Heat transfer engineering an international journal, tập 32, số 2, 2011

Heat Transfer Engineering, 32(2):87–89, 2011 Copyright C Taylor and Francis Group, LLC ISSN: 0145-7632 print / 1521-0537 online DOI: 10.1080/01457631003769070 editorial Selected Papers From the Sixth HEFAT Conference JOSUA P MEYER Department of Mechanical and Aeronautical Engineering, University of Pretoria, South Africa In 2002, the 1st International Conference on Heat Transfer, Fluid Mechanics, and Thermodynamics (HEFAT2002) was hosted in the Kruger National Park, South Africa In 2003, the 2nd conference (HEFAT2003) was hosted at the Victoria Falls, Zambia The 2004 conference (HEFAT2004) was in Cape Town and the 4th conference (HEFAT2005) took place in Cairo, while the 5th conference (HEFAT2007) was at Sun City The 6th International Conference on Heat Transfer, Fluid Mechanics and Thermodynamics (HEFAT2008) was held in Pretoria, South Africa, from 30 June to July 2008 It was part of the University of Pretoria’s 100-year celebrations with the theme “A century in the service of knowledge.” For this conference and proceedings, all papers were peerreviewed by at least two reviewers and almost 150 papers were accepted The review policy was that only original research papers that were recommended unconditionally by two reviewers who are distinguished subject specialists in the field of the relevant paper were accepted The papers were read in 17 parallel lecture sessions over a period of three days during which five keynote papers were presented The purpose of most conferences, including this one, was to provide a forum at which specialists in heat transfer, fluid mechanics, and thermodynamics from all corners of the globe could present the latest progress and developments in the field This not only allowed the dissemination of the state of the art, but also served as a catalyst for discussions on future directions and priorities in the areas of heat transfer, fluid mechanics, and thermodynamics The additional purpose of this conference was to introduce Africa to the rest of the world and to initiate collaboration in research The current edition of Heat Transfer Engineering, therefore, is a special issue covering the HEFAT2008 conference It contains nine papers that were nominated by the conference session chairs and co-chairs as the best papers from each session These papers dealt with several topics as summarized by the authors: • The first paper was on condensation heat transfer and pressure loss measurements of high- and low-pressure refrigerants flowing in a 0.96-mm single minichannel The refrigerants considered were R32 and R245fa, as they display a wide range of fluid properties and therefore they could be used for proper validation of predicting models The condensation tests were performed in a unique measuring test section, at around 40◦C, and the pressure drop tests were performed in adiabatic flow conditions to measure only the pressure losses due to friction The experimental heat transfer data were compared with predicting models to provide a guideline for the design of minichannel condensers It has also been found that the heat transfer coefficients were roughly the same for the two fluids at the same experimental conditions and the condensation was shear stress dominated for most of the data points However, the frictional pressure drop was significantly higher in the case of the low-pressure refrigerant, as would be expected • The second paper was on quantifying mixing in penetrative convection experiments where penetrative convection in a stable stratified fluid has been reproduced under laboratory conditions It was done by employing a tank filled with water and subjected to heating from below The purpose of the experiments was to predict the mixing layer growth as a function of Address correspondence to Prof Josua P Meyer, Department of Mechanical and Aeronautical Engineering, University of Pretoria, Pretoria, 0002, South Africa E-mail: jmeyer@up.ac.za 87 88 J P MEYER initial and boundary conditions and describing the outcome of a tracer dissolved in the fluid phase The equipment used made it possible to simultaneously provide temperatures inside the domain through thermocouples and Lagrangian particle trajectories by feature tracking The results demonstrate the validity of Deardorff mixed-layer similarity for the turbulent structure of the boundary layer Also, the comparison with similar experiments described in literature shows good agreement with measurements taken at both bench and real scale, signifying the legitimacy of the experimental work and applicability to the real atmospheric boundary layer and its monitoring for environmental purposes • The purpose of the third article was to determine experimentally the local stretching rate distribution along the limit methane/air and propane/air flames Particle image velocimetry measurements were used to obtain moving flame velocity fields in a standard flammability column and also to recognize the flame structures The methodology that was developed proved to be reliable and able to supply analyses with repeatable data From the experiments, it was possible to derive the flame stretching rate that causes its extinction in both mixtures • Because of the heat capacity of pressure vessel walls, the heat transfer from the compressed gas to the vessel wall strongly influences the temperature field of the gas Until now, no correlations were available for the heat transfer coefficient between the inflowing gas and inner surface of the vessel To develop such a correlation, in the fourth article computational fluid mechanics was used to determine the gas velocities at the vicinity of the inner surface of the vessel for a number of discrete surface elements A large number of numerical experiments show that there exists a unique relationship between the gas velocity at the inlet and the tangential fluid velocity at the vicinity of the inner surface of a pressure vessel Once this unique relationship is known, the complete velocity distribution at the vicinity of the inner surface can be determined from the inlet gas velocity The near-wall velocities at the outer limit of the boundary layer are substituted into the heat transfer correlation for external flow over flat plates The method was applied to the special case of filling a 70-MPa composite vessel for hydrogen fuel cell vehicles • In the fifth article an air-side data analysis method was developed for flat-tube heat exchangers under partially wet conditions It was done by making the simplification that combined, sensible, and latent heat transfer assumed that drainage paths developed such that, at steady state, water does not spread to noncondensing surfaces, which therefore remain dry The air dewpoint was compared with local fin-tip and fin-base temperatures, and a partially wet flat-tube heat exchanger was divided into fully wet, partially wet, and dryfin regions, which were subsequently analyzed as separate heat exchangers Using an enthalpy-based effectiveness NTU method, the average fin efficiency was calculated for each region, and the locations of region boundaries were determined heat transfer engineering iteratively The method was compared with experimental data of a flat-tube louver-fin heat exchanger under various latent loads • For temperature-dependent heat transfer coefficients and heat capacities, fast approximation methods were considered for the estimation of the effective overall heat transfer coefficient in the sixth paper The heat transfer coefficients were determined for two, three, or four reference temperatures For parallel and countercurrent flow, a known method was described, which used a hypothetical fluid temperature for the iteration-free consideration of variable heat capacities For the mixed–unmixed cross-flow, a previous method for temperature-dependent heat transfer coefficients was refined to allow also for variable heat capacities A new iterative fast design and rating method was developed for the mixed–mixed crossflow, which was a suitable model for special multipass shell-and-tube heat exchangers The accuracy of the methods was tested against numerical calculations with good results • The seventh paper is related to the operation of proton exchange membrane fuel cell stacks, which require careful thermal and water management for optimal performance Appropriate placement of cooling plates and appropriate cooling conditions are therefore essential To study the impact of these design parameters, a two-phase model accounting for the conservation of mass, momentum, species, energy, and charge, a phenomenological model for the membrane, and an agglomerate model for the catalyst layer were developed and solved The models were validated for a single cell, in terms of both the local and the global current density, and good agreement was found Four repetitive computational units were then identified for the number of single cells placed between the coolant plates varying from one to four cells The flow fields in the single cells and the cooling plates were of a net type • The thermodynamic stability of gas hydrates was investigated in the presence of electrolyte solutions in the eighth article The proposed model was based on the Van der Waals–Platteeuw model for gas hydrate equilibrium, and the Pitzer and Mayorga model was employed to calculate the water activity in electrolyte solutions Available values for the Pitzer and Mayorga model parameters were usually adjusted using experimental data at 25◦C, which was usually higher than the gas hydrate formation temperature In order to eliminate this problem, those adjustable parameters were re-optimized using experimental data from the literature at the lowest temperature In the case of mixed electrolyte solutions and without using any adjustable parameters, a mixing rule was proposed to estimate the water activity The new mixing rule was based on the ionic strength of the mixture and estimated the mixture water activity by using properties of the single electrolytes which constituted the mixture The results show the proposed model can calculate hydrate equilibrium conditions with good accuracy, especially at low concentrations, which is the case for most industrial applications vol 32 no 2011 J P MEYER • In the last article, the flow behavior within an interrupted fin design, the inclined louvered fin, was investigated experimentally through visualization and numerically through CFD simulation The inclined louvered fin was a hybrid of the offset strip fin and standard louvered fin, aimed at improved performance at low Reynolds numbers for compact heat exchangers The flow behaviors was studied in six geometrically different configurations over a range of Reynolds numbers and quantified using the concept of “fin angle alignment factor ζ,” which was related to the flow efficiency η in louvered fins The experimental data resulted in a discrete data set of local ζ values, which was used to validate the simulations Using these validated cases, it was shown that the graphical measurement method can be distorted by recirculation zones resulting in erroneous values Care should thus be taken when performing graphical measurement of the mean flow angle based on dye injection images The transition from steady laminar to unsteady flow in inclined louvered fins was geometrically triggered and occurred at lower Reynolds numbers compared heat transfer engineering 89 with slit fins and standard louvered fins This property can potentially be used to further improve on the performance of interrupted fin surfaces Josua P Meyer is a professor and chair of the School of Engineering and also head of the Department of Mechanical and Aeronautical Engineering of the University of Pretoria, South Africa He specializes in heat transfer, fluid mechanics, and thermodynamic aspects of heating, ventilation, and air conditioning He is the author and co-author of more than 250 articles, conference papers, and patents, and has received various prestigious awards for his research He is also a fellow or member of various professional institutes and societies and is regularly invited as a keynote speaker at local and international conferences He has received various teaching awards as lecturer of the year, and