x 43.2cm and was sized to the specifications of a stationary plate. The steel pipe had an inner diameter, wall thickness, and pipe length of 19.05mm, 6.35mm, and 612mm, respectively. The lower carrier, which is connected to the conveyor-chain system, was positioned above the jet in the test section prior to the plate being removed from the furnace, which functioned as the test jig during the experiment. The upper carrier, shown in Figure 3-6, was placed on top of the lower carrier before the jet was turned on. T/C Wires out to Data Acquisition Steel Pipe Thermocouple Wires 5 1 Impingement Surface Angle Iron HSLA Steel Plate Figure 3-6: Schematic of the upper carrier for stationary plate. A schematic of the upper carrier for moving experiments is shown in Figure 3-7. The primary building materials used on this carrier were from angle iron making a rectangular frame to mount the test plate on and two steel pipes welded to the frame for the purpose of removing the plate from the furnace. The welded angle section had approximate dimensions of 50.8 x 132.1cm. Thermocouples were instrumented on the top surface of the plate and were not inserted within the steel pipe as in the stationary plate case, but were attached to the outside of steel pipe. Procedures and Methodology 42 T/C Wires out to Data Acquisition Steel Pipe HSLA Steel Plate Angle Iron z SEE Impingement Surface Figure 3-7: Schematic of the upper carrier for moving plate. 3.1.4 Data Acquisit ion Signals from all thermocouples were collected on a DaqBook/2000 Series data acquisition system. The flow meter data was collected by an I-NetlOO data acquisition board. All thermocouple and flow meter data was transferred to a PC by DASYLab 8.0 data acquisition software. In order to capture the potentially rapid changes of the temperature during cooling, a maximum data polling frequency was selected at 40 Hz. All thermocouple and flow meter data were recorded in both raw and filtered forms. However, the raw data was used in the analysis. 3.1.5 Test Procedure Preparation of the test plates were initiated by drilling threaded holes and through-holes that would later be used to anchor the thermocouples to the test plate and anchor the test plate to the upper carrier, respectively. The flat-bottom thermocouple holes were cleaned with an injection of methyl alcohol from a hypodermic needle followed by a blast from an air jet. The depth of each thermocouple hole was measured using a micrometer and a 1.6mm diameter pin of known length. Procedures and Methodology 43 Once instrumented with the 20 type-K thermocouples for stationary plate, or 23 thermocouples for each moving plate, the test plate was bolted to the upper carrier. Fibre insulation (Fibrefrax) was set in between the test plate and upper carrier before being bolted together for the purpose of limiting conductive heat transfer between the carrier and the plate during jet cooling. The thermocouples were un-sheathed to expose the thermocouple wires for welding. These wires were inserted into a ceramic tube to prevent contact between the wires and ensured that the thermocouple junction was made on the base of the flat-bottom hole. The internal thermocouple wires were spot-welded separately to the plate, each wire being separated by a distance of approximately 1mm. The thermocouples were bolted to the plate using the aforementioned threaded holes by 10-24NC bolts. This prevented dislocation of thermocouples during plate transport. On the opposite end of the thermocouple wire, the male adapter end was fixed to the upper carrier handles by a clamp. Then each thermocouple wire was connected to the data acquisition hardware before insertion into the furnace cavity. After the plate was inserted into the furnace cavity enriched with nitrogen the furnace temperature was raised to 930°C. After insertion of the test plate into the furnace, the jet nozzle was inclined to a pre-determined angle. The angle measurement was made by laying a flat plate across the channelling, which was pushed adjacent to the bottom nozzle, making a level platform to take vertical and diagonal measurements of the nozzle with respect to the plate. J In the case of stationary plate experiments, the lower carrier was placed in its proper position above the nozzle to ensure thai the centreline of the jet impinged on the central thermocouple position. Procedures and Methodology 44 The water was circulated throughout the ROT system to make the water temperature uniform. Circulation was done by partially restricting water flow through the bottom nozzle, and passing the remaining flow to the upper tank for heating. The heated water was gravity fed down to the containment tank. The water temperature was maintained at 30°C by a temperature gauge and a 30kW heater in the upper tank, and another temperature gauge in the lower containment tank. Flow rate had to be set to a pre-determined value just before removal of the test plate from the furnace. The main line that connects the top tank with the containment tank was closed with valve#l to isolate water flow through the bottom jet system. The bottom jet flow rate was measured by an OMEGA FTB-905 turbine flow meter and values were acquired and displayed on a computer using DASYLab 8.0 software. Flow through the bottom nozzle was controlled. Once the valve was correctly adjusted and the pre-determined flow rate was reached valve#3, found downstream from the flow meter, was closed to restrict water flow. This was done to ensure that the stationary plate was positioned properly before the jet contacted the plate. Approximately 3 hours of heating time was required to reach the desired plate temperature and to ensure uniformity of temperature across the plate. The data acquisition system was started immediately before the plate was removed from the furnace. For stationary experiments, the test plate was manually removed from the Lindberg/Blue M furnace and placed on the test jig above the bottom nozzle. Once the plate was in position and the plate temperatures were at approximately 820°C the bottom jet was turned on by fully opening valve #3. Procedures and Methodology 45 For the moving experiments, the test plate was pulled from the furnace pocket onto the lower carrier. The conveyor-chain system was activated on a pre-determined voltage, which corresponds to a set strip speed. Data was collected until the temperature of the plate reached a temperature of approximately 150°C. Each experiment took approximately 1-5 minutes depending on flow rate, inclination angle, and strip speed. All cooling jet experiments were filmed on a Panasonic PV-GS500 video recorder with a shutter frequency at approximately 30 Hz on mini-DV cassette tapes. Each still picture could record up to 4.0 Mega-pixels. Computer software converted the video film into digital video files which were divided further into snapshot images for the purpose of estimating the initial size of the impingement zone and the progression of the wetting front. Still photographs were taken during experimentation from a Canon EOS 5D digital camera to compliment video images. 3.2 Image analysis A high resolution video camera was used to record videos for each test. Once the video file was digitised using computer software, the video file could be isolated frame by frame. Figure 3-8 shows a typical video snapshot taken after start of jet impingement. Procedures and Methodology 46 Figure 3-8: Images taken from stationary experiment #5 at 2s after jet impingement (i.e. water temperature = 30°C, flow rate = 45 l/min, and inclination angle = 20°). Initially, the entire bottom surface is bright red just before the jet impinges the bottom surface. Immediately, after impingement an area on the plate just above the jet turns black, which corresponds to a zone that is highly wetted. The bright red zone is assumed to be experiencing air convection and radiation. At the interface between the dark and bright zones, the wetting front occurs. Matrox Inspector 4.0 image analysis software is used to measure the size of the wetted zone and the position of the wetting front with respect to the centreline of the jet in the lateral flow direction for each frame. This data coupled with temperature and heat flux data, the measurements obtained using image analysis generate a database used for the analysis of the behaviour of the wetted region during jet impingement (Prodanovic et al., 2006). As well, it provides a means of comparing numerical solutions to what is happening physically. Procedures and Methodology 47 3.3 Numerical Analysis An inverse heat conduction (IHC) model developed previously at UBC (Zhang, 2004) was used to analyze the data and calculate surface heat fluxes and temperatures at known locations along the forward, reverse and lateral directions along the test specimen based on the measured internal thermal history. This model was modified to account for the asymmetry in the forward and reverse flow directions, as well as, remove the TC hole geometry in the model by using an "Equivalent Depth" technique, which will be described in more detail later. 3.3.1 I H C P Mode l The IHC model used in the current study was developed by Zhang (2004) to analyse data collected for his bottom jet experiments. Solution of the inverse heat conduction problem consists of two steps: solving the direct conduction problem and solving the inverse problem. The direct problem is solved by taking assumed heat fluxes to calculate temperatures at all the measured locations. In solving the inverse step, the calculated and the measured temperatures are compared and their differences are used for calculating the increments of the assumed heat fluxes. These two steps are repeated until the difference between the calculated and the measured heat fluxes are smaller than a pre-set convergence criterion (Zhang, 2004). The heat transfer in the plate during bottom jet impingement cooling with inclined jets is complicated by the fact that in the forward and reverse flow direction there is asymmetric cooling while in the lateral direction symmetry can be assumed. As a result two separate geometries were used: one to solve for the heat Procedures and Methodology 48 fluxes along the forward and reverse axis (x-axis) and the other to solve for the heat fluxes along the lateral flow axis (y-axis). The sensitivity of the IHC code was evaluated by varying a number of input parameters to see the effect on the final solution. The verification of the sensitivity of the inverse heat conduction model (IHCM) was performed by comparing the calculated heat fluxes with those heat fluxes generated from modified input parameters. The four evaluated input parameters studied for model sensitivity are as follows: mesh size, variance of conductivity (i.e. both temperature dependent and non-temperature