he has received two awards from the University of Pretoria as an exceptional achiever In 2006, he was evaluated by the National Research Foundation (NRF) as an established researcher who enjoys considerable international recognition for the high quality and impact of his recent research outputs He is an associate editor of Heat Transfer Engineering vol 32 no 2011 Heat Transfer Engineering, 32(2):90–98, 2011 Copyright C Taylor and Francis Group, LLC ISSN: 0145-7632 print / 1521-0537 online DOI: 10.1080/01457631003769104 Condensation Heat Transfer and Pressure Losses of High- and Low-Pressure Refrigerants Flowing in a Single Circular Minichannel ALBERTO CAVALLINI, STEFANO BORTOLIN, DAVIDE DEL COL, MARKO MATKOVIC, and LUISA ROSSETTO Dipartimento di Fisica Tecnica, University of Padova, Padova, Italy A 0.96 mm circular minichannel is used to measure both heat transfer coefficients during condensation and two-phase pressure losses of the refrigerants R32 and R245fa Test runs have been performed at around 40◦ C saturation temperature, corresponding to 24.8 bar saturation pressure for R32 and 2.5 bar saturation pressure for R245fa The pressure drop tests have been performed in adiabatic flow conditions, to measure only the pressure losses due to friction The heat transfer experimental data are compared against predicting models to provide a guideline for the design of minichannel condensers INTRODUCTION not been reported in the open literature Coleman and Garimella [1] reported flow patterns for R134a condensing inside horizontal tubes and square minichannels of hydraulic diameter ranging from to 4.9 mm At mass velocities G > 150 kg m−2 s−1 the authors observed annular, wavy, intermittent (slug, plug), and dispersed (bubble) flow patterns At hydraulic diameters Dh = mm the wavy regime was not observed, while at high flow rates and qualities annular film with mist core or mist flows were observed The hydraulic diameter has a substantial effect on flow transitions, but the tube shape was found to be less significant Several other authors have performed flow visualizations in minichannels, as reported in Cavallini et al [2], but no specific flow visualization with high-pressure fluids has been done Considering the pressure drop behavior of minichannels, very few data are available in the open literature regarding highpressure fluids In fact, most data refer to medium-pressure refrigerants, such as R134a Cavallini et al [3] measured pressure drops during adiabatic flows of R410A, R134a, and R236ea at 40◦ C inside a multiport minichannel having a square cross section with hydraulic diameter 1.4 mm, length 1.13 m, and with mass velocities ranging from 200 to 1400 kg m−2 s−1 The multiport minichannel tested is characterized by a square cross section and a low value of surface roughness (Ra = 0.08 µm and Rz = 0.43 µm), whose effect can thus be neglected The authors compared their data against models, either developed for conventional macrochannels or specifically developed A significant and still growing part of the engineering research community has been devoted to scaling down devices in the last few decades, while maintaining or improving their functionality The introduction of minichannels in the field of enhanced heat and mass transfer is surely one of those attempts As a result, compact heat exchangers and heat pipes are used in a wide variety of applications: from residential air conditioning to the spacecraft thermal control Growing interest for different solutions can also be found in electronic cooling, though these applications are less interesting from the condensation point of view due to its exothermal nature Compact elements work with small refrigerant charge and can usually withstand extremely high system pressures Two-phase flow in rough minichannels is affected by gravity, inertia, viscous shear, and surface tension forces These forces influence flow regimes, pressure drop, and heat transfer characteristics of minichannels Some researchers reported flow regimes during condensation of R134a in minichannels, but general flow regime maps have Address correspondence to Dr Davide Del Col, Dipartimento di Fisica Tecnica, University of Padova, Via Venezia 1, 35131 Padova, Italy E-mail: davide.delcol@unipd.it 90 A CAVALLINI ET AL for minichannels None of the models was able to predict frictional pressure drops of R410A, and many models were not able to predict R236ea trends However, better predictions were obtained for the frictional pressure drops of R134a A new model for the frictional pressure gradient valid for adiabatic flow or for flow during condensation of halogenated refrigerants inside minichannels was then suggested by Cavallini et al [4] As is the case for flow visualization and pressure drop, most of the heat transfer data available in the literature for condensation inside minichannels were measured with R134a and in most cases multiport channels have been used For multiport tubes, averaged values over a number of parallel channels are measured instead of in one single channel It is not an easy task to measure local heat transfer coefficients during condensation inside a single minichannel, and it is complicated as compared to the flow boiling case, where electrical heating can be adopted In this context, a new experimental apparatus for the measurement of the local heat transfer coefficients inside a single minichannel has been recently constructed at the University of Padova With this apparatus, condensation tests have been performed in a 0.96 mm diameter circular channel The fluids investigated in this study are single-component refrigerants R32 and R245fa, a high- and a low-pressure fluid, respectively Their main physical properties at 40◦ C compared to the corresponding values of R134a are shown in Table R32 is considered a higher pressure refrigerant compared to R134a; its vapor density exceeds R134a by 46%, while its liquid viscosity is significantly lower R245fa has an opposite behavior It has much lower vapor density and higher liquid viscosity Surface tension force of R32 is slightly lower than that of R134a (by 20%), while its liquid thermal conductivity is higher by 50% The surface tension force of R245fa is twice that of R134a and 2.7 times of R32 In practical applications, the use of a high-pressure refrigerant can mitigate a disadvantage of the high pressure drop with small channels R32 also has high thermal conductivity, which is favorable to high heat transfer coefficients during condensation On the contrary, R245fa displays much lower saturation pressure and can be used when looking for low system pressures Table Properties of saturated R32 and R245fa compared to R134a at 40◦ C Properties R245fa R32 R134a Liquid density [kg m−3] Vapor density [kg m−3] Liquid thermal conductivity [mW m−1 K−1] Vapor thermal conductivity [mW m−1 K−1] Liquid viscosity [µPa s] Vapor viscosity [µPa s] Surface tension [mN m−1] 1297.0 893.04 1146.7 14.08 73.27 50.08 85.42 114.58 74.72 15.10 18.65 15.45 336.50 94.88 161.45 10.76 14.44 12.37 12.12 4.47 6.13 Note Corresponding saturation pressure is equal to 24.8 bar for R32, 2.5 bar for R245fa, and 10.2 bar for R134a heat transfer engineering 91 Figure Experimental test rig (DESUP = desuperheater, MF = mechanical filter, HF = drier, PV = pressure vessel, CFM = Coriolis-effect mass flow meter, P = pressure transducer, T = temperature transducer, DP = differential pressure transducer) The reason for studying those two fluids is that they display a wide range of fluid properties and therefore they can be used for proper validation of predicting models CONDENSATION TESTS In order to investigate condensation heat transfer within a single minichannel, a unique measuring test section has been constructed [5] The test rig designed for heat transfer and pressure drop measurements during condensation is shown in Figure It consists of the primary refrigerant loop and four auxiliary loops The subcooled refrigerant is circulated through a filter drier, then passes into the gear pump that is coupled with a variable-speed electric motor It is then pumped through the Coriolis-effect mass flow meter into the evaporator, where the fluid is heated up, vaporized, and superheated At the evaporator exit, the state of the superheated vapor is determined from temperature and pressure The superheated vapor enters the test section, which is composed of two countercurrent heat exchangers The first heat exchanger of the test section (desuperheater) is used to cool down the fluid to saturation state before entering into the second heat exchanger, which is the actual measuring sector The saturation temperature is obtained from the pressure in the two adiabatic sectors upstream and downstream of the measuring sector There, the refrigerant temperature is also measured by means of adiabatic wall temperature measurements After the test section, the fluid is sent to the post-condenser, where it is condensed and subcooled The temperatures and flow rates of the secondary loops are controlled by a closed hot-water loop, two thermal baths, and vol 32 no 2011 92 A CAVALLINI ET AL Figure Schematic of the experimental test section (the refrigerant flows from left to right) an additional resistance heater arranged in series at the inlet of the desuperheater In this way, it is possible to control the temperatures of four different heat sinks or heat sources within the test rig A sketch of the test section is shown in Figure The measuring section is a commercial copper tube with inner diameter 0.96 mm Single-phase tests and forced convective condensation and flow boiling tests can be done with the present test facility The test section is a straight, single minichannel with two diabatic sectors (desuperheater and measuring section in Figure 2) divided by an adiabatic capillary tube The two diabatic sectors are made from an mm external diameter copper rod with a 0.9 mm internal bore The desuperheater has a length of 50 mm and the measuring section is 228.5 mm long The thickwalled tube (8 mm diameter) was machined externally and then closed with a plastic sheath in order to obtain a cooling water channel within the wall thickness The tortuous path of the secondary fluid enables good water mixing and thus allows precise local coolant temperature measurements In this test section, 15 T-thermocouples have been inserted into the water channel along the measuring sector (MS) with 15 mm step to obtain the coolant temperature profile The enhanced coolant heat transfer surface area moves the dominant thermal resistance from the external to the internal side; with this solution the internal thermal resistance (refrigerant to wall) is increased and thus the experimental heat transfer coefficient uncertainty due to the refrigerant-to-wall temperature difference is reduced In order to measure precise local heat transfer coefficient values, 15 T-type thermocouples have been inserted into the wall thickness, near the minichannel along the measuring sector, without having the thermocouple wires cross the coolant path Furthermore, thermocouples are used to measure the refrigerant temperature in the adiabatic segments by recording the external wall surface temperature of the stainless-steel capillary tubes at the two extremes of the measuring section When operating in condensation mode, the first diabatic sector works as a desuperheater To avoid large temperature gradients at the inlet of the measuring sector, the desuperheater is used to cool down the superheated refrigerant to the saturation state at the inlet of the measuring sector Vapor quality is calculated from the energy balance on the coolant side, and saturation conditions are checked using the adiabatic wall temperature and the pressure measurement in the adiabatic sectors heat transfer engineering The following three parameters are used for determination of the local heat transfer coefficient: the local heat flux, saturation temperature, and wall temperature The heat flux is determined from the temperature profile of the coolant in the measuring sector The wall temperature is directly measured along the test section and the saturation temperature is obtained from the pressure measured at inlet and outlet of the test section The coolant side temperature profile is obtained