dependent), initial temperature, and depth variance. The density of the mesh used in the analysis was varied such that close to the impingement surface, where the steepest thermal gradient would occur, very fine elements were used whereas further away a coarser mesh could be tolerated. Mesh sensitivity was performed to ensure the mesh density was not influencing the results predicted by the IHC model (Appendix A). Since the jet is inclined, heat conduction is not symmetric about the central thermocouple (i.e. TCI). The 3 flow directions are selected due to the asymmetry of jet hydrodynamics. A planar (i.e. rectangular coordinate) moving boundary code is applied to all three flow regions (i.e. forward flow, reverse flow, and lateral flow) separately. Figures 3-9 a) and b) illustrate the methodology behind the moving boundary code applied to the forward and reverse flow region, and the lateral flow region, respectively. Procedures and Methodology 49 Wetting Front Different Heat Flux Conditions on Impingement Surface a) Forward and Reverse flow direction Wetting Front b) Lateral flow direction Figure 3-8: Schematic of IHC boundary conditions in a) Forward and Reverse flow direction, and b) Lateral flow direction Procedures and Methodology 50 The boundary condition on the top side of the test plate is considered to be under air convection and radiation during the entire cooling jet process. The boundary condition of the left surface is assumed to be adiabatic since the central thermocouple (TCI) experiences the largest heat fluxes thus no heat is conducted through. The boundary condition of the right side of the test plate is assumed to be conduction through more test material based on Fourier's law. The governing equation describing the flow of heat in the forward and reverse flow axes, and lateral flow axis, are shown in Equation 3-1 a) and b), respectively: dx d_ dy k — k^ + — dz d k3T dz = / * , f . (3-1.) f dT\ „ dT dz\ The boundary conditions are defined according to the four boundaries identified on Figures 3-7 a) and b). At the centreline of the sample, x = 0 or y = 0 at all z-positions (i.e. z = 0 to 7.6mm), heat flow across this boundary is assumed to be zero as shown in equations 3-2 a) and b). 57/ dx = 0 (forward and reverse flow axes) (3-2a) lx=0 which is due to the highest heat flux being present at x = 0. = 0 (lateral flow axis) (3-2b) 57/ dy IjteO which is due to symmetry. On the un-impinged part of the sample, z = 7.6mm, and at all x and y positions, heat is lost due to radiation and air convection as shown in equation 3-3: Procedures and Methodology 51 dn , =h(T-TJ = q" (3-3) z=1.6mm where h is estimated according to equation 3-4 to 3-5 and n is the outward pointing normal to the boundary. h = hr+hc (3-4) hr=crSBsCT2+T>)(T + TJ (3-5) Since the amount of heat transfer by convection in stagnant air is less than radiation at elevated temperatures (i.e. hr » hc), the effects of air convection are neglected. At the right hand side (x = 128mm or y = 80mm) for all z locations, a semi-infinite solid is assumed such that thermal conduction is present as shown in equations 3-6 a) and b). 8T\ -k dx dy = q" (forward and reverse flow axes) (3-6a) x=\2Smm = q" (lateral flow axis) (3-6b) y=80mm At the impingement surface, z = 0mm, and at all x and y positions = h(T-TJ = q" (3-7) -k—| dn z=Q mm Where q " is the applied heat flux in W/m2 and n is the outward pointing normal to the boundary. The heat flux at this boundary is quite complicated and changes not only spatial but with time. The wetting front times, as will be discussed in Section 3.3.2, were used as queues to tell the IHC code to change boundary conditions on the impingement surface from nucleate boiling to radiation and air convection depending on the rate of progression of the wetting front. The region on the impingement surface that the wetting front occupies sees a step function shift from one boundary condition to another. Thus, Procedures and Methodology 52 the primary inputs into the IHC code are thermocouple data arranged in sequence starting from TCI for each flow region plus wetting front times. The heat fluxes on the impingement surface boundary are unknown and have to be estimated. The initial condition used in the model is shown in equation 3-8 a) and b). T(x,z,t0) = T0(x,z) (3-8a) T(y,z,t0) = T0(y,z) (3-8b) The initial temperature distribution in the direction of plate thickness is assumed constant. The thermo-physical properties of the tested material used in the analysis are shown in Table 3-2. Table 3-2: Thermo-physical properties of the plate material Property Value Reference Conductivity: k 60.571-0.03849T [W/m°C] Hauksson (2001) Density: p 7800 [kg / mJ] Hauksson (2001) Specific heat: cp 470 [J / kg °C] Hauksson (2001) The presence of a thermocouple hole causes a local disturbance of the temperature field in a sample and leads to a discrepancy between the thermal history measured by the thermocouple and the thermal history in an area without a thermocouple present (Caron et al., 2006). This phenomenon will lead to errors in the surface heat flux calculated using an inverse heat conduction model which relies on the accurate acquired thermal history data. A technique called the Equivalent Depth (ED) technique was proposed by Caron et al. (2006) to compensate for the disturbance of the temperature field due to the presence of a thermocouple hole. It has been employed in this study to account for the thermal disturbance that a thermocouple and a thermocouple hole make on the local temperature Procedures and Methodology 53 field. Validation of the ED technique is shown in Appendix B. This technique involves off-setting the true depth (TD) location of a thermocouple with a more accurate equivalent depth (ED) based on the specific radius (TR) and at the specific depth (TD) of the thermocouple, as shown in equation 3-9. 77? ED = TD (3-9) 3.3.2 Wetting Front Propagation Due to the nature of bottom jet cooling as well as the hydrodynamic asymmetry of an inclined jet, taking video images was not a feasible option to measure the progression of the wetting front. Instead, measured interior cooling curves were used to estimate the wetting front since this process could be applied to every thermocouple regardless of its position. The video imaging method was carried out in the lateral flow direction for all experiments to confirm the results made by the numerical method. In both the video image method and numerical method, the radius of the wetted zone was measured as a function of time from the instant the jet impacted the test plate. Knowing the precise times at which the wetting front passes over each thermocouple position is an integral part in characterizing the boiling regimes during the jet impingement cooling process that is used in the numerical analysis. The wetting front times were determined based on the 2nd order derivative of the recorded thermocouple data. Zhang (2004) applied this approach for each recorded thermocouple measurement in the IHC code. The associated times at which the wetting front passed over each Procedures and Methodology 54 thermocouple position was used as a queue to tell the IHC code when to change boundary conditions. In the current study, the times at which the wetting front reached each thermocouple position in each of the three flow regions (i.e. forward flow, reverse flow, lateral flow) were plotted. It was observed that the first three thermocouples were normally within the initial wetted zone. As well, a delay in the progression of the wetting front occurred for approximately 0.5s after jet impingement. Figure 3-10 is a typical wetting front progression plot taken from experiment#4 (i.e. flow rate = 35 1/min, inclination angle = 20°) in the forward flow direction. 250 i 0 V - , • , n , • , 0 20 40 60 80 100 Time [s] Figure 3-9: Wetting front progression taken from stationary experiment #4 (i.e. flow rate = 35 1/min, and inclination angle = 20°) in the forward flow direction The discrete points represent the actual wetting front times estimated based on the maximum in the 2nd order derivative of the thermocouple data for each thermocouple location (Zhang, 2004). As can be seen, the wetting front progressed in a linear manner Procedures and Methodology 55 in the forward flow direction with a velocity of 1.76mm/s. This velocity was used in the IHC model to specify the rate of progression of the BC on the quench surface for stationary experiments. This analysis was done in each direction (forward, reverse and lateral) as well as for each test. 3.3.3 E r r o r In Table 3-2 shows estimated errors associated with some of the measured experimental variables. The errors in water flow rate, water temperature, and inclination angle measurements were estimated from experience in measuring the quantities. Table 3-3: Measurement errors Quantity Error Temperature [Loveday etal., 1982] ±2°C Temp < 277°C ±0.75% Temp>277°C Flow Rate ± 1 l/min Inclination Angle ± 0.5° Thermocouple Depth ± 0.003mm Thermocouple Location ± 0.1mm (relative to centre of plate) Water Temperature ± 1°C The jet from the nozzle to the impingement surface was observed to be transient in nature, which made it difficult to hit the central thermocouple location precisely. When positioning the lower carrier below the nozzle required accurate measurements in which some observational error was present. The error of having the centreline of the jet impinge on the central thermocouple was estimated to be approximately +/- 5mm. Results and Discussion 56 Chapter 4 - Results and Discussion A combined total of 15 stationary and moving tests were conducted out on UBC's pilot-scale ROT facility. The first 11 tests were stationary and the last 4 tests were moving. The effect of inclination angle and flow rate, and in the case of moving experiments, the additional effect of strip speed, was studied. The experimental matrices are shown in tables 4-1 and 4-2, for stationary and moving plate experiments, respectively. Work by Zhang (2004) on bottom jet cooling of stationary samples using a 0° inclination was added for comparison purposes. Table 4-1: Experimental matrix for stationary plate Flow Rate / Inclination Ang le 10° 20° 30° 35 1/min 1 4 7 45 1/min 2 5 8 55 1/min 3 6 9 Table 4-2: Experimental matrix for moving p Strip S p e e d / Inclination Angle 0° 20° 0.3 m/s 10 11 0.9 m/s 12 13 ate Results and Discussion 57 4.1 Typical Test Data 4.1.1 Cooling Curves The cooling curves are a representation of the change in internal temperatures with time. Internal temperatures were measured for all stationary and moving plate experiments at approximately 1mm below the impingement surface. These temperatures were recorded from times immediately after the test plate was removed from the furnace, at approximately 920°C, until the plates cooled down to room temperature. The experimental start temperature is considered the temperature at which the water jet first impinged on the plate, which took place at approximately 800±20°C. Figures 4-1 show the cooling curves of a typical stationary plate taken in the forward flow (a), reverse flow (b), and lateral flow (c) direction. As can be seen in Figure 4.1, the thermocouple temperatures were uniform before the water jet impinged on the test surface. The thermocouple positions that were very close to the centreline of the jet saw substantial cooling very soon after the jet impinged on the plate. It was observed that the further thermocouple positions are away from the centreline 6f~the jet the longer it took to see a rapid drop in temperature. As expected, there was an asymmetry in the cooling pattern in the forward, reverse and lateral directions. Results from the measured thermal history have shown that the 3 aforementioned flow directions cool down differently for all cases when a jet is inclined with the impingement surface. Results and Discussion 58 20 40 60 T ime [s] 80 100 a) Forward flow direction 0mm +32mm +64mm • +96mm 20 40 60 T i m e [s] 80 100 b) Reverse flow direction Results and Discussion 59 900 -i-800 -j 700 -o p 600 -3 500 -ra k_ 0) 400 -Q.E 300 -H 200 -100 -— 0mm — +16mm +32mm — +48mm — +64mm 0 20 40 60 80 100 Time [s] c) Lateral flow direction Figure 4-1: Measured thermal history for positions away from the centre of the jet for experiment #4 (35 l/min flow rate, and 20° inclination) showing: a) Forward flow direction, b) Reverse flow direction, and c) Lateral flow direction. Positions at 0mm, 16mm, and 32mm from the impingement point in the forward flow direction show very similar cooling trends indicating that these thermocouples lie within a zone of relatively uniform heat flux. This zone is sometimes denoted as the impingement zone (Hauksson, 2001); however, for the purpose of clarity the term initial wetted zone of the jet will be used. The size of the initial wetted zone in the forward flow direction lies between 32mm to 48mm for this experiment. Findings from the reverse flow direction in Figures 4-lb) show that no thermocouples other than the central thermocouple are within the initial wetted zone. However, from Figure 4-lc) the thermocouple at an approximate distance of 16mm from the impingement point in the lateral flow direction is seen to cool within 0.5s after experimental start. Results and Discussion 60 As can be seen in Figure 4-1, some TC's outside the initial wetted zone experience a rapid decrease in temperature followed by a recovery and then a much lower drop in temperature (48 mm in the forward flow direction and 32 mm in the reverse and lateral flow directions). Video imagery revealed that when water first contacts the hot plate it covers most of the surface and then recedes back to the initial wetted zone as the vapour layer builds-up outside in the un-wetted zone. The water is diverted away from the plate surface by the stable vapour-film barrier on the surface of the test plate. The subsequent rise in temperature is due to the relatively ineffective cooling in the un-wetted zone due to a combination of air convection, radiative cooling, and film boiling (Mitsutake etal., 2001). Thermocouple positions greater than 32mm from the impingement point in the forward flow direction show a gradual decline in temperature after recovery depending on the distance between the respective thermocouple and the impingement point. After a given time, a sharp decrease in temperature is noted for each thermocouple outside the wetted zone, which can be attributed to the progression of the wetting front moving across the plate at those thermocouple locations (Filipovic et al., 1995). 4.1.2 Boil ing Curves The boiling curve is represented as the variation in heat flux versus surface temperature, which can be taken for each measurement location. These curves can facilitate the comparison of heat fluxes between various thermocouple locations within the same experiment as well as similar locations for different experiments. Figure 4-2 shows the calculated heat fluxes as a function of time in the forward flow direction. Results and Discussion 61 9.00E+06 7.00E+06 H 8.00E+06 A Omm (TC1) +16mm (TC2) +32mm (TC3) E 6.00E+06 A -••-+64mm (TC5) — +80mm (TC6) +96mm (TC7) 0.00E+00 0 5 10 15 20 Time [s] 25 Figure 4-2: Calculated heat flux versus time in the forward flow direction for experiment #9 (55 l/min flow rate and 30° inclination) As can be seen in Figure 4-2, the boiling curves at thermocouple locations TCI (Omm), TC2 (16mm), and TC3 (32mm), show almost identical characteristics suggesting the size of the initial wetted region to be greater than 32mm in the forward flow direction. The maximum heat fluxes occur at similar times for all three locations; however, the maximum heat flux for TC3 is approximately 2.5MW/m less than that of TCI. It should be mentioned that the highest heat flux was always obtained from TCI regardless of inclination angle or flow rate. The maximum heat flux was observed to decrease with increasing distance from the impingement point (i.e. centerline of the jet). Maximum heat fluxes from thermocouples TC5, TC6, and TC7, which are found outside the initial wetted zone, show the same trend in heat flux with respect to distance from centerline. However, the difference in maximum heat flux between each respective thermocouple location is approximately 0.25MW/m . A leveling of maximum heat flux was observed Results and Discussion 62 for positions far from the impingement point such that they became approximately constant. Figure 4-3 shows calculated boiling curves in the forward (a), reverse (b) and lateral (c) flow directions with respect to surface temperature. It is observed that the shapes of boiling curves at locations within the impingement zone are similar to each other. Also boiling curves at locations found outside the initial wetted zone are similar in shape. It should be noted the point of maximum heat flux was observed to shift to lower surface temperatures at locations far away from the impingement point, from approximately 350°C at the impingement point to approximately 250°C at locations TC5, TC6, and TC7 (i.e. +64mm, 80mm, and 96mm from jet centerline). The rapid rise in heat flux for locations close to the impingement point are shown in the wide spacing between each calculated point. At locations farther away the cooling is more gradual and the heat flux slowly decreases beyond the point of maximum heat flux. As can be observed in Figure 4-3 b) and c), all thermocouple positions except for TCI show similar trends in maximum heat flux, which range between 3.0-4.0 MW/m2. These thermocouple locations fall outside the jet impingement zone (i.e. the initial wetted zone). In the case of reverse flow, as seen in Figure 4-3 c), the distance between thermocouple location TCI4 and TCI6 is 64mm. Even though the maximum heat fluxes are reached at times approximately 25s apart, both curves are quite similar in shape. Results and Discussion 63 J 3.00E+06 2.00E+06 1.00E+06 0.00E+00 200 400 600 800 Surface Temperature [°C] 1000 a) Forward flow direction 9.00E+06 8.00E+06 7.00E+06 l7 6.00E+06 5.00E+06 X LI 4.00E+06 ® 3.00E+06 2.00E+06 1.00E+06 0.00E+00 0mm (TC1) +32mm (TC14) +96mm (TC16) 200 400 600 800 Surface Temperature [°C] 1000 b) Reverse flow direction Results and Discussion 64 9.00E+06 8.00E+06 7.00E+06 -\ 6.00E+06 5.00E+06 A x 3 4.00E+06 * 3.00E+06 2.00E+06 A 1.00E+06 0.00E+00 • Omm (TC1) * +32mm (TC18) — +47mm (TC19) * +64mm (TC20) 200 400 600 800 Surface Temperature [°C] 1000 c) Lateral flow direction Figure 4-3 Calculated boiling curves for stationary experiment #9 (55 l/min flow rate and 30° inclination) showing: a) Forward flow direction, b) Reverse flow direction and c) Lateral flow direction. Figure 4-4, which shows a heat flux curve taken from experiment #9, is provided to further illustrate the transitions in a typical boiling curve. It has been taken from a position outside the impingement zone at 32mm in the lateral flow direction. The vertical lines represent the change between different heat transfer regimes. Two heat flux peaks are observed in which one occurs at approximately 580°C and another at approximately 200°C. The maximum heat flux occurs at approximately 200°C surface temperature. It can be considered as the demarcation between transition and nucleate boiling where transition boiling exists on the side of higher surface temperatures. Results and Discussion 65 Nucleate Transition Film Initial Boiling Boiling Boiling Cooling 4 . 5 0 E + 0 6 -, : : : : 9 0 0 Surface Temperature [°C] Figure 4-4: Boiling curve outside impingement zone at +32mm from jet centreline in the lateral flow direction from stationary experiment #9 (45 1/min and 20°) The heat flux peak at approximately 580°C (i.e. Point A) is considered the time at which the IHC code has stabilised during the initial cooling stage into the stable film boiling regime. It should be mentioned that this stage is not present in the typical boiling curve, as seen in Figure 2-7. This point corresponds to a relative maximum vapour layer thickness and it decreases with decreasing wall superheat (Hall et al., 2001). Thus radiation and forced convection through the vapour barrier are the governing modes of heat transfer. This could explain the relatively high heat transfer in the first moments of cooling with a leveling in cooling at point A. A stable vapour barrier is assumed to exist as the surface temperature drops from 580°C to 520°C. However, the effectiveness of radiation and convection are reduced, Results and Discussion 66 resulting in the decrease in heat flux. During this time the wetted zone increases in size. The wetting front (i.e. the demarcation between the wetted and un-wetted zones) approaches thermocouple locations outside the initial wetted zone. At the transition between region 2 and 3, the boiling curve exhibits a pronounced local minimum at a surface temperature of approximately 520°C. This point, is assumed to demarcate a shift from stable film boiling to transition boiling since the surface is unable to supply enough * heat to support a stable film barrier. After the wetting front reached the measuring location, the heat flux increased monotonically to the critical heat flux and then dropped in the nucleate boiling regime. These trends are typical of all tests carried out. Further along the curve, the heat flux increases to a maximum heat flux with decreasing surface temperature to approximately 200°C. Once past the critical heat flux, nucleate boiling then single-phase convection occurs with decreasing surface temperature. Similar trends were found in other tests under different inclination angles and jet flow rate conditions. A transition denoted as point B has been