from the thermocouples inserted in the water channel along the measuring sector (Figure 3) The derivative of the temperature profile is proportional to the local heat flux: q (z) = −m˙ w · c pw dTw (z) · π · di dz (1) and it is associated to the local heat transfer coefficient: HTC(z) = q (z) (Tsat (z) − Twall (z)) (2) The wall temperature (Twall (z)) is measured locally with the thermocouples embedded in the wall Figure Temperature measurements within the single minichannel test section vol 32 no 2011 A CAVALLINI ET AL 25000 20000 15000 10000 5000 z (3) heat transfer engineering 0.0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 VAPOUR QUALITY [/] 0.9 1.0 Figure Heat transfer coefficient measured during condensation of R32 in the channel versus vapor quality R245fa, a reduction in vapor density will increase the vapor velocity in the channel A higher vapor velocity leads to higher interfacial shear stress On the other hand, it must be remembered that the liquid thermal conductivity of R32 exceeds by 34% that of R245fa In the case of R32 data (Figure 4), the experimental heat transfer coefficients measured at 100 kg m−2 s−1 and those at 200 kg m−2 s−1 are very close to each other, showing little effect of mass velocity at these conditions This overlapping of 12000 11000 10000 9000 8000 -1 -2 In the present technique, the dominant thermal resistance during the condensing process is on the refrigerant side, as shown in Figure This favors minimizing the experimental uncertainty associated with the determination of the heat transfer coefficient In fact, one contribution of the experimental heat transfer coefficient uncertainty is the saturation-to-wall temperature difference The other main uncertainty terms are associated to the heat flux, the mass flow rate, and the hydraulic diameter Since the heat flux is obtained from the temperature gradient of the water, this uncertainty contribution depends on the operating conditions, (mass flux and vapor quality), yielding higher uncertainty at lower mass fluxes At highest mass velocity, 1200 kg m−2 s−1, the total heat transfer coefficient uncertainty is below 5%, while at the lowest mass velocity, 100 kg m−2s−1, it ranges between 10% and 25% More details on the error analysis are reported in Matkovic et al [5] Prior to the tests, all instruments were carefully calibrated Besides, several tests have been run to verify that the nondependency of heat transfer coefficient on the conditions of the secondary fluid The local heat transfer coefficient has been measured during condensation of R32 and R245fa The R32 experiments have been performed for the entire range of vapor quality at 40◦ C saturation temperature and mass velocity ranging from 100 kg m−2s−1 up to 1200 kg m−2 s−1 The experimental heat transfer coefficients for condensation of R32 are shown in Figure As expected for forced convective condensation inside conventional pipes, the heat transfer coefficient increases with mass velocity and vapor quality It is worth remembering that the lower the mass velocity, the higher is the experimental uncertainty of the heat transfer coefficient, due to the low local heat flux The heat transfer coefficient measured for R245fa, with mass velocity ranging from 100 up to 500 kg m−2 s−1, is shown in Figure By comparing the values measured for R32 and R245fa, one can see that refrigerant R32 displays roughly the same or a slightly higher coefficient at the same mass velocity and vapor quality This may be surprising since, in the case of HTC [W m K ] π · di · q (z)dz x(z) = xin − m˙ r · h LG G1200 G1000 G800 G600 G400 G200 G100 -2 -1 HTC [W m K ] The local saturation temperature, Tsat (z), of the refrigerant along the channel is calculated from the measured pressure at inlet and outlet and a proper prediction of the pressure profile along the channel using the model by Cavallini et al [4] This calculation is obtained by estimating the pressure gradient The calculated pressure at outlet is checked against the measured pressure at channel outlet and the pressure gradient profile is multiplied by a constant factor to match the calculated and the measured pressure drop The saturation temperature (Tsat (z)) is then evaluated, from the pressure, using Refprop7 [6] By considering the conservation of energy in the sector, the coolant temperature change is directly associated to the corresponding enthalpy variation of the refrigerant Therefore, the local vapor quality is calculated as follows: 93 7000 6000 5000 G500 4000 G400 3000 G300 2000 G200 1000 G100 0.0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 VAPOUR QUALITY [/] 0.8 0.9 1.0 Figure Heat transfer coefficient measured during condensation of R245fa in the channel versus vapor quality vol 32 no 2011 94 A CAVALLINI ET AL the heat transfer coefficients at the lower tested values of mass velocity is not found with the refrigerant R245fa, whose trend at low mass velocity is as expected from macroscale condensation The behavior experienced with R32 may be explained with a different flow pattern occurring in the channel heat transfer engineering 80 PRESSURE DROP [kPa] The present authors report here the pressure drop measured during adiabatic two-phase flow of R32 and R245fa inside the same test section previously described For most of the frictional tests, the desuperheater is used to achieve partial condensation of the refrigerant and then the pressure drop is measured adiabatically in the following sector Inlet vapor quality has been controlled through the thermal balance in the desuperheater Some frictional tests, at low vapor qualities, have been performed bypassing the evaporator and sending the refrigerant to the test section as a subcooled liquid; the desuperheater is then used as a preheater for the liquid Indeed, saturation conditions are achieved by partial vaporization before the measuring sector The present mini-tube has a much higher surface roughness as compared to the previously tested multiport minichannel [3] Therefore, the effect of tube wall roughness to the frictional pressure drop was investigated Some single-phase flow tests have previously been performed with R134a to measure the friction factor in the minichannel Experimental values of R134a friction factor have been compared against equations for both laminar and turbulent flows for rough tubes, and good agreement between calculated and experimental values was found [4] The test tube is the same as used for heat transfer tests The arithmetical mean deviation of the assessed profile Ra of the copper channel inner surface is Ra = 2.3 µm, the maximum height of profile Rz is 18 µm The inlet and outlet pressure ports are inserted in two stainless-steel tubes 24 mm long, attached to the ends of the copper tube (adiabatic sectors, Figure 2) The stainless-steel tubes have 0.762 mm inner diameter, Ra = 2.0 µm, and Rz = 10.2 µm The total frictional pressure drop is then the sum of the frictional pressure drop in the two stainlesssteel tubes, of the frictional pressure drop in the 228.5 mm long copper tube, and of the pressure variations due to one abrupt enlargement (from 0.762 mm diameter to 0.96 mm diameter) and one contraction (from 0.96 mm to 0.762 mm) Pressure losses due to abrupt geometry changes account for to 8% of total pressure drop in the R245fa data and for to 10% in the case of R32 data, according to the calculation by means of the Paliwoda [7] equations The experimental uncertainty for the measured pressure difference is ±0.1 kPa, for the absolute pressure is ±3 kPa, for the refrigerant flow rate is ±0.2%, and for the vapor quality ±1% The total experimental pressure drop for R32 at 40◦ C versus vapor quality, at mass velocities of 200, 400, 600, 800, and 1000 kg m−2 s−1, is shown in Figure In Figure 7, the 90 70 G1000 G800 G600 G400 G200 60 50 40 30 20 10 0.0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 VAPOUR QUALITY [/] 0.8 0.9 1.0 Figure Cumulative experimental pressure drop in the test channel during adiabatic two-phase flow of R32 cumulative pressure drop measured during adiabatic two-phase flow of R245fa is reported The combined effect of low vapor density and high liquid viscosity explains the significant pressure drop increase that is measured for R245fa as compared to R32, at the same mass velocity and vapor quality 60 G300 50 G250 G200 PRESSURE DROP [kPa] PRESSURE DROP TESTS 100 40 30 20 10 0.0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 VAPOUR QUALITY [/] 0.8 0.9 1.0 Figure Cumulative experimental pressure drop in the test channel during adiabatic two-phase flow of R245fa vol 32 no 2011 A CAVALLINI ET AL 95 100.00 25 R245fa +20% R32 10.00 -2 -1 PREDICTED HTC [kW m K ] 20 -20% Jg [/] 15 1.00 R245fa G100 10 R245fa G200 0.10 R245fa G400 R32 G100 R32 G200 R32 G400 0.01 0.01 10 15 20 -2 -1 EXPERIMENTAL HTC [kW m K ] 0.10 25 Figure Calculated versus experimental heat transfer coefficient: The model by Moser et al [8], modified by Zhang and Webb [9], is applied to R32 and R245fa data ASSESSMENT OF HEAT TRANSFER CORRELATIONS Experimental results have been compared against two models available in the open literature developed for HTC predictions inside macro-scale tubes The first correlation has been presented by Moser et al [8], and was initially developed for conventional pipes and later modified by using the Zhang and Webb [9] method for pressure drop calculation inside small-diameter tubes Although in this paper all the experimental data points have been compared to the models, the Moser et al [8] correlation was developed for and should be applied only to annular flow condensation The comparison between experimental and predicted data is depicted in Figure As one can see, the model by Moser et al [8] modified with the Zhang and Webb [9] pressure drop correlation is in good agreement with R32 data but overestimates R245fa data by 30% The heat transfer coefficients measured with R32 at 100 kg m−2 s−1 are not in satisfactory agreement with the model This may be due to the different flow pattern occurring in the channel at these conditions As stated earlier, the correlation was developed only for annular flow condensation and therefore it is questionable whether the data at 100 kg m−2 s−1 mass velocity may be included in the comparison The second model used for comparison was developed by Cavallini et al [10] for macroscale condensation It also accounts for the transition from the T-independent to T-dependent region Here T is the saturation minus wall temperature difference However, this transition is defined for heat transfer engineering 1.00 10.00 Xtt [/] Figure Condensation test runs plotted on the flow pattern map (Cavallini et al [10]) conventional tubes, i.e., for channels with hydraulic diameters higher than or equal to mm From flow pattern visualization available in the open literature, one should expect that the stratified flow region is reduced in the case of minichannels, as compared to conventional tubes Matkovic et al [5] reported that the saturation to wall temperature difference has no effect on the heat transfer coefficient at 200 kg m−2 s−1 mass velocity with R32, confirming that the effect of gravity forces in a channel of around mm diameter is not significant in comparison with the other forces that influence the condensation heat transfer at this mass velocity In Figure 9, the test runs at mass velocity ranging between 100 and 400 kg m−2 s−1 are plotted in the diagram of dimensionless vapor velocity versus Martinelli parameter The transition curve provided by Cavallini et al [10] for macro tubes is also plotted This transition curve divides the map in two regions: the upper area characterized by annular flow condensation (where the heat transfer coefficient does not depend on the saturation minus wall temperature difference), and the bottom area where the heat transfer coefficient is dependent on the temperature difference already described The transition line is defined by Eq (4) for HFC refrigerants: JGT = 7.5/ 4.3X tt1.111 + −3 + 2.6−3 −1/3 (4) According to this map, all the data points at mass velocity higher or equal to 200 kg m−2 s−1 lay in the T-independent region and may be predicted by using a model for annular flow condensation In this case the heat transfer coefficient is vol 32 no 2011 174 T0 v− v+ vi x z M