established in the literature (Ochi et al., 1984; Ishigai et al. 1978, Hall et al., 2001) which has been attributed to a transition from nucleate boiling to single-phase convection. Although, surface temperature at which this transition occurs is much higher than the inception temperature for nucleate boiling, and there is no obvious change in slope of the boiling curve in the vicinity of this shoulder. Results and Discussion 67 4.1.3 Visual Observations The plate was bright red in colour once removed from the furnace just before jet impingement commenced. No obvious scale was observed on the plate surface. Figure 4-5 a) shows the water jet impinging on the bottom of the plate at 0.0333s after the start of the experiment. Observations show that a zone immediately underneath the jet turns black. In comparison to the cooling curve and heat flux data, the dark zone on the figure represents a significant temperature drop and shows very large heat fluxes immediately after jet impingement, which is consistent with the wetted region. The size of the initial black area depended on the water flow rate and inclination angle. Shortly after jet impingement water covered nearly the entire surface until a stable vapour barrier was established. A grey ring appeared beyond the black zone in which partial wetting is assumed to occur as shown in Figure 4-5 b). Further along the plate water was forced from the plate surface by gravity and by the formation of a stable vapour film barrier on the surface. As the water jet cooling continued, the black zone grew in the radial direction. As shown in Figure 4-5 c), the black zone is apparently bigger at 1.066s after impingement than at 0.267s. Results and Discussion 68 i * X T i Flow a) 0.0333s after jet impingement b) 0.267s after jet impingement Results and Discussion 69 c) 1.333s after jet impingement Figure 4-5: Video images taken during experiment # 5 (45 1/min and 20° inclination) for: a) 0.0333s after jet impingement, b) 0.267s after jet impingement and c) 1.333s after jet impingement. To visualise the effects of different flow configurations, calculated surface temperature contours were developed based on the calculated surface temperatures from each thermocouple location. Figure 4-6 shows the surface temperature contours after 2s for a bottom jet inclination from 0° to 30°. As can be seen in Figure 4-6 a), taken at a 0° inclination the data is symmetric about the impingement point. In contrast, as seen in Figure 4-6 d), the jet at 30° is highly asymmetric with the largest spatial temperature drop occurring in the forward flow region. The blue zone on each curve represents an area on the plate that has seen significant cooling. It can be seen that the size of the blue zone decreases as inclination angle is increased. It is largest for the case of 0° inclination and Results and Discussion 70 smallest in the case of 30° inclination, suggesting that 0° inclination is the most efficient at removing heat. 0 50 100 -100 -50 0 Along Length of Plate from Jet Centreline [mm] 8 <a i s 2 50 o -50 £-100 < -100 -50 0 50 100 -100 -50 0 Along Length of Plate from Jet Centreline [mm] Figure 4-6: Effect of inclination angle on the symmetry of the surface temperature profile at 2s after turning on the jet at 45 1/min and a) 0° inclination, b) 10° inclination, c) 20° inclination, and d) 30° inclination Results and Discussion 71 4.1.4 The Wetting Front The progression of the wetting front was calculated numerically based on the method outlined in section 3.3.2. The wetting front is determined only at the thermocouple locations where internal temperatures were measured. Figure 4-7 shows the radial progression of the wetting front based on video imagery (i.e. discrete points), as well as, numerical calculations of the 2 n d order derivative of the measured thermal history (i.e. straight line). Only the 2 n d order derivative measurements were utilised in the IHC code and in the remainder of the wetting front analysis. 10 -0 -I 1 1 . 1 1 0 10 20 30 40 50 60 Time [s] Figure 4-7: Comparison of wetting front estimates between 2 n d order derivative method and video images from stationary experiment #4 (35 1/min and 20°) in the lateral flow direction As can be seen, the results obtained by the two methods match well, and are consistent with previous work (Zhang, 2004). Results from the 2 n d order derivative calculations have indicated that on the plate surface at lateral locations 16mm from the centreline of the jet was wetted immediately upon jet impingement. The slope of the Results and Discussion 72 curve, which is taken after the wetting front starts to progress, represents the rate of the progression of the wetting front with respect to time. These slopes were calculated for every stationary plate experiment in each of the three flow directions. Figures 4-8 show the influences of the flow rate on the progression of the wetting front in the forward flow (a), lateral flow (b), and reverse flow (c) directions, respectively. As observed in these figures, increasing the flow rate causes the wetting front velocity to increase in all flow directions. These figures also show the influences of inclination angle on the progression of the wetting front. As seen in Figure 4-8 a), the wetting front velocity in the forward flow direction is enhanced as inclination angle is increased since there is an increase in water flow in that direction. Conversely, as seen in Figures 4-8 b) and 4-8 c), the wetting front velocity decreases in the lateral and reverse flow directions as inclination angle increases due to a decrease in flow rate in those directions. Similar trends are seen for different flow rates. Results and Discussion 73 40 45 50 Flow Rate [l/min] 55 60 30 35 a) Forward flow direction 40 45 50 Flow Rate [l/min] 55 60 b) Lateral flow direction Results and Discussion 74 4 T 3.5 -(A Flow Rate [1/min] c) Reverse flow direction Figure 4-8 Effect of inclination angle and flow rate on the wetting front progression for: a) Forward flow direction, b) Lateral flow direction and c) Reverse flow direction. Once the wetting front velocity was known quantitatively for each flow direction, a 2D spatial and temporal progression of this front could be plotted across the plate surface. This was achieved by plotting the spatial position of the wetting front in the forward, reverse, and lateral directions in their respective axis at specific times. In the off-axis quadrants it was assumed that the wetting front had an elliptical profile. Figure 4-9 shows the 2D wetting front contours and the influences of inclination angle on the progression of the wetting front. In general, decreasing inclination angle decreased the progression of the wetting front in the forward flow direction; however, it was enhanced in the lateral and reverse flow directions. At an inclination angle of 0° the plot is circular since an axi-symmetric condition applies. Increased jet flow rate was also shown to enhance the velocity of the wetting front in all flow directions. Results and Discussion 75 E E c c a> O E o -89-60 0s - Experimental Start - - - 2 s 4s -• -8s -••-16s * -go m a. •a -60 •] -89-120 Length of Plate from Centreline [mm] a) 0° (Zhang, 2004) E E c o u E o * -80 2 re a. •a 5 -89--60 1-40'. 0 -40' -60 A - 8 9 -0s - Experimental Start - - - 2 s 4s - — 8s -•-16s /4.0 .' 60 /'80 100 120 Length of Plate from Centreline [mm] b) 10° (stationary experiment#2) Results and Discussion 76 E E c <u o E * -80 3 to a. •a -89--40 -60--69-— 0s - Experimental Start - 2 s •• 4s - 8 s -16s 60 80 /' 100 120 Length of Plate from Centreline [mm] c) 20° (stationary experiment#5) E E o c c c o E o •fc -80 o « Q. -86- -0s - Experimental Start -2s 4s T3 -60 4 0 / / .60 / 80 100 120 -60 H -89-Length of Plate from Centreline [mm] d) 30° (stationary experiment#8) Figure 4-9: 2D wetting front progression over time taken from stationary experiments at 45 1/min flow rate and showing a) 0°, b) 10°, c) 20° and d) 30°. Results and Discussion 77 Validation of the elliptical model used in the wetting front analysis was carried out using off-axis thermocouple positions. Details of this analysis are shown in Appendix C. Table 4-3 shows the size of the initial wetted zones for all experimental conditions. As can be seen, increasing inclination angle, as well as, flow rate tends to increase the initial size of the wetted zone before wetting front progression begins. However, in the case of 35 l/min, as inclination angle is increased the initial wetted zone decreases since the was jet barely able to reach the plate surface. Table 4-3: Area of initial wetted region (mm2) as a function of inclination angle and water flow rate. Inclination 0° 10° 20° 30° angle/Flow rate (Zhang, 2004) (l/min) 35 ~ 1045 647 512 45 1588 1456 1646 1840 55 — 1766 1825 2010 The size of the wetted zone was plotted versus time is shown in Figure 4-10 for all inclination angles with a flow rate of 45 l/min. As can be seen, initially the 30° inclination angle has the largest wetted zone. However, by 8s after the start of wetting front progression, 0° inclination has surpassed the other inclination angles in wetting zone size. Knowing that the highest heat fluxes occur within the wetted zone, this suggests that 0° inclination is the most efficient at removing heat. Figure 4-10: Size of wetted zone with respect to time for stationary experiments with 451/min flow rate Results and Discussion 79 4.1.5 Quantif ication of Heat Extract ion In order to calculate the overall heat extraction, thus the optimal flow configuration, it was necessary to determine discrete values of the heat flux spatially across the plate at each time step. To determine these discrete values at a given time step the following procedure was employed: 1) Determination of the heat fluxes along the x and y axes at each thermocouple location using the IHC model as described earlier based on the measured thermal history 2) Overlay of the wetting front contours on a nodal grid as shown by the black curves in Figure 4-11. Each curve was positioned so that it intersected a thermocouple location (denoted as X's) along the x-axis (i.e. the forward flow direction) in the forward section of the plate. In the reverse section of the plate, each curve was positioned so that it intersected a thermocouple location along the y-axis (i.e. the lateral flow direction). 3) Determination of heat flux values at each nodal position (black dot) via an interpolation technique which is described below. 4) Integration of these values spatially and temporally to quantify the overall heat extraction for each cooling experiment. As shown in Figure 4-11, a grid of discrete nodal points at 8mm intervals were used to calculate heat fluxes and surface temperatures at selected times. In combined total for both sections of the plate, 29x29 points were used along the x and y axes to cover an area of ~0.05m2. The heat flux curves were established from the wetting front Results and Discussion 80 contours, as discussed earlier. The heat fluxes across each heat flux curve were approximately constant, and variance of heat fluxes was found to be small. 120 120 Length of Plate from Centreline [mm] (X-axis) Figure 4-11: Schematic showing heat flux curves overlaid on nodal grid based on wetting front contours in forward flow section. Data from stationary experiment #4 (35 1/min, 20° inclination) The first step involved determining the heat fluxes at each discrete point by a linear interpolation method using the surrounding heat flux curves as a framework. A typical calculation is shown in Figure 4-12 and involves the following: a) Determination of heat flux values along each curve on either side of the discrete point at a given time, b) Determination of the radial distances from the centre to the discrete point (R*) and centre to each heat flux curve (Ri, R2), and c) Calculation of the incident angle (QDP) between the forward flow (x) axis to the discrete point. Results and Discussion 81 Impingement; Point Heat flux curves X - a x i s Forward f low Figure 4-12: Schematic indicating technique used to estimate the heat flux (QDP) at a discrete nodal point between two heat flux curves at a given time, t The following equations were applied to define the heat flux value for each discrete point (QDP)-where 90 - 1 90 -9DP = tan 1 - 1 Kx) (4-1) (4-2) (4-3) QDP is the heat flux value at the discrete point, Qi and Q2 are the heat flux values of the surrounding curves, R* is the radial distance of the discrete point from the centre (origin) of the plate, and Ri and R2 are the radial distances of the heat flux curve from the center point, 6DP is the incident angle between the discrete point and the forward flow axis. Results and Discussion 82 Figure 4-13 shows the calculated heat fluxes in the forward flow section of a typical stationary experiment after 7.5s of cooling. Length of Plate from Centreline [mm] Length or Plate from Centreline [mm] Length of Plate from Centreline [mm] Figure 4-13: Heat flux contours [W/m2] taken in forward flow section at 2, 5, and 7.5s into stationary experiment #4 (35 1/min flow rate and 20° inclination angle) The linear interpolation method misses the sharp transition associated with the progression of the wetting front. To account for this discrepancy, the wetting front was obtained spatially based on the calculated heat flux plots versus time. For example, Figure 4-14 shows the position of the wetting front at 20s into cooling in the forward section for stationary experiment#4. As can be seen, the wetting front is between TC5 and TC6. Figure 4-15 illustrates the method employed in determining the heat flux spatial distribution of the wetting front. In situations were the wetting front lay between two heat flux contours the following technique was used. Results and Discussion 83 Length of Plate from Centreline [mm] (X-axis) Figure 4-14: Schematic showing heat flux contours with position of wetting front at 20s into cooling in forward flow section. Data from stationary experiment #4 (35 1/min, 20° inclination) 4.00E+06 Polled Heat Fluxes Figure 4-15: Weighted average heat fluxes with wetting front time from stationary experiment#4 (35 1/min, 20° inclination) r Results and Discussion 84 Since the wetting front progresses across the plate, it will occupy different positions at any given time. As shown in Figure 4-15, at 20s the wetting front for stationary experiment #4 (35 l/min, 20° inclination) is positioned between TC5 and TC6. In general, the heat flux curves were calculated using a weighted average from the calculated heat fluxes at two thermocouple positions. In this example, TC5 and TC6 were used. Each heat flux versus time curve can be considered associated with a measurement location in between the two thermocouple locations. All heat flux values on Figure 4-15 that occurred at 20s on all curves are polled and used in determining the spatial distribution of the wetting front. The result is seen in Figure 4-16, which shows the wetting front heat fluxes spatially distributed non-dimensionally between the two thermocouple positions. Heat fluxes were fixed with a lower bound cut off of 750KW/m2 for all wetting front contours. 4.00E+06 3.50E+06 3.00E+06 "E" 2.50E+06 3 2.00E+06 8 1.50E+06 x 1.00E+06 5.00E+05 0.00E+00 0 T C 5 0.2 0.4 0.6 Non-Dimensional Distance 0.8 T C 6 Figure 4-16: Characteristic spatial distribution of heat fluxes on the wetting front. Results and Discussion 85 From Figure 4-16, locations found before TC5 would lie in the wetted zone where nucleate boiling and single phase convection regions occur. Beyond TC6 would lie in the un-wetted zone with air convection and radiation being the primary modes of heat transfer. Once these wetting front heat fluxes were known spatially, then those discrete nodal points which fell beneath the wetting front were removed from the overall heat flux calculation. As can be seen in Figure 4-14, the heat flux curves for TC5 and TC6 form an elliptical ring surrounding the wetting front contour in the forward flow section. Those discrete points which fell between the elliptical ring would be removed from the overall heat flux calculation. The heat fluxes obtained from the wetting front spatial distribution are then superimposed over the linearly interpolated heat flux map to obtain the more accurate representation of the overall heat fluxes for that instant in time. Figure 4-17 illustrates the technique used to calculate the overall heat removal effectiveness. Each point on the curve represents the total calculated heat flux across the plate at an instant in time. This was calculated as described previously.. The overall heat removal can be calculated by the total area underneath curve by adding each quadrilateral segment that is formed between two instances in time. Between the discrete calculated points it was assumed that a linear interpolation was adequate. The integrated value is then multiplied by the total surface area defined in Figure 4-11 (i.e. grid size which corresponds to ~0.05m ) to derive the overall effectiveness at heat dissipation in Joules. Results and Discussion 86 1.20E+07 T 2 4.00E+06 -> O 2.00E+06 -j I 0.00E+00 / ' ' I ' 1—'—I 1 ' 1 1—, 1 0 2 4 6 . 8 10 12 Time [s] Figure 4-17: Variation in overall calculated heat flux across an area of the plate as a function of time. Data illustrated was taken from stationary experiment#3 (55 l/min, 10° inclination). 4.2 Effect of Inclination Angle 4.2.1 Boiling Curves This research indicated that the inclination angle o f the bottom jet had a significant impact on the heat transfer. Figure 4-18 a) shows the effect of inclination angle on the boiling curves within the initial wetted zone. A s can be seen, as the inclination angle was reduced the heat transfer rate within the initially wetted zone increased. Similar trends are observed regardless of flow rate. Figure 4-18 b) shows the effect o f inclination angle outside the initially wetted zone in the forward flow direction. A s can be seen, an enhancement of heat transfer occurs when inclination angle is increased due an increase of fluid momentum in that flow direction. Results and Discussion 87 9.00E+06 8.00E+06 7.00E+06 - 0° (TC2) 10° (TC2) - - - 20° (TC2) - - 30° (TC2) 0 100 200 300 400 500 600 700 800 900 Surface Temperature [°C] a) Initial wetted zone (16mm from the centre) at 45 1/min 9.00E+06 8.00E+06 7.00E+06 6.00E+06 i . 5.00E+06 i l 4.00E+06 4-1 re X 3.00E+06 2.00E+06 1.00E+06 0.00E+00 - - 20° (TC5) - - - 30° (TC5) X X» •> • \ » . .X-0 100 200 300 400 500 600 700 800 900 Surface Temperature [°C] b) Outside initial wetted zone (64mm from the centre) at 55 1/min Figure 4-18: Effect of inclination angle in forward flow direction showing a) initial wetted zone (16mm from the centre) and b) outside initial wetted zone (64mm from the centre). Results and Discussion 88 Figures 4-19 a) and b) show the effects of inclination angle in the lateral flow direction within and outside the initial wetted zone, respectively. Results for the reverse flow direction are not displayed since they exhibit similar trends to the lateral flow direction. It can be seen that reducing inclination angle increases heat transfer in both figures. A decreased inclination angle will result in an increase in water flow in the lateral direction leading to increases in heat fluxes. Considering conservation of mass, a decrease in flow in the forward flow direction will have a tendency to increase flow in the lateral and reverse directions. 9.00E+06 8.00E+06 A 7.00E+06 A 6.00E+06 A i . 5.00E+06 -x 3 C 4.00E+06 A ra X 3.00E+06 -2.00E+06 A 1.00E+06 A 0.00E+00 0 100 200 300 400 500 600 Surface Temperature [°C] 700 800 900 a) Initial wetted zone (16mm from the centre) at 45 l/min Results and Discussion 89 9.00E+06 8.00E+06 7.00E+06 6.00E+06 i . 5.00E+06 H x E 4.00E+06 re I 3.00E+06 2.00E+06 -I 1.00E+06 0.00E+00 - -10° (TC18) - - - 20° (TC18) - • 30° (TC18) V \ J>--S \ \ \ \ r7 \ \ \ \ 0 100 200 300 400 500 600 700 800 900 Surface Temperature [°C] b) Outside initial wetted zone (32mm from the centre) at 55 1/min Figure 4-19: Effect of inclination angle in the lateral flow direction showing a) within the initial wetted zone (16mm from the centre) and b) outside the initial wetted zone (32mm from the centre). 4.2.2 Overall Energy Dissipation The overall energy dissipation reflects the overall energy extracted from the unit area of the plate during 10s of jet impingement. As illustrated in Figure 4-20, the overall energy dissipation is highest for a perpendicular jet. It decreases slightly as inclination is increased. At 30° inclination a significant deviation occurs, due to the jet barely reaching the plate. This is especially evident at the lowest flow rates. It should be mentioned that though the area of the initial wetted zone of an inclined jet is larger than a perpendicular jet, overall heat transfer is larger for the perpendicular jet. The probable cause is that an inclined jet enhances heat transfer in one flow direction but reduces it in the other two. As well, a perpendicular jet impinges the impingement surface with a higher velocity. Results and Discussion 90 This has been found to act like added subcooling since higher impact velocity increases local pressure near to the impingement point, thus increasing saturation temperature. Up to 20° there is a slight decrease in energy dissipation as a function of inclination angle. With increasing inclination, there is a considerable drop-off due to the large angle between the jet with respect to the bottom of the plate that the jet is not able to effectively reach it. The amount of this decrease lessens as the flow rate is increased. 