M SHABANI ET AL reference temperature, 273.15 K anion stoichiometric coefficient cation stoichiometric coefficient number of cavity type i in hydrate phase per water molecules mole fraction ions electrical charge number [4] [5] Greek Symbols [6] β(0) MX β(1) MX ø CMX γ η µ adjustable parameter adjustable parameter adjustable parameter difference in properties between real and hypothetical empty hydrate phase ionic activity coefficient equation constants chemical potential, J osmotic coefficient [7] [8] [9] Subscripts + CO2 el m mix s T w [10] anion cation CO2 electrolyte mean mixture solvent (water) total amount in the mixture water [11] [12] Superscripts ∗ calc exp H L β pseudo-single electrolyte with ionic strength equal to total mixture ionic strength calculated values using the model experimental data hydrate phase liquid water phase hypothetical empty hydrate phase [13] [14] REFERENCES [1] Sloan, E., and Koh, C A., Clathrate Hydrate of Natural Gases, 3rd ed., pp 1–29, CRC Press, Boca Raton, FL, 2008 [2] Carroll, J J., Natural Gas Hydrate a Guide for Engineers, Gulf Professional Publishing, Burlington, MA, pp 17–38, 2003 [3] Larsen, R., Lund, A., Andersson, V., and Hjarbo, K W., Conversion of Water to Hydrate Particles, 2001 SPE Anheat transfer engineering [15] [16] nual Technical Conference and Exhibition, New Orleans, SPE 71550, pp 15, 2001 Lund, A., and Larsen, R., Conversion of Water to Hydrate Particles—Theory and Application, 14th Symposium on Thermophysical Properties, Boulder, CO, 2000 Parrish, W R., and Prausnitz, J M., Dissociation Pressures of Gas Hydrates Formed by Gas Mixtures, Industrial & Engineering Chemistry Process Design and Development, vol 11, no 1, pp 26–35, 1972 Van Der Waals, J H., and Platteeuw, J C., Clathrate Solutions, Advances in Chemical Physics, vol 2, no 1, pp 1–57, 1959 Holder, G D., and Hand, J H., Multiple-Phase Equilibria in Hydrates from Methane, Ethane, Propane and Water Mixtures, AIChE Journal, vol 28, no 3, pp 440–447, 1982 Holder, G D., and John, V T., Thermodynamics of Multicomponent Hydrate Forming Mixtures, Fluid Phase Equilibria, vol 14, pp 353–361, 1983 John, V T., Papadopoulos, K D., and Holder, G D., A Generalized Model for Predicting Equilibrium Conditions for Gas Hydrates, AIChE Journal, vol 31, no 2, pp 252–259, 1985 Englezos, P., and Bishnoi, P R., Experimental Study of the Equilibrium Ethane Hydrate Formation Conditions in Aqueous Electrolyte Solutions, Industrial & Engineering Chemistry Research, vol 30, no 7, pp 1655–1659, 1991 Tohidi, B., Burgass, R W., Danesh, A., and Todd, A C., Hydrate Inhibition Effect of Produced Water: Part 1—Ethane and Propane Simple Gas Hydrates, Offshore Europe, Aberdeen, UK, SPE 26701, pp 255–264, 1993 Masoudi, R., Tohidi, B., Danesh, A., and Todd, A., A New Approach in Modelling Phase Equilibria and Gas Solubility in Electrolyte Solutions and its Applications to Gas Hydrates, Fluid Phase Equilibria, vol 215, no 2, pp 163–174, 2004 Javanmardi, J., Moshfeghian, M., and Maddox, R N., An Accurate Model for Prediction of Gas Hydrate Formation Conditions in Mixtures of Aqueous Electrolyte Solutions and Alcohol, Canadian Journal of Chemical Engineering, vol 79, no 3, pp 367–373, 2001 Shabani, M M., Rashtchian, D., Ghotbi, C., Taghikhani, V., and Khayat, G., Prediction of Hydrate Formation for the Systems Containing Single and Mixed Electrolyte Solutions, Iranian Journal of Chemistry and Chemical Engineering, vol 26, no 1, pp 35–45, 2007 Peng, D Y., and Robinson, D B., A New Two-Constant Equation of State, Industrial and Engineering Chemistry Fundamentals, vol 15, no 1, pp 59–64, 1976 Nasrifar, K., and Moshfeghian, M., A Model for Prediction of Gas Hydrate Formation Conditions in Aqueous Solutions Containing Electrolytes and/or Alcohol, Journal of Chemical Thermodynamics, vol 33, no 9, pp 999–1014, 2001 vol 32 no 2011 M M SHABANI ET AL [17] Pitzer, K S., and Mayorga, G., Thermodynamics of Electrolytes II Activity and Osmotic Coefficients for Strong Electrolytes With One or Both Ions Univalent, Journal of Physical Chemistry, vol 77, no 19, pp 2300–2308, 1973 [18] Patwardhan, V S., and Kumar, A., Unified Approach for Prediction of Thermodynamic Properties of Aqueous Mixed-Electrolyte Solutions, AIChE Journal, vol 32, no 9, pp 1419–1428, 1986 [19] Lobo, V M M., and Quaresma, J L., Handbook of Electrolyte Solutions, Elsevier Science, Amsterdam, 1989 [20] Dinane, A., and Mounir, A., Water Activities, Osmotic and Activity Coefficients in Aqueous Mixtures of Sodium and Magnesium Chlorides at 298.15 K by the Hygrometric Method, Fluid Phase Equilibria, vol 206, no 1–2, pp 13–25, 2003 [21] El Guedouzi, M., and Azougen, R., Thermodynamic Properties of the Ternary System {yNH4 Cl + (1−y)MgCl2 } (aq) at 298.15 K, Fluid Phase Equilibria, vol 253, no 2, pp 81–87, 2007 [22] Dinane, A., El Guendouzi, M., and Mounir, A., Hygrometric Determination of Water Activities, Osmotic and Activity Coefficients of (NaCl + KCl)(aq) at T = 298.15 K, Journal of Chemical Thermodynamics, vol 34, no 4, pp 423–441, 2002 [23] Roo, J D., Peters, C J., Lichtenthaler, R N., and Diepen, G A M., Occurrence of Methane Hydrate in Saturated and Unsaturated Solutions of Sodium Chloride and Water in Dependence of Temperature and Pressure, AIChE Journal, vol 29, no 4, pp 651–657, 1983 [24] Dholabhai, P D., Englezos, P., Kalogerakis, N., and Bishnoi, P R., Equilibrium Conditions for Methane Hydrate Formation in Aqueous Mixed Electrolyte Solutions, Canadian Journal of Chemical Engineering, vol 69, no 3, pp 800–805, 1991 [25] Dholabhai, P D., Kalogerakis, N., and Bishnoi, P R., Equilibrium Conditions for Carbon Dioxide Hydrate Formation in Aqueous Electrolyte Solutions, Journal of Chemical and Engineering Data, vol 38, no 4, pp 650–654, 1993 heat transfer engineering 175 [26] Englezos P., and Bishnoi P R., Prediction of Gas Hydrate Formation Conditions in Aqueous Electrolyte Solutions, AIChE Journal, vol 34, no 10, pp 1718–1721, 1988 [27] Javanmardi, J., Moshfeghian, M., and Maddox, R N., Simple Method for Predicting Gas-Hydrate Forming Conditions in Aqueous Mixed-Electrolyte Solutions, Energy and Fules, vol 12, no 2, pp 219–222, 1998 Mohammad M Shabani is a Ph.D student at the Norwegian University of Science and Technology, Department of Energy and Process Engineering He received his master’s degree in chemical engineering from Sharif University of Technology, Tehran, Iran, in 2003 Since then, he has been working on gas hydrate-related issues He is currently working on multiphase flow with gas hydrate particles both on multiphase flow and phase equilibrium aspects Ole J Nydal is professor at the Norwegian University of Science and Technology (NTNU), Department of Energy and Process Engineering His educational background is from NTNU (M.Sc physics) and the University of Oslo (Ph.D fluid mechanics) Before joining NTNU in 1998 he spent more than 10 years at the Institute for Energy Technology working with multiphase transport of oil–gas mixtures in pipelines His research interest is mainly within transient multiphase flows, both experimental and numerical Most of the work is made in collaboration with industry Roar Larsen is chief scientist at SINTEF Petroleum Research, and adjunct professor of gas hydrates at the Norwegian University of Science and Technology (NTNU), both in Trondheim, Norway He received his Dr Ing degree from NTNU in 1997, having divided his study between NTNU and the Colorado School of Mines in the United States He devotes most of his time to developing technologies for allowing oil and gas transport systems to operate free of gas hydrate deposition problems In addition, he teaches one of the world’s few university-level gas hydrate courses to master students at NTNU vol 32 no 2011 Heat Transfer Engineering, 32(2):176–188, 2011 Copyright C Taylor and Francis Group, LLC ISSN: 0145-7632 print / 1521-0537 online DOI: 10.1080/01457631003769377 Combined Experimental and Numerical Flow Field Study of Inclined Louvered Fins CHRISTOPHE T’JOEN, HENK HUISSEUNE, ARNOUT WILLOCKX, ` HUGO CANIERE, and MICHEL DE PAEPE Department of Flow, Heat and Combustion Mechanics, Ghent University, Gent, Belgium In this study the flow behavior within an interrupted fin design, the inclined louvered fin, is investigated experimentally through visualization and numerically through computational fluid dynamics (CFD) simulation The inclined louvered fin is a hybrid of the offset strip fin and standard louvered fin, aimed at improved performance at low Reynolds numbers for compact heat exchangers The flow behavior is studied in six geometrically different configurations over a range of Reynolds numbers and quantified using the concept of “fin angle alignment factor,” which is related to the flow efficiency in louvered fins The experimental data resulted in a discrete data set of local fin angle alignment factor values, which were used to validate the simulations Using these validated cases it is shown that the graphical measurement method can be distorted by recirculation zones, resulting in erroneous values Care should thus be taken when performing graphical measurement of the mean flow angle based on dye injection images The transition from steady laminar to unsteady flow in inclined louvered fins is geometrically triggered and occurs at lower Reynolds numbers compared to slit fins and standard louvered fins This property can potentially be used to further improve on the performance of interrupted fin surfaces INTRODUCTION For refrigerant-to-air heat exchangers, the air-side thermal resistance can be dominant, representing up to 85% of the total heat transfer resistance To increase the heat transfer rate, the exterior surface area is enlarged by adding fins Manufacturers continuously seek to increase the fin performance in order to reduce the heat exchanger size and cost Today, highly interrupted surfaces are widely used to enhance the thermal performance of compact heat exchangers These interrupted fins exploit two mechanisms to provide a performance improvement compared to continuous fins: (1) Interrupted surfaces restart the thermal boundary layer, and because the resulting average thermal boundary layer thickness is smaller for several short plates than for one long plate, the average convective heat transfer coefficient is higher for the interrupted surfaces; and (2) above a critical Reynolds number, interrupted surfaces can cause vortex shedding and the resulting mixing and flow unsteadiness result in an increased heat transfer Address correspondence to Dr Christophe T’Joen, Department of Flow, Heat and Combustion Mechanics, Ghent University, Sint-Pietersnieuwstraat 41, 9000, Gent, Belgium E-mail: christophe.tjoen@ugent.be Two widely used interrupted fin designs are the offset strip or slit fin and the louvered fin Both consist of arrays of flat plates In offset strip fins these plates are aligned to the main flow, while in louvered fins the plates are set at an angle to the flow The inclined louvered fin is a hybrid design of louvered and offset strip fins, as described by Shah et al [1] The plates are aligned with the main flow (Figure 1) but are set out in a staggered layout, forming a deflecting channel In Figure the main geometric parameters are indicated: the fin pitch Fp , the louver pitch Lp , the fin thickness t, the fin angle ϕ , and the number of louvers N Just as in louvered fins there are distinct inlet louvers, turnaround louvers, and exit louvers The design shown in Figure represents the interrupted section of the fin surface between the tubes This interrupted section needs to be connected to the tubes to form the heat exchanger In modernday louvered fin heat exchangers this is done through a flat fin surface, a landing To manufacture these fins a complex series of machining processes such as stamping, rolling, etc will be required This study is focused solely on the characteristics of the flow passing through the interrupted element, and is aimed at determining possible future uses of this fin type No further study has been performed regarding the manufacturing, as first it should be established whether this design can result in improved 176 C T’JOEN ET AL 177 Figure The inclined louvered fin array and relevant geometric parameters The flow passes from left to right heat exchanger