4.00E+05 3.50E+05 A LU O >> " 3.00E+05 -| 2.50E+05 -g 2.00E+05 -«0 Q_ o 1.50E+05 4 1.00E+05 0 5 10 15 20 25 30 35 Inclination Angle [°] Figure 4-20: Overall energy dissipation as a function of inclination angle and water flow rate. Results and Discussion 91 4.3 Effect of Water Flow Rate 4.3.1 Boi l ing Curves As shown in Figures 4-21 through 4-23, an increase of jet velocity enhances heat transfer regardless of inclination angle or flow direction. Particularly, the region close to the impingement point shows that maximum heat fluxes are more pronounced. It was observed that higher flow rates were observed to yield higher heat fluxes. The surface temperature linked with the maximum heat flux shifted to higher temperatures with increasing flow rate, which is consistent with previous findings (Ishigai et al., 1978; Kumagai et al., 1995b; Robidou et al., 2002). 9.00E+06 X 3.00E+06 -_, 6.00E+06 -E I, 5.00E+06 -4.00E+06 \ x 2.00E+06 j 0.00E+00 7.00E+06 H 1.00E+06 4 8.00E+06 j 0 100 200 300 400 500 600 Surface Temperature [°C] 700 800 900 a) Initial wetted zone (16mm from the centre) Results and Discussion 92 9.00E+06 8.00E+06 7.00E+06 ^ 6.00E+06 E I, 5.00E+06 x 4.00E+06 3.00E+06 2.00E+06 1.00E+06 0.00E+00 • - - 35 l/min (TC8) — 45 l/min (TC8) 55 l/min (TC8) 100 200 300 ' 400 500 600 Surface Temperature [°C] 700 800 900 b) Outside initial wetted zone (112 mm from the centre) Figure 4-21: Effect of flow rate for a jet with a 10° inclination angle in the forward flow direction showing a) initial wetted zone (16mm from the centre) and b) outside initial wetted zone (112mm from the centre). 9.00E+06 8.00E+06 7.00E+06 „ 6.00E+06 E ^, 5.00E+06 x 3 4.00E+06 a 1 3.00E+06 2.00E+06 1.00E+06 - - - 35 l/min (TC4) — 45 l/min (TC4) / - . N . \ *. \ A \ • • s •••— , v. N - - - - - - . \ 100 200 300 400 500 600 700 800 Surface Temperature [°C] 900 Figure 4-22: Effect of flow rate for a jet with a 30° inclination angle in the forward flow direction in a region close to the initial wetted region (48mm from centre). Results and Discussion 93 9.00E+06 8.00E+06 7.00E+06 -I , 5 - 6 .00E+06 E I. 5.00E+06 -| x 3 - - - 35 1/min (TC14) - -45 l/min (TC14) 55 l/min (TC14) 0 100 200 300 400 500 600 700 800 900 Surface Temperature [°C] Figure 4-23: Effect of flow rate at 20° inclination in reverse flow direction outside the initial wetted zone (32mm from the centre). 4.3.2 Overall Energy Dissipation Figure 4-24 shows the overall energy dissipation with respect to flow rate. As can be seen as the flow rate increases the overall energy dissipation increases. The effect of inclination angle is most pronounced at the lower flow rates. At the higher flow rates the difference between the inclination angles is reduced. It is observed that a reduction in slope occurs along the curves from 45-55 1/min. This suggests that increasing flow rates might reach a plateau or maximum effectiveness beyond which an increase in flow rate yields no enhancement of heat transfer. Differences between inclination angles are very large at low flow rates. As flow rate increases the effect of inclination angle decreases and approaches that calculated for 0°. Results and Discussion 94 4.00E+05 T 3.50E+05 -— 3.00E+05 -g re a | 2.50E+05 -b ai gj 2.00E+05 -UJ 1.50E+05 -1.00E+05 -25 35 45 55 65 Flow Rate [l/min] Figure 4-24: Overall energy dissipation versus flow rate 4.4 Effect of Sample Movement A total of twenty-three thermocouple measurements were taken for each of the four moving plate experiments. In order to facilitate analysis of boiling curves only those thermocouples that fell along the jet impingement line were used. A characteristic cooling curve of a thermocouple on the impingement line is shown in Figure 4-25, which is taken from moving experiment #2 (45 1/min flow rate, 0° inclination, 1 m/s strip). It should be noted that a data acquisition frequency of approximately 34 Hz was utilised. Each pass is revealed by a sharp drop in temperature. It should be noted that the largest temperature drop does not happen during the first pass. Rather, the largest temperature drops happen in the range of approximately 400-600°C. This is consistent with work done by Hernandez (1995, 1999) and Chen,et al. (1991). It is assumed that the initial Results and Discussion 95 surface temperature was elevated enough to reduced liquid-solid contact greatly. Thus, reducing the amount of heat dissipation since water has a higher capacity to remove heat than water vapour. As the surface temperature dropped, the surface could not supply enough heat energy to maintain a vapour-film barrier such that higher liquid-solid contact increased. This was observed visually during experimentation by the increased amount of steam produced. Another interesting phenomenon is the recovery of plate temperatures after the jet left the vicinity of the thermocouple location. Initially, recovery is not observed 1200 1000 A 0 A 1 1 I I , 1 1 0 100 200 300 400 500 600 700 Time [s] Figure 4-25: Characteristic cooling curve for moving strip for 10 passes from moving plate experiment (1 m/s strip speed, 0° inclination angle, 45 l/min flow rate) Figures 4-26 and 4-27 show the effects of strip motion with a vertical water jet measured from the same thermocouple location on two different moving plate Results and Discussion 96 experiments (i.e. 0.3m/s and l.Om/s). It is observed that jet cooling commences with a steep temperature drop when the thermocouple location is in close proximity to the impinging jet. Hernandez (1999) explained that this sharp drop occurred once the plate reached the countercurrent parallel flow region (i.e. before the jet is underneath a thermocouple location). This was followed by even steeper cooling in the impingement region. After leaving the impingement region, recovery of the surface occurs increasing the measured temperature. By this point air convection and radiative cooling are the primary modes of heat transfer. Interestingly, recovery is higher for the moving plate and happens sooner for the faster moving plate (i.e. l.Om/s) since the duration of the plate underneath the jet is less. Thus, the amount of energy removal is less, meaning more heat is available to compensate for heat loss. 450 400-k — 0.3 m/s — 1.0 m/s ,2 150 ] 100 -\ 50 H 0 0 20 40 60 80 Time [s] 100 120 140 160 Figure 4-26: Effect of strip speed (pass #5, flow rate = 45 1/min, and inclination angle = 0°) Results and Discussion 97 400 350 A — 0.3 m/s 1.0 m/s 100 A 50 A 0 0 20 40 60 80 Time [s] 100 120 140 160 Figure 4-27: Effect of strip speed (pass #6, flow rate = 45 l/min, and inclination angle = 0°) The heat fluxes calculated from the temperature measurements are presented versus time and versus surface temperature in Figure 4-28 and 4-29, respectively. It can be seen for the moving plate data that heat fluxes increase monotonically in the countercurrent region up until the plate reaches the impingement region where the heat flux reaches a maximum. After the jet has passed the thermocouple locations entering into the concurrent flow region, the heat transfer also decreases monotonically. The stationary plate experiment (i.e. 0° inclination) using a thermocouple location at stagnation is added for comparison. As can be seen, as the strip speed is reduced the heat fluxes increase slow at first then substantially for the stationary plate experiment. The measured boiling curves for the parallel flow region of the jets impinging on the moving plate are compared with some reported measurements for non-moving surfaces and with predictions by the transition boiling model. Results and Discussion 98 1.00E+07 T 9.00E+06 -8.00E+06 -7.00E+06 £ 6.00E+06 j§ 5.00E+06 LL | 4.00E+06 I 3.00E+06 -2.00E+06 -1.00E+06 -0.00E+00 -0 100 200 300 400 500 600 700 800 900 Surface Temperature [°C] Figure 4-28: Effect of strip speed on heat flux versus surface temperature (0° inclination angle and 45 1/min flow rate) The boiling curves are compared versus time in Figure 4-29, where maximum heat fluxes are linked by a line. The heat fluxes measured for the moving plate vary considerably for the each pass, even though they operate under the same experimental conditions (i.e. flow rate, strip speed, etc.), illustrating the effect of surface temperature. With respect to the effect of the strip motion on the boiling curves, the results presented here and those reported by other researchers (Filipovic et al., 1992, 1994; Chen et al., 1991) for boiling reveal that the strip motion decreases heat transfer. This is a result of a decreased amount of time the plate spends above the jet. Hernendez (1999) assumed that other processes were also dependent on the strip motion, such as the mixing of water, which affects the local subcooling, and the length of the different cooling zones. Results and Discussion 5.00E+06 4.50E+06 Time [s] Figure 4-29: Effect of strip speed on heat flux versus time Conclusions and Future Recommendations 100 Chapter 5 - Conclusions and Future Recommendations 5.1 Summary An experimental study of the boiling heat transfer during impingement of an inclined bottom water jet on a downward facing hot steel plate was conducted using UBC's pilot-scale ROT facility. The focus of the study was to observe and analyse the temperature changes during the cooling process. The effects of jet inclination angle and jet flow rate on heat transfer were investigated for stationary experiments, which ranged from 10° to 30° and 35 l/min to 55 l/min, respectively. Previous work that focused on 0° inclination was also included into the current study for comparison purposes. The experimental part of this research consisted of measuring the internal temperature change in instrumented steel samples at various locations in the forward, reverse and lateral axis positions as well as in some off-axis locations. The TC's were instrumented at approximately 1mm from the impingement surface. Intrinsic thermocouples were used to increase data responsiveness of the temperature measurements. An inverse heat conduction model, which employed the wetting front progression across the impingement surface, has been utilised. The model used a two dimensional, planar finite element method to solve the nonlinear transient heat conduction problem. Measured internal temperatures were employed in the inverse heat conduction model to generate surface heat fluxes and surface temperatures. All cooling tests were filmed and video was digitised into still-images for data analysis. Conclusions and Future Recommendations 101 Maximum energy dissipation calculations were used to compare overall effectiveness at heat removal between each experimental configuration and an optimal setup has been suggested. 5.2 Conclusions Video imagery and temperature data collected from experiments indicated that the entire cooling process of the plate consisted of a number of phases and were dependent on the position of the measurement location. Two distinct cooling zones were observed on the bottom surface of the plate during impingement: The central wetted zone with a distinct elliptical shape, and the non-wetted zone. The dynamic boundary between the two zones is called the wetting front. The initial wetted zone is formed immediately at the start of cooling. It will keep a fairly constant size for a certain period of time (from approximately 0-ls after jet impingement) until the wetting front starts progressing outward. A sharp decrease in temperature observed for each thermocouple can be attributed to the progression of the wetting front moving across the impingement surface. Calculated boiling curves exhibited different heat transfer regimes depending on the thermocouple measuring location with respect to the impingement point and with respect to time. At positions within the wetted zone, heat transfer is primarily by single phase convection and nucleate boiling. Some evidence also suggests that transition boiling is present during the initial stages of cooling. For all measurement locations within the initial wetted zone, the maximum heat flux took place at the same time, Conclusions and Future Recommendations 102 although the value of the maximum heat flux decreased with increasing distance from the impingement point. Immediately outside the wetted zone, transition and film boiling is the predominant mode of boiling heat transfer until the wetting front passes over those thermocouple locations. With increasing distances from the impingement point, the plate surface is untouched by water and heat transfer is by air convection and radiation. The point of maximum heat flux for thermocouple locations outside the initial wetted zone was observed to shift to a lower surfaces temperatures and lower heat flux with increased distance from the impingement point. The effect of varying inclination angle in the range of 10° to 30° on heat transfer from the test plate was investigated and compared with previous investigation of vertical bottom jets. The maximum heat flux was always attained for thermocouple locations at the impingement point with higher calculated heat fluxes when using lower inclination angles. A vertical jet (i.e. inclination angle 0°) was found to yield the highest cooling. However, this effect reversed outside the initial wetted zone in the forward flow direction in which higher inclination angle yielded higher heat fluxes. This is a result of an enhancement of water flow in that flow direction. In the reverse and lateral flow directions, however, the effect of increasing inclination angle resulted in a decrease in heat fluxes. Inclination angle was also found to affect the size of the initial wetted zone and the rate of progression of the wetting front. The size of the initial wetted zone increased and the wetting front progressed faster with increased inclination in the forward flow direction with the opposite effect in the reverse and lateral flow directions. Conclusions and Future Recommendations 103 The effect of flow rate in the range of 35 1/min to 55 1/min on heat transfer from the test plate was investigated. The maximum heat flux occurred at the impingement point but decreased rapidly at increasing distances from the impingement point until a levelling of heat fluxes occurred far outside the wetted zone. Jet velocity influenced the size of the initial wetted zone and the rate of wetting front progression. Flow rate was found to increase the size of the initial wetted zone and the rate of progression of the wetting front in all flow directions. Investigations into the overall energy dissipation indicated that the best flow configuration at removing heat was a vertical jet and at highest flow rates. The following conclusions can be made: 1) Inclination angle did have a significant effect on the heat transfer occurring on stationary hot steel plates. As the inclination angle increases the overall effectiveness of heat removal decreased. This effect was most pronounced at the lowest flow rates studied. Part of the reason for this is that as the inclination angle increased the ability of the water to reach the plate surface at low flow rates was compromised. The other noticeable effect was that as the inclination increased the degree of asymmetry in the temperature distribution between the forward, reverse and lateral axes also increased. However, these changes could be counterbalanced by the introduction of moving plate. As the inclination angle increased the wetting front velocity along the forward flow axis increased whereas in the reverse and lateral flow directions it decreased. These trends were consistent at all the flow rates studied during this research. However, results showed that the initial wetted zone was larger for higher inclination angles. Conclusions and Future Recommendations 104 2) Flow rate also was seen to have a significant effect on heat transfer occurring on stationary hot steel plates. As expected as the flow rate increased and so did the effectiveness of the heat transfer regardless of the inclination angle. 5.3 Recommendations for Future W o r k Heat transfer behaviour in jet impingement cooling is very complicated, which has been shown in the current study. Further work is required to improve our knowledge of heat transfer on the industrial runout table. Suggestions for future work include: • Focusing primarily on a moving plate condition.. To simulate industrial conditions better, the influences of plate speed, the effect of multiple jet lines, as well as, top and bottom jets working in tandem should be considered. Employing water curtains could facilitate investigations since temperature gradients along the width of the plate would be reduced. • Conducting a detailed analysis of the hydrodynamics of jet impingement, including the size of the initial wetted zone and the rate of the progression of the wetting front for different flow configurations. This could be a helpful method in investigating the effectiveness of heat removal. • Observing the influence df other parameters such as subcooling, nozzle to plate distance, and the effect of different steel grades should be studied on the boiling heat transfer that occurs. • Refining the current IHC code to account for heat generation due to phase transformation. Moreover, the code should be expanded to account for heat transfer in three dimensions. References 105 References Beitelmal, A.H., Saad, M.A., and Patel, CD. , 2000, Effect of inclination on the heat transfer between a flat surface and an impinging two-dimensional air jet, International Journal of Heat and Fluid Flow, 21: 156-163. Beitelmal, A.H., Saad, M.A., and Patel, CD., 1998, Heat transfer by a two-dimensional air jet impinging obliquely on a flat surface, ASME International Mechanical Engineering Congress and Exposition, 361: 219-227. Berenson, P.J., 1962, Experiments on pool-boiling heat transfer, Int. J. 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Ma, C.F., Zheng, Q., Sun, H., Wu, K., Gomi, T., and Webb, B.W., 1997, Local characteristics of impingement heat transfer with oblique round free-surface jets of large Prandtl number liquid, Int. J. Heat Mass Transfer, 40: 2249-2259. References 109 Marcus, B.D., 1963, Experiments on the mechanism of saturated nucleate pool boiling heat transfer, PhD Thesis, Cornell University. Meng, Q., 2002, Experimental study transient cooling of a hot steel plate by an impinging circular jet, MASc. Thesis, MMAT, UBC. Militzer, M., Hawbolt, E.B., Meadowcroft, T.R., 2000, Microstructural model for hot strip rolling of high-strength low-alloy steels, Metall Mat Trans A Phys Metall Mat Sci, 31A: 1247-1259. Mitsutake, Y., and Monde, M., 2001, Heat transfer during transient cooling of high temperature surface with an impingement jet, Heat and Mass Transfer, 37: 321-328. Monde, K. Kitajima, T. Inoue, and Y. Mitsutake, 1994, Critical heat flux in a forced convective subcooled boiling with an impinging jet, 10th International Heat Transfer Conference, paper 18-FB-19, 515-520. Mozumber, A.K., Monde, M., Woodfield, P.L., 2005, Delay of wetting propagation during bottom jet impingement quenching for a high temperature surface, Int. J. Heat Mass Transfer, 48: 5395-5407. Nishikawa, K., Uchida, S., Ohta, H., and Fujita, Y., 1983, Effect of heating surface orientation on nucleate boiling heat transfer, ASME/JSME Thermal Engineering Conf. Vol.1,129-136. Ochi, T., Nakanishi, S., and Kaji, M. et al., 1984, Cooling of a hot plate with an impinging circular water jet, Multiphase Flow and Heat Transfer III: part A: Fundaments, ed. By Veziroglu, T.N. et al., Elsevier Science Publishers, Amsterdam, B.V., 671-681. References 110 Patel, P., and Roy, S., 2003, Study of jet impingement heat transfer for varying fluid flow characteristics, Proceedings of the ASME Fluids Engineering Division Summer Meeting, 2: 197-204. Prieto, M.M., Ruiz, L.S. and Menendez, J.A., 2001, Thermal performance of numerical model of hot strip mill runout table, Ironmaking and Steelmaking, 28(5): 474-480. Prodanovic, V., Fraser, D., and Militzer, M., 2004, Simulation of runout table cooling by water jet impingement on moving plates-A novel experimental method, 2nd International Conference on Thermomechanical Processing of Steel, ed. M. Lamberights, Liege, Belgium, 25-32. Prodanovic, V., and Wells, M., 2006, Propagation of the wetting front during jet impingement boiling on a horizontal flat surface, ECI international Conference on boiling heat transfer, IN PRINT. Robidou, H., Auracher, H., and Gardin, P., 2002, Controlled cooling of a hot plate with a water jet, Experimental Thermal and Fluid Science, 36: 123-129. Robidou, H., Auracher, H., and Gardin, P., 2003, Local heat transfer from a hot plate to a water jet, Heat and Mass Transfer, 39: 861-867. Summerfields, S., 2004, Boiling heat transfer for an impinging jet on a hot plate and the development of a new correlation, MSc. Thesis, Mech. Eng., U of M. Tacke,G., Litzke,H., and Raquet, F., 1985, Investigations into the efficiency of cooling systems for wide-strip hot rolling mills and computer-aided control of strip cooling, Proc. Symp. on Accelerated Cooling of Steel, The Metallurgical Society of the AIME, Pittsburgh, PA, Aug. 19-21, 35-54. References 111 Tong, A.Y., 2002, Hydrodynamics and heat transfer of an oblique plane jet impinging onto a substrate, ASME International Mechanical Engineering Congress and Exposition, 1: 23-30. Wang, X., and Monde, M., 1997, Critical heat flux in forced convective subcooled boiling with a plane wall jet (effect of subcooling on CHF), Heat and Mass Transfer 33, 167-175. Webb, B.W., Ma, C.A., 1995, Single-phase liquid jet impingement heat transfer, Advances in Heat Transfer, 26: 105-217. Wolf, D.H., Incropera, F.P., and Viskanta, R., 1993, Jet impingement