performance However, in view of current multilouvered and interrupted designs in production, there appear to be no significant barriers to the production of the proposed design Shah et al [1] report that the inclined louvered fin offers a significant improvement in the ratio of heat transfer to pressure drop based on earlier work by Tanaka et al [2] and Suzuki et al [3] To the authors’ knowledge no other studies have been presented on this design Shah et al [1] conclude that the best potential for future improvements in numerical modeling of louvered fin designs lies within the understanding of the performance of inclined louvered fins, which is why this research was undertaken Tanaka et al [2] previously studied inclined louvered fins and considered another parameter: the louver angle, which is set at 0◦ in the current study This value was fixed to reduce the number of geometrical parameters in the experimental survey Both louvered fins and offset strip fins have been studied extensively in the past Chang and co-workers [4, 5] developed a correlation for the Colburn and friction factor of louvered fins based on a large set of experimental data Manglik and Bergles [6] presented an overview of the previous experimental studies on offset strip fins and, using a multi-regression analysis, determined a correlation for the heat transfer and pressure drop Both these data sets were obtained through Wilson-plot testing (the method is described by, e.g., [7, 8]) of full-scale heat exchangers Wilson-plot testing treats the heat exchanger in essence as a black box Testing the heat exchanger with different geometric parameters will show the thermo-hydraulic impact of the variation, but it won’t reveal the actual physics behind it To better understand the coupling between the flow behavior and resulting thermo-hydraulics, several authors have focused solely on the interrupted fin pattern and studied the local flow behavior and heat (or mass) transfer coefficients This has been done both experimentally (using scaled models) and numerically (computational fluid dynamics [CFD] simulations), and this is the approach that will be followed in this research program as well Figure In a standard louvered fin array two extremes in flow pattern can occur: In louver-directed flow the streamlines are parallel to the louvers; in duct-directed flow the streamlines pass straight from left to right with the louvers at high Reynolds numbers These findings were later confirmed by other researchers, e.g., Davenport [10], Cowell et al [11], and DeJong and Jacobi [12] Davenport [10] speculated that the flow angle dependency on the Reynolds number was a result of boundary layer growth on the louvers At low Reynolds numbers, the thick boundary layers block the passage between the louvers, forcing the flow to pass in between the different fins This is referred to as “duct-directed flow” in Figure As the Reynolds number increases, the flow passages open up and “louver-directed flow” becomes apparent The flow is deflected, extending the flow path throughout the fin array and aligning the flow with the louvers, increasing the heat transfer rate (as shown by [11, 12]) But as the flow path is extended, the frictional pressure drop increases The degree of flow deflection is usually quantified using the concept of “flow efficiency,” η This is the ratio of the mean flow angle to louver angle For small louver angles, this can be approximated by the ratio of the actual transverse distance D (as shown in Figure for an inclined louvered fin array) traveled by a streamline to the ideal distance Did the flow would travel if the flow were fully aligned with the fin angle as expressed in Eq (1) [12] η= tan α D α ≈ = θ tan θ Did (1) In standard louvered fins, the flow angle approaches the louver angle at high ReLp , as can be seen in Figure The change FLOW BEHAVIOR IN LOUVERED FINS Standard louvered fins have two distinct flow regimes Beauvais [9] was the first to show, using smoke trace visualization, that the flow in louvered fins is deflected and becomes aligned heat transfer engineering Figure Graphical approximation to the fin angle alignment factor ζ: the ratio of the actual transverse distance D traveled by a streamline to the transverse distance Did the streamline would travel if it were perfectly aligned with the fin angle vol 32 no 2011 178 C T’JOEN ET AL Figure Example flow efficiency data for three different arrays obtained by flow visualization [12] in flow behavior is due to the flow following the path of least resistance—the path corresponding to the lowest overall pressure drop A detailed discussion on the flow efficiency of louvered fins (the transition from duct to louver directed flow and the impact the geometric parameters and the Reynolds number) is provided by DeJong and Jacobi [13] and DeJong [14] Sahnoun and Webb [15] derived an analytical model to predict the heat transfer and pressure drop of standard louvered fin automotive heat exchangers This model included a correlation for η based on the measurements of Webb and Trauger [16] The model showed reasonable agreement (± 20%) with experimental data of Davenport [10] Achaichia and Cowell [17] found a strong coupling between the heat transfer rate of the tested flat tube heat exchangers with louvered fins and the mean flow behavior, as they were able to relate the Stanton number with η (Eq (2)) St = 1.18 · η · Re−0.58 Lp (2) Zhang and Tafti [18] presented an overview of previously measured flow efficiency data for louvered fins, and found considerable discrepancy This is most likely due to the impact of the channel walls on the flow, as shown by DeJong and Jacobi [12] Some studies were preformed using only three to seven fins, which is too few to prevent significant flow distortion due to wall proximity Using a large database of validated two-dimensional numerical simulations, Zhang and Tafti [18] determined a correlation for the flow efficiency Results show that flow efficiency is strongly dependent on geometrical parameters, especially at low Reynolds numbers Flow efficiency increases with ReLp and louver angle, while decreasing with fin pitch and thickness ratio Local heat transfer coefficients in standard louvered fins presented by Lyman et al [19] showed the impact of the local thermal field surrounding a louver If a louver is surrounded by the hot thermal wake of an upstream louver, it will have a lower heat transfer coefficient It is therefore important, as noted by Tanaka et al [2], to ensure a good positioning of the different louvers to improve the heat transfer rate of a given fin heat transfer engineering design This concept is known as “thermal wake management.” Using numerical simulations, Zhang and Tafti [20] studied the wake behavior within louvered fin arrays, and distinguished two types of wake interference: inter- and intrafin The first type of wake interference is caused by the thermal wakes of louvers of a different fin, while intrafin wake interference is caused by the thermal wakes of louvers of the same fin Although both modes of interference are present in duct and louver directed flow, from their description it is clear that duct directed flow will surely give rise to intrafin interference, while louver directed flow is more prone to interfin interference Summarizing, previous studies clearly show a strong coupling between the mean flow behavior and the thermohydraulics in standard louvered fins, and this is due to the nature of the flow, which is driven by boundary-layer growth on the louvers Because this phenomenon can be described as a function of the louver pitch, this value has been used as the reference length scale in the Reynolds number (ReLp ) by many researchers in the past, e.g., DeJong and Jacobi [12, 13, 21] and Tafti et al [22–24] These studies were focused solely on the louvered element For full-scale louvered fin-and-tube heat exchangers it is not as clear which length scale is appropriate: Some correlations use the tube exterior diameter or collar diameter as reference length scale in the Reynolds number (ReDc ) (e.g., Chang et al [4, 5] and Wang et al [25]), while others consider the hydraulic diameter of the channel (ReDh ) (e.g., Malapure et al [26]) Because only the interrupted fin element is considered in this study, and similar to standard louvered fins a boundary-layer-driven flow was found to be present (T’Joen et al [27]), it was chosen to use the louver pitch as a reference length for the Reynolds number in this research as this is related to the flow physics If the geometry of the inclined louvered fin (Figure 1) is considered, a priori it is clear that similar phenomena as in standard louvered fins will occur At low ReLp the thick boundary layers will block the passages between the louvers, forcing the flow to wind up and down As ReLp increases, the flow passages will open up and the flow becomes more offset strip fin like, aligned with the louvers Tanaka et al [2] studied the flow through an inclined louvered fin using a two-dimensional analytical model The primary aims were to ensure that the airflow passed straight through the fin array, avoiding deflection to reduce the air-side pressure drop, and to enhance the heat transfer coefficient by positioning each louver outside of the boundary layer of an upstream louver The flow in various inclined louver models was visualized using aluminum powder as a tracer For high fin angles (35◦ ) strong wakes and deflection were noted, while for small angles (10◦ ) a strong developing boundary layer was detected An intermediate fin angle (20◦ ) resulted in an almost straight flow with little separation or deflection Tanaka et al [2] concluded that a high average heat transfer coefficient could be realized in inclined louvered fins provided that (1) the minimum traverse space to the adjacent louver is wide enough to include the boundary layers and (2) the longitudinal distance between the vol 32 no 2011 C T’JOEN ET AL 179 Jacobi [21] visualized the flow in three slit fin arrays with different geometric parameters and by combining this information with mass transfer data were able to show the impact of the onset of unsteady flow onto the local mass transfer coefficients Using numerical simulations, Tafti and co-workers [22–24] studied the transition from steady laminar to unsteady flow in a standard louvered fin and also reported increased heat transfer rate as flow became unsteady EXPERIMENTAL SETUP AND PROCEDURE Figure Fin designs studied by Suzuki et al [3] Top: offset strip fin; middle: parallel louver fin “PL1”; bottom: parallel louver fin “PL2.” stream-wise louvers is wide enough to eliminate wakes After a series of large-scale model tests, a single heat exchanger was built, tested, and compared to a standard louvered fin In the considered velocity range (1–5 m/s) the average heat transfer coefficient of the inclined louvered fin core was 16% higher and the pressure drop 21–27% lower compared to the standard louvered fin heat exchanger This example clearly indicates the potential of this fin type Suzuki et al [3] used numerical simulations to compare three different fin designs: a slit fin and two “parallel louver fins.” The designs are shown in Figure The design “PL2” is very similar to the geometry studied here except for the lack of entry, turnaround, and exit louvers In the numerical simulations the fins were assumed to be infinitely thin They found that the inclined louvered fin PL2 showed the highest Nu, being 25% higher at ReLp = 60 and 100% higher at ReLp = 250 than that of the slit fin The geometry PL1 resulted in an intermediate increase compared to the slit fin Similarly, the PL2 geometry had the largest pressure drop, PL1 the intermediate, and the slit fin the lowest pressure drop By studying the air velocity plots, Suzuki et al [3] showed that at low Reynolds numbers the flow deflected in geometry PL2, which resulted in an increase of the pressure drop Suzuki et al [3] concluded that the improved heat transfer performance of the PL1 and PL2 design was due to longer recovery lengths for the louver wakes: By increasing the stream-wise (following the flow) distance between louvers the wake can recover, reducing the velocity defect and temperature defect, and this increases the mean convective