boiling, Advances in Heat Transfer, 23: 1-132. Woodfield, P.L., Monde, M., Mozumber, A.K., 2005, Observations of high temperature impinging jet boiling phenomena, Int. J. Heat Mass Transfer, 48: 2032-2041. Zhang, P., 2004, Study of boiling heat transfer on a stationary downward facing ' hot steel plate cooled by a circular water jet, MASc. Thesis, MMAT, UBC. Zumbrunnen, D.A., Viskanta, R., and Incropera, F.P., 1989, Effect of surface motion on forced convection film boiling heat transfer, ASME J. Heat Transfer, 111: 760-766. Zumbrunnen, D.A., Viskanta, R., and Incropera, F.P., 1990, Method and apparatus for measuring heat transfer distributions on moving and stationary plates cooled by planar liquid jets, Experimental thermal and fluid science, 3, 202. Zumbrunnen, D.A., Viskanta, R., and Incropera, F.P., 1991, Convective heat and mass transfer in the stagnation region of a laminar planar jet impinging on a moving surface, ASME J. Heat Transfer, 113: 563-570. References 112 Zumbrunnen, D.A., Viskanta, R., and Incropera, F.P., 1992, A laminar boundary layer model of heat transfer due to a non-uniform planar jet impinging on a moving plate, Warme- and Stoffubertragung, 27: 311-319. Appendix A 113 Appendix A - Sensitivity Analysis A - l Mesh Sensitivity Mesh size relates to the number of nodes used in the code. The influence of mesh size on the final solution was studied by independently varying the number of nodes perpendicular and parallel to the impingement surface. A 2D single-hole axisymmetric finite element IHC model was applied to see the effect of varying mesh size. The input temperature data used was taken from a thermocouple position at the stagnation point. The FEM simulations were first conducted to see the effect mesh variances above the thermocouple tips had on the generated heat fluxes. Four vertical mesh densities were selected at 5, 10, 25, and 50 number of nodes within the same 6mm physical space above the thermocouple tip. The number of nodal columns were varied from 5 to 100. In this instance, the difference between the obtained heat fluxes from all the different meshes can be considered to be independent of the mesh size, which is shown in Figure A - l . In the case of mesh variance below the thermocouple tip, the effect was found to vary considerably. Eight different vertical mesh densities were selected at 4, 5, 1.0, 15, 20, 25, 40 number of nodes within the same 1mm physical space occupied below the thermocouple tip. Heat flux convergence was reached at 25 nodes per 1mm spacing, in between the measurement location and the impingement surface, as shown in Figure A-2. These results helped determine the proper mesh size for all IHC calculations conducted. Appendix A 114 7.00E+06 1.00E+06 O.OOE+00 200 400 600 800 Surface Temperature [°C] 1000 F i g u r e A - l : Change i n heat f lux for mesh variance above thermocouple t ip 0.00E+00 D i r e c t i o n o f increasing nodal density 200 400 600 800 Surface Temperature [°C] 1000 F i g u r e A - 2 : Change i n heat f lux for mesh variance b e l o w thermocouple t ip Appendix A 115 A-2 Conductivity Sensitivity The material properties of the test plates are essential in calculating the heat fluxes and surface temperatures by using an inverse heat conduction method. Perhaps the most important property in this study is thermal conductivity since it has a large influence on the estimates of the generated heat fluxes and surface temperatures. The material properties of the test specimens have a profound influence on the calculated heat fluxes and surface temperatures. None more so than thermal conductivity, which is considered the primary property of the present study. Predicting the properties of HSLA steel such as thermal conductivity accurately is made difficult since these changes vary not only with temperature, but are dependent on the phase, and that the phase is dependent on the thermal history. When a solid to solid phase change occurs there is a release of latent heat as the plate passes through the phase change. However, this is well beyond the scope of this study. The code itself does not take into consideration the solid to solid state phase transformations that happen during cooling. For instance, HSLA steel is in the austenitic phase when in the maximum limits of the furnace heating, however, upon removal from the furnace and subsequent cooling, the plate goes through an austenite to ferrite phase transformation. The thermal conductivity was studied using a temperature dependent value and a non-temperature dependent value. The temperature dependent thermal conductivity was found by a linear regression for the temperature range studied of AISI 1008, which gave the following correlation of conductivity: k = 60.571 - 0.03849T [W/m°C] (Hauksson, 2001). The non-temperature dependent thermal conductivity value that was used was the Appendix A 116 intercept of the temperature dependent thermal conductivity correlation (i.e. 60.571 [W/m°C]). As seen in Figures A-3, the shape of the heat flux curves remained relatively similar as the temperature dependent conductivity was varied. 7.00E+06 i , 1000 Surface Temperature [°C] Figure A-3: Effect of thermally dependent conductivity on heat flux However, when comparison of the relative error to the 0% conductivity case was plotted versus surface temperature, as seen in Figure A-4, the relative error varied similarly to the variation in conductivity in most regions except above 800°C and below 200°C. Meaning that across most temperature regions concerned there is approximately a 1:1 relationship in conductivity variance and relative error associated with the generated heat fluxes. Figure A-5 illustrates that the largest relative error occurs near the start of jet impingement when the plate is going through a highly transitory state as boiling regime are just beginning to develop on the impingement surface. Appendix A 117 45 0 -I , , , , 1 0 200 400 600 800 1000 Temperature [°C] Figure A - 4 : Relative error in heat flux with respect to surface temperature 100 90 0 2 4 6 8 10 12 14 16 Time [s] Figure A - 5 : Relative error in heat flux with respect to time Appendix B 118 Appendix B - Validation of ED Technique Numerical tests were conducted to observe the effect the equivalent depth (ED) technique had on the generated heat fluxes. The following curves were generated from thermocouple data taken at the stagnation point from two different experiments using both planar (i.e. rectangular) and axi-symmetric (i.e. cylindrical) coordinate systems. As can be seen in the curves without the ED technique applied, the planar model tended to consistently under-predict the heat flux value in comparison to the axi-symmetric model. However, in both instances the thermocouple hole was considered as a material void. The ED technique assumes that there is no thermocouple hole, and compensates for this added thermal mass by offsetting the depth of thermocouple hole. In both planar and cylindrical coordinates with the applied ED technique the effects disappeared and the two curves overlapped. Thus, the error associated when generating heat fluxes can be reduced if not eliminated with the ED technique. This is shown in Figures B-l. This method has been applied in the current analysis. Appendix B 119 8.00E+06 7.00E+06 6.00E+06 5.00E+06 2 4.00E+06 g 3.00E+06 2.00E+06 1.00E+06 O.OOE+OO Axi bU_AXl P l a n e ED Plane L i X I J I 1 1 1 1 1 , . 100 200 300 400 500 600 700 Surface Temperature [°C] 800 900 Figure B - l : Equivalent depth technique validation of TCI for experiment^ (0° inclination and 30 1/min flow rate) data taken from (Zhang, 2004) 2.50E+06 900 Surface Temperature [°C] Figure B-2: Equivalent depth technique validation Appendix C 120 Appendix G - Validating Interpolation Technique with Off-Axis Locations C - l Comparison of Wetting Front Times in El l ipt ical Mode l Figures C-l demonstrates the validity of the elliptical model in describing the progression of the wetting front with an inclined jet. These figures are taken from high to low experimental conditions using off-axis thermocouple locations for comparison. For each stationary plate experiment, three thermocouple locations were positioned off-axis such that they were not located on any flow axis (e.g. forward flow axis). The temperature profiles shown are from those off-axis locations. A time estimate was calculated when the wetting front reached the off-axis thermocouple location using the 2D wetting front contours based on the elliptical model as discussed in the previous section. As can be seen, the off-axis thermocouple position experienced a dramatic shift in cooling close to the time the wetting front reached the thermocouple position. This is consistent with work done by Filipovic et al. (1995), which showed that the wetting front reached thermocouple locations at the same time as the thermocouple cooling curve saw a sharp shift in cooling. Additionally, the 2n d order derivative method of measured thermocouple data was applied in calculating the wetting front time. It compared very well to the wetting front times estimated from the 2D wetting front contours. The maximum observed error was found to be approximately 2 seconds; however, in most cases it was less than 1 second. From Figure C-l , which is taken from TC#13 of experiments (i.e. 10° inclination and 35 1/min flow rate), the difference in time between the 2D contour estimate and the 2nd order derivative method of the measured thermocouple data was 0.025s. Appendix C 121 900 0 5 10 15 20 25 30 Time [s] Figure C-l: Measured internal temperature curve with inclination angle of 10° and 35 l/min flow rate in forward flow direction C-2 Comparison of Surface Temperatures in El l ipt ical Mode l To verify the elliptical model further, surface temperatures from off-axis locations were compared with off-axis temperatures estimated using on-axis temperatures. For each stationary plate experiment, four on-axis thermocouple locations were used in estimating the temperature profile for the off-axis location. Figure C-2 illustrates a characteristic example of this comparison using internal thermocouple data. This estimate was based on the assumption that an elliptical temperature profile existed and interpolation between was valid. Very good agreement is observed in the case of stationary experiment #8 from TCI2 where in the high temperature region (>300°C) less than 10% relative error is shown. Appendix C 0 5 10 15 20 25 Time [s] Figure C-2: Comparison between internal measured and estimated off-axis internal temperature curves from experiment #8 (45 l/min flow rate and 30° inclination angle)
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A study of boiling heat transfer on a hot steel plate cooled by an inclined circular bottom water jet
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