heat transfer coefficient Both in slit fins and standard louvered fins it has been shown that starting from a certain ReLp self-sustained oscillations are present within the fin array These oscillations result in increased mixing of the flow and enhanced local heat transfer DeJong and heat transfer engineering The aim of the current research is to study the flow in inclined louvered fins, to determine the average flow behavior over a range of ReLp , as well as to study the transitional behavior from laminar to unsteady flow As shown earlier, these data can provide valuable insights for possible future applications of this fin type as well as help focus the future thermo-hydraulic study on configurations and Reynolds numbers of interest To this end, a series of visualization experiments have been performed in a water tunnel In order to quantify the flow behavior in the inclined louvered fins, the “fin angle alignment factor” ζ (defined in Eq (3) as the ratio of the mean flow angle α to the fin angle ϕ ) is introduced ς= tan α D α ≈ = ϕ tan ϕ Did (3) Note that the definition is identical to that of the “flow efficiency” η for standard louvered fins For standard louvered fins, a high value of η corresponds to a flow almost parallel to the louvers, resulting in large heat transfer coefficients Thus, high values of η are good from a heat transfer perspective, resulting in the term flow efficiency for η For the inclined louvered fin array, high values of ζ indicate that the flow along the louver surface is dramatically reduced, and the local heat transfer coefficient can be expected to be reduced This would make the term “flow efficiency” misleading for inclined louvered fins; thus an alternate term is used, the “fin angle alignment factor.” In this paper no thermo-hydraulic data will be presented, but just as for standard louvered fins, this flow behavior parameter will be an interesting tool to further understanding in the thermo-hydraulics of the inclined louvered fin, and a strong link between ζ and the thermo-hydraulic behavior is expected It should also be pointed out that the definition of ζ (Eq (3)) will have to be modified in the case of louvers with a nonzero angle of attack This is not considered here As can be seen in Figure 1, there are five potentially important geometric parameters: fin pitch, louver pitch, fin angle, number of louvers, and fin thickness In a previous study (T’Joen et al [28]) a numerical screening was performed to determine which parameters had the strongest impact on the thermo-hydraulic behavior It was found that the fin pitch, the fin angle, and louver thickness had the largest impact This is consistent with the results found for standard louvered fins [4, 12] Similarly for offset strip fins, Manglik and Bergles [6] found that the fin thickness vol 32 no 2011 180 Table C T’JOEN ET AL Geometric data of the tested samples Fin parameters Configuration Configuration Configuration Fp ϕ t Fin parameters Fp ϕ t 0.0340 m 12.64◦ 0.002 m Configuration 0.0225 m 30.96◦ 0.002 m 0.0225 m 12.64◦ 0.002 m Configuration 0.0276 m 19.78◦ 0.002 m 0.0340 m 30.96◦ 0.002 m Configuration 0.0276 m 19.78◦ 0.004 m and fin spacing had a strong impact on the thermo-hydraulic behavior Therefore six different fin configurations were studied, with varying fin pitch, fin angle, and louver thickness, to determine the impact of these parameters on the flow behavior The geometric parameters of the six configurations can be found in Table The different configurations all had five louvers (N = 5) Scaled models (20:1) of the inclined louvered fin were placed in the test section of a water tunnel The scaling factor was selected in order to obtain sufficient spatial resolution, while maintaining a sufficiently large number of fin rows to ensure periodic flow behavior This resulted in a louver pitch of cm The fin length was 0.2 m DeJong and Jacobi [12] previously studied the flow in scaled models of standard louvered fins and found that the channel walls can have a profound impact on the measured flow efficiency A sufficiently large number of fins is required to accurately represent the fin array and to avoid these wall effects A mirror mounted at 45◦ underneath the test section allows for convenient recording of the images of the flow A standard digital camera was used to obtain the images In order to visualize the flow, a solution of food dye and water was prepared and injected upstream of the test section This dye had a density slightly higher than that of the water in the test section, and although the orientation of the test section was such that buoyancy did not influence the measurement of ζ, experiments were limited to Reynolds numbers for which buoyancy effects were insignificant (ReLp ∼ 130) The dye was fed to the 1-mm-diameter injection needle by gravity, and the injection rate was adjusted with a small valve to try to match the flow velocity at the point of injection Dye color, concentration, and delivery rate were adjusted in a trial-and-error fashion to achieve good image contrast and a faithful representation of the flow More details about the experimental setup and procedure can be found in T’Joen et al [27] For each of the six configurations (shown in Figure 6; the geometric parameters are listed in Table 1), images of the flow field were recorded at various Reynolds numbers At each ReLp the dye injection point was varied over the channel width This resulted in a series of images of various streamlines, which, once combined, provide an overview of the total flow field at a given Reynolds number For each of the images ζ can be determined graphically The images were imported into a graphical software package The mean flow angle was determined between the point where the streamline entered the fin array and the midpoint of heat transfer engineering Figure Geometric representation of the six different configurations studied through flow visualization the turnaround louver, as shown in Figure To correct for any optical distortion both at the inlet and at the turnaround louver the distances were expressed as a ratio of the local fin pitch This results in Eq (4) for the transverse distance D An extra parameter τ is introduced, representing the number of times the streamlines passes a fin before reaching the endpoint So if the streamline remains within the same fin passage as the injection point, τ = 0, for each fin the streamline crosses τ is increased by Thus, for the streamline in Figure 3, τ = The ideal traveled distance Did can be expressed based on the fin geometry as in Eq (5) This results in Eq (6) for the fin angle alignment factor D = Did + y x − − τ · Fp Fp1 Fp2 (4) Did = (N + 1.15) · Lp · tanϕ ς≈ (N + 1.5) · Lp · tanϕ + (5) x Fp1 − y Fp2 (N + 1.5) · Lp · tanϕ − τ · Fp (6) In order to get an indication of the quality of the measurements, a thorough error analysis was made using the procedures found in Taylor [29] The water temperature uncertainty was 0.5◦ C The water velocity was determined by measuring the time required for a dye marker to travel a given distance This measurement was repeated five times for each ReLp and the mean value was used in the calculations The standard deviation of these five measurements was used to determine the uncertainty value for the water velocity The louver plates were cut using a precision tool resulting in an uncertainty of 0.5 mm These values resulted in average uncertainty of 5% on ReLp A 2% uncertainty was assumed for physical properties (density and dynamic viscosity) To determine the fin angle alignment factor using Eq (6) the distance traveled by the flow ( x and vol 32 no 2011 C T’JOEN ET AL y ) and the local fin pitch (Fp1 and Fp2 ) were measured graphically The local width of the dye streakline at the inlet and at the turnaround louver was used as uncertainty value for x and y; the width of a pixel was used as the uncertainty on Fp1 and Fp2 Assuming an uncertainty of 1◦ on ϕ, the uncertainty for ζ could be determined For configuration the relative uncertainty on ζ ranged from 3.5% to 18% Similar values were found for configurations 2, 5, and For configurations and much greater values for the relative uncertainty were found due to the small flow deflection and the finite thickness of the dye streak The uncertainty on ζ for configuration and ranged from 30% to 90% NUMERICAL MODEL AND METHODS The graphical measurement method to determine averaged ζ values is very tedious Considering that a numerical simulation contains all information on the local flow angle, it should be possible to derive an average ζ directly from the CFD data In order to distinguish between the graphically determined (based on either experimental or numerical flow images) and the numerically determined ζ values (Eqs (7) and (8)), the latter is referred to as ζT This was shown by Zhang and Tafti [18] for louvered fins For each louver (excluding the inlet, turnaround, and exit louver) an average flow angle αL is determined using the average velocity components over the fin passage, as in Eq (7) The average velocities were determined on a cut perpendicular to the louver spanning the fin passage, intersecting the louver at the center The αL values are then averaged out over all the louvers and divided by the fin angle to determine the fin angle alignment factors; see Eq (8) Downstream of the turnaround louver, the average flow angle is divided by the negative fin angle as the flow now points down instead of up This definition of ζT makes sense if we recall the original goal of this parameter, which is to quantifiably describe the mean flow behavior in the louver passages αL = tan−1 ςT = αL ϕ Fp v.dy Fp u.dy 181 Figure Basic representation of the two-dimensional geometry that is studied in the CFD simulations was simulated with an entry region (two fin pitches) and exit region (five fin pitches) This represents the periodic unit of an infinite stack of fins The height of the computational domain is set to one fin pitch with periodic boundary conditions on the top and bottom of the domain This is schematically shown in Figure At the inlet a uniform velocity in the x direction (parallel to the louvers) was imposed At the outlet the static pressure was set to Pa (pressure outlet boundary condition) The double precision segregated solver was used The standard Navier–Stokes equations were used As these equations are well known and can be found in many textbooks, they are not presented here; consult e.g., the Fluent manual [31] Convergence criteria were set to 1e-8 for continuity and velocity components and 1e-6 for energy Second-order discretization was used in combination with the SIMPLE algorithm Substance properties were set to a constant value The average pressure drop and outlet temperature were monitored during the iterations to determine if the simulations had converged If no steady-state convergence was obtained, an unsteady simulation was performed till the different monitored signals (mean inlet pressure, local velocity in the wake of the louvers and mean outlet velocity) showed a clear periodic behavior For those cases data were averaged out over 10–15 periods of the signal A grid independence study was performed; the results are presented in T’Joen [32] In order to determine ζ from the simulations, the same procedure was used as for the water tunnel experiments The streamlines were visualized using the “path lines” option in Fluent and the resulting image was exported to the graphical software The same data reduction was performed as for the experimental flow images, which resulted in a set of numerical data points, spread evenly out over the fin spacing (7) RESULTS AND DISCUSSION (8) Local Fin Angle Alignment Factors The flow through the different configurations was simulated using Fluent Two-dimensional cases were studied using unstructured quadrilateral meshes generated by Gambit The highest value of cell skewness in the different grids was 0.45 and only 1% of all the cells had a skewness larger than 0.14 The largest cell aspect ratio was 1.57, but only 1% of the cells had an aspect ratio larger than 1.1 Using two-dimensional cases resulted in a strong reduction of computational time compared to three-dimensional cases, and Tafti and Cui [30] had shown that in louvered fins the flow over the louvers is mainly twodimensional for the considered ReLp range A single louver row heat transfer engineering For each configuration a series of graphs can be produced, one per Reynolds number, showing the local values of ζ set out against the injection location This value represents the position of the injector relative to the fin passage and is scaled with the fin pitch, so it ranges from (top of the fin passage) to (bottom of the fin passage) An example is shown in Figure (filled dots) The error bars are indicated There is clearly a strong variation depending on the injection location Similar findings were reported by Wartick and Jacobi [33] and DeJong and Jacobi [13], who stated that the flow efficiency value depends on the vol 32 no 2011 182 C T’JOEN ET AL Figure Comparison of the experimentally and numerically determined local ζ values set out against the injection point, configuration 5, with ReLp = 402 injection location They stated that the maximum value was found for a streamline in the center of the fin passage, and the minimum for a streamline touching the inlet louver In their data reduction the average of this minimum and maximum was used It is clear from Figure that using the average of the maximum and minimum value of the experimental data points would not represent a fair average of ζ To determine a correct average value the gaps between the data points should be filled with additional data This is done by using the CFD data as described earlier The results are shown for the same case in Figure (white dots) As can be seen, the agreement between the simulations and the experimental data is very good (the difference is for most data points smaller than 0.05) This was the case for all of the considered Reynolds numbers To verify the assumption that the flow is two-dimensional (2D), three-dimensional (3D) simulations were performed on configuration The 2D domain was extended into the z direction over a distance of 10 cm (half the fin length) Symmetry conditions were imposed on one of the sidewalls while the other one was set to a nonslip wall The remaining boundary conditions and solver settings were identical to those of the 2D simulations The same data reduction procedure was used (Eqs (7) and (8)): ζT was determined based on the average velocity angle over a series of cuts perpendicular to the louvers at different distances from the wall The results are illustrated in Figure As can be seen in Figure 9A, the difference between the average ζT values determined using 2D and 3D simulations is negligible in the considered Reynolds number range Figure 9B shows the spanwise variation of the average ζT values for ReLp = 523 As can be seen, similar to the findings of Tafti and Cui [30] in louvered fins, the flow over the interrupted fin element is primarily two-dimensional, resulting in a flat ζT profile Only in a small zone near the wall a slight reduction can be seen As stated earlier, in this research only the interrupted fin element is considered In future applications a landing will be required to connect the interrupted element to the tube wall This will result in local strong three-dimensionality of the flow Average Fin Angle Alignment Factors Using the validated numerical data it is possible to determine a well-defined average value for ζ In essence the measured data points (from the simulations) are a discretization of the continuous fin angle alignment function ζ(x), spanning between and Due to the way these values are determined (graphic measurement from path lines) no continuous data could be extracted from the numerical results So it is important to verify that the discretization is sufficiently fine, capturing the function shape By integrating a series of constrained cubic splines (based on the dataset), it was shown (T’Joen [32]) that a minimum of 18 points was required for a good estimation of the average ζ Figure Comparison of the numerically determined average ζT values based on 2D and 3D simulations for configuration 2: (a) ζT set out against ReLp , (b) ζT set out against z (ReLp = 523) (distance from the wall) heat transfer engineering vol 32 no 2011 C T’JOEN ET AL Figure 10 Averaged ζ values for the different configurations over the studied Reynolds number range The filled symbols represent validated numerical simulations The open symbols are purely numerical cases without experimental data The resulting averages can be seen in Figure 10 The filled symbols represent cases that were also measured in the water tunnel, while the open symbols are cases that are only simulated numerically For ReLp < 130 it was not possible to visualize the flow due to the density difference between the dye and the surrounding fluid At high Reynolds numbers the flow became unsteady, making measuring ζ experimentally impossible Using CFD simulations it was possible to visualize the flow beyond the experimental range When comparing the averaged fin angle alignment values of the different configurations, a large difference is apparent The flow is not deflected by configuration (ζ ≈ 0), and only very slightly by configuration The flow deflection by configuration 1, 2, and are very similar, with values ranging between 0.4 and 0.6 Configuration presents an intermediate case with values ≈ 0.2 It is clear that the fin geometry has a strong impact on the ζ values, just as it has on η for louvered fins This large variation between the fin angle alignment values of the different configurations can be explained when considering the geometries, Figure As can be seen, both configuration and have very large passages in between the louvers, resulting in very “open” configurations The flow deflection is triggered by boundary layer growth blocking the passage between the louvers In these open configurations, the impact of the boundary layer growth is small to negligible in the considered ReLp range, so the flow is hardly deflected Reducing the size of the louver passage increases the fin angle alignment factor, or, conversely said, increases the flow deflection If these results are compared to the mean flow behavior in louvered fins, considering, e.g., Figure 4, two key differences can be seen: • The flow deflection is considerably lower than those found in standard louvered fins, considering, e.g., the work by DeJong and Jacobi [12] reporting values between 0.6 and 0.9, and the numerical work by Zhang and Tafti [18], who reported η values between 0.5 and 0.95 in the same ReLp range So it is clear that inclined louvered fins generate flow deflection, similar heat transfer engineering 183 Figure 11 Comparison of graphically determined ζ values for configuration (using the path lines) and the numerically determined ζ values to louvered fins, but that the deflection is not as pronounced as in louvered fins • Figure 10 also shows that for most configurations (3, 4, 5, 6) the flow deflection is nearly constant in the considered Reynolds range For configurations and the values decrease but asymptotically approach a value of ∼0.4 The decreasing trend in configurations and is consistent with the boundarylayer growth model This is a large difference with standard louvered fins, which show a sharp reduction in η (and thus become less effective) once ReLp drops below 300 The flow deflection also shows an opposite trend in inclined louvered fins, increasing as the Reynolds number decreases A comparison between the numerically determined average ζT value (diamond symbols) and the graphically determined ζ value (round symbols) is shown in Figure 11 As can be seen, there is a considerable difference between the two values, especially at low ReLp To further investigate this difference, Figure 12 Graphical determination of the fin angle alignment factor on L5 and the center of the turnaround louver vol 32 no 2011 184 C T’JOEN ET AL Figure 13 Comparison between the flow visualization of the local flow behavior around the turnaround louver and the numerical simulation of the streamlines, configuration 5, with ReLp = 402 additional graphical measurements were performed on the same images (both the numerical path lines and the flow visualization), but with different starting points and endpoints These measurements were done to determine whether the graphically determined values are independent of the chosen measurement location, which is an important requirement, considering the goal of ζ is to be a measure of the mean flow behavior around the louvers Two other measurement locations were chosen upstream of the turnaround louver: the center of L5 and the center of L4 This resulted in the data labeled as ζL5 and ζL4 Just as for the measurements at the center of the TL, very good agreement was found between the experimentally and numerically determined data (flow visualization) and the numerical simulations (path lines) The results are also shown in Figure 11, with square symbols for ζL5 and triangle symbols for ζL4 As can be seen there is a nearly constant difference over the considered Reynolds number range between ζ and ζL5 The ζL5 data nearly coincides with the ζL4 data Table contains an overview of the averaged absolute value of the difference between the various data sets for the six configurations As can be seen, the difference between ζ and ζL5 for all configurations is ∼0.08 This value can be deduced theoretically Consider the streakline in Figure 12 Both ζ and ζL5 are determined; see Eqs (9) and (10) If one assumes the area underneath the turnaround louver is filled with a recirculation zone (Figure 13), it is inaccessible for the streamlines The streamlines are thus displaced over a given distance, and a relationship can be determined between Table Average absolute difference between the numerically determined ζT and the graphically determined ζ, ζL5 , ζL4 , and ζL1-L5 , with ReLp ranging from 20 to 600 Configuration |ςT − ς| 0.09 0.06 0.03 0.01 0.04 0.14 |ς L5 − ς| |ς L5 − ς L4 | 0.084 0.080 0.083 0.076 0.078 0.078 0.009 0.005 0.01 0.01 0.008 0.006 |ς L1−L5 − ςT | 0.018 0.017 0.012 0.01 0.006 0.009 heat transfer engineering D1 and D2, and in Eqs (11)–(15) ς= D1 α1 ≈ ϕ D1id (9) ςL5 = α2 D2 ≈ ϕ D2id (10) D2 = D1 + LP tan α1 − LP tan ϕ D1 D1id D2 LP LP LP tan ϕ = − tan α1 + D1id D2id D2id D2id 2 D2id (11) (12) D1id = 5.5 · LP · tan ϕ (13) D1 5.5 · LP (14) tan α1 = ς L5 = ς + = ς + 0.076 · 5.5 (15) The difference between ζL5 and ζ for the different configurations has been set out against the ReLp in Figure 14, and, as the previous relationship suggests, it is independent of the Reynolds number This quantifies the impact of the separation zone underneath the turnaround louver and shows that the separation zone underneath the turnaround louver distorts the graphical measurements Determining the fin angle alignment factor at the center of L5 is a better option, as it is more representative of the flow around the louvers Graphically determined fin angle alignment factors at the center of L5 and the center of L4 resulted in almost the same value, as shown in Table Figure 11 shows that at very low ReLp , ζL5 values coincide with the numerically determined ζT values, but as the Reynolds number increases they slowly diverge A possible explanation is the occurrence of the separation zone underneath the inlet louver (Figure 15) At very low Reynolds numbers this separation does not occur, but as the Reynolds number increases a small recirculation zone appears (which increases in size as the Reynolds number increases), pushing the streamlines lower, vol 32 no 2011 C T’JOEN ET AL 185 (standard louvered fins or a variation) because local separation zones can result in a distorted result These also show that a detailed analysis of the local data is required in order to verify that the presented average values are indeed true averages as the local variations can be strong This can be done using validated numerical simulations Transition to Unsteady Flow Figure 14 Difference between ζL5 and ζ set out against the Reynolds number for the different configurations similar to what happens underneath the turnaround louver To eliminate the impact of this recirculation zone, a final series of graphical measurements was performed on the data set, determining ζ by measuring graphically between the center of L1 and the center of L5 The result is shown in Figure 11 as ζL1—L5 As can be seen, these data points agree very well with the numerically determined ζT values, and this agreement was found for all configurations Zhang and Tafti [18] presented a comparison of their numerically determined η values to the graphically measured values by DeJong and Jacobi [12] The agreement was good, with a difference of ∼0.1 for ReLp = 160 and 0.05 for ReLp = 600 They did not elaborate on the possible source of these small differences, but most likely similar effects take place in the louvered fins; e.g., if a recirculation zone were to form on the bottom side the turnaround louver, it would force the streamlines to shift lower, resulting in a higher value for η These findings have a practical significance, as they show that great care should be taken when determining the reference points to measure the mean flow angle in a interrupted fin design During the flow visualization the transition from steady laminar flow to unsteady flow was also studied in detail A description of the transition for each configuration is presented in T’Joen et al [27] and T’Joen [32] It was found that the initial instability appears in the wake of the exit louver As the Reynolds number increases the flow instabilities move upstream into the fin array The onset of unsteady flow occurred at low Reynolds numbers (200–300) Figure 16 shows the Reynolds numbers at which the first unsteady flow patterns were detected at the inlet louver, turnaround louver, and exit louver for five configurations It can be seen that the fin pitch and fin angle both impact the transition to unsteady flow: e.g., decreasing the fin pitch postponed transition to much higher Reynolds numbers for a small fin angle As a reference, the experimental data of DeJong and Jacobi [21] for a slit fin and numerical data by Tafti and Zhang [22] for two louvered fin cases were added As can be seen, the transition in the inclined louvered fin configurations occurs at much lower Reynolds numbers It should be pointed out, however, that the slit fin configuration has a very low fin pitch ratio, which will further increase the velocity required for unsteady flow Tafti and Zhang [22] showed that both in louvered and slit fins the transition is due to the accumulation of perturbations in the flow The various louvers can be seen as individual roughness elements perturbing the flow as a fluid element passes over the fin The cumulative effect of these perturbations causes the flow to develop instabilities These instabilities manifest themselves around a louver as leading-edge shear layer instability or as wake instability The similarity between the transitional flow behavior Figure 15 Comparison between the flow visualization of the local flow behavior around the inlet louver and the numerical simulation of the streamlines, configuration 5, with ReLp = 402 heat transfer engineering vol 32 no 2011 186 C T’JOEN ET AL Figure 16 ReLp for which unsteady flow was found at the entry, turnaround, and exit louver for five inclined louvered fin configurations For comparison: experimental slit fin data from DeJong and Jacobi [21] and numerical louvered fin data from Tafti and Zhang [22] in standard louvered fins and slit fins is not unexpected Due to the high flow efficiency in standard louvered fins (Figure 4), the central core flow in standard louvered fins is very similar to the flow within a slit fin The transition in the inclined louvered fins is however not due to the accumulation of flow perturbations It is geometrically triggered by large recirculation zones that are present on the inclined parts of the inlet and turnaround louver These are clearly visible in Figure 13 At low ReLp , steady recirculation bubbles form, but as the velocity increases the shear layers trigger unsteadiness and a pattern of vortex shedding (Von Karman vortex street) becomes apparent Previously studied louvered fin designs all had a much smaller inclined section of the turnaround louver, as is clear from Figure 17 The data in Figure 16 also show that the unsteady flow behavior moves upstream a lot more rapidly in the inclined louvered fins than in the slit fins or standard louvered fins This again is due to the nature of the driving force In the standard louvered fins and slit fins the perturbations build up, resulting in a gradual shift to unsteady flow In the inclined louvered fins the recirculation zones become unsteady quite abruptly at a Reynolds number not much higher than that of the wake of the exit louver The results of this study show that it is possible to lower the Reynolds number at which unsteady flow is present in an interrupted fin design considerably by modifying the geometry This will result in increased heat transfer at low Reynolds numbers, which can be very significant for specific applications such as space heating and cooling (air conditioning) It was also found that the same geometric modification will determine the location of the onset of unsteady flow So adding additional turnaround louvers (W-shaped flow path in stead of a V-shaped one) will result in a larger part of the fin experiencing unsteady flow at larger Reynolds numbers These results have a practi- Figure 17 Standard louvered fin geometry as studied by Tafti and Zhang [22] heat transfer engineering cal significance for the further improvement of interrupted fin designs: the use of larger inclined sections on the turnaround louvers or multiple turnaround louvers can result in increased heat transfer This idea should be further studied and evaluated using heat transfer and pressure drop data combined with performance evaluation criteria for compact heat exchangers to better assess the potential However, it clearly must have some merit, as a commercially available fin type for air conditioning has a W shape and very large inclined sections, as shown by T’Joen et al [34] CONCLUSIONS The mean and transitional flow behavior within the inclined louvered fin (a hybrid of the slit fin and the standard louvered fin) has been studied experimentally through flow visualization in a water tunnel To quantify the flow behavior, the fin angle alignment factor ζ was introduced, which is closely related to the well-known flow efficiency η in standard louvered fins Using a graphical measurement technique, local ζ values were determined that showed considerable variation, depending on the location of the injection point The local data was used to validate CFD data Using the validated simulations it was shown that the graphical measurement method is sensitive to distortions due to local flow separation So care should be taken when choosing the reference points to determine the mean flow angle α The results of this study also showed that it is possible to lower the Reynolds number at which unsteady flow is present in an interrupted fin design considerably by modifying the geometry and at the same time determine where the unsteady flow occurs This will result in increased heat transfer at low Reynolds number, which can be very significant for specific applications such as air conditioning These results have a practical significance for the further improvement of interrupted fin designs: Use of larger inclined sections on the turnaround louvers vol 32 no 2011 C T’JOEN ET AL or multiple turnaround louvers (W-shaped fins) can result in increased heat transfer This idea should be further studied and evaluated using heat transfer and pressure drop data combined with performance evaluation criteria for compact heat exchangers to better assess the potential [6] NOMENCLATURE D Did Fp Lp N ReLp St t u v x, y x, y, z [5] actual transversal distance traveled by the flow, measured from the dye injection point—defined in Figure (m) ideal transversal distance traveled by the flow, measured from the dye injection point (fully aligned with the fin angle)—defined in Figure (m) fin pitch (m) louver pitch (m) number of louvers in the upstream half of the fin array (—) Reynolds number, based on the louver pitch and the mean inlet velocity (-) Stanton number (—) fin thickness (m) velocity in the x direction (m/s) velocity in the y direction (m/s) distance defining the steakline location in Figure (m) spatial coordinates (m) [7] [8] [9] [10] [11] [12] Greek Symbols α ζ θ η φ τ mean flow angle (◦ ) fin angle alignment factor (—) louver angle (◦ ) flow efficiency (—) fin angle (◦ ) number of fin passages a streakline crosses before reaching the center of the turnaround louver (—) [13] [14] [15] REFERENCES [1] Shah, R K., Heikal, M R., Thonon, B., and Tochon, P., Progress in The Numerical Analysis of Compact Heat Exchangers, Advances of Heat Transfer, vol 34, pp 363–443, 2001 [2] Tanaka, T., Itoh, M., Kudoh, M., and Tomita, A., Improvement of Compact Heat Exchangers With Inclined Louvered Fins, Bulletin of JSME, vol 27, pp 219–226, 1984 [3] Suzuki, K., Nishihara, A., Hayashi, T., Shuerger, M J., and Hayashi, M Heat Transfer Characteristics of a Two-Dimensional Model of a Parallel Louver Fin, Heat Transfer—Japanese Research, vol 19, pp 654–669, 1990 [4] Chang, Y J., and Wang, C C., A Generalized Heat Transfer Correlation for Louvered Fin Geometry, International heat transfer engineering [16] [17] [18] [19] 187 Journal of Heat and Mass Transfer, vol 40, pp 533–544, 1997 Chang, Y J., Hsu, K C., Lin, Y T., and Wang, C C., A Generalized Friction Correlation for Louver Fin Geometry, International Journal of Heat and Mass Transfer, vol 43, pp 2237–2243, 2000 Manglik, R., and Bergles, A., Heat Transfer and Pressure Drop Correlations for the Rectangular Offset Strip Fin Compact Heat Exchanger, Experimental Thermal and Fluid Science, vol 10, pp 171–180, 1995 Rose, J W., Heat Transfer Coefficients, Wilson Plots and Accuracy of Thermal Measurements, Experimental Thermal and Fluid Science, vol 28, pp 77–86, 2004 Taler, D., Prediction of Heat Transfer Correlations for Compact Heat Exchangers, Forschung Im IngenieurwesenEngineering Research, vol 69, pp 137–150, 2005 Beauvais, F N., An Aerodynamic Look at Automotive Radiators, Proceedings of the SAE Automobile Engineering Meeting, Detroit, MI, USA, 1965 Available online at: http://papers.sae.org/650470 SAE Paper No 650470, 1965 Davenport, C J., Heat Transfer And Fluid Flow in Louvred Triangular Ducts, Ph.D thesis, Lanchester Polytechnic, Lancaster, UK, 1980 Cowell, T A., Heikal, M R., and Achaichia, A., Flow and Heat Transfer in Compact Louvered Fin Surfaces, Experimental Thermal and Fluid Science, vol 10, pp 192–199, 1995 DeJong, N C., and, Jacobi, A M., Flow, Heat Transfer and Pressure Drop in the Near-Wall Region of Louvered-Fin Arrays, Experimental Thermal and Fluid Science, vol 27, pp 237–250, 2003 DeJong, N C., and Jacobi, A M., Localised Flow and Heat Transfer Interactions in Louvered Fins, International Journal of Heat and Mass Transfer, vol 46, pp 443–455, 2003 DeJong, N C., Flow, Heat Transfer and Pressure Drop Interactions in Louvered Fin Arrays, Ph.D, thesis, University of Illinois, Urbana, IL, 1999 Sahnoun, A., and Webb, R L., Prediction of Heat Transfer and Friction for the Louver Fin Geometry, Transactions of the ASME: Journal of Heat Transfer, vol 114, pp 893–900, 1992 Webb, R L., and Trauger, P., Flow Structure in the Louvered Fin Heat-Exchanger Geometry, Experimental Thermal and Fluid Science, vol 4, pp 205–217, 1991 Achaichia, A., and Cowell, T A., Heat Transfer and Pressure Drop Characteristics of Flat Tube and Louvered Plate Fin Surfaces, Experimental Thermal and Fluid Science, vol 1, pp 147–157, 1988 Zhang, X., and Tafti, D K., Flow Efficiency in MultiLouvered Fins, International Journal of Heat and Mass Transfer, vol 46, pp 1737–1750, 2003 Lyman, A C., Stephan, R A., Thole, K., Zhang, L W., and Memory, S B., Scaling of Heat Transfer Coefficients vol 32 no 2011 188 [20] [21] [22] [23] [24] [25] [26] [27] [28] [29] [30] [31] [32] [33] C T’JOEN ET AL Along Louvered Fins, Experimental Thermal and Fluid Science, vol 26, pp 547–563, 2002 Zhang, X., and Tafti, D K., Classification and Effects of Thermal Wakes on Heat Transfer in Multilouvered Fins, International Journal of Heat and Mass Transfer, vol 44, pp 2461–2473, 2001 DeJong, N C., and Jacobi, A M., An Experimental Study of Flow and Heat Transfer in Parallel-Plate Arrays: Local, Row-by-Row and Surface Average Behaviour, International Journal of Heat and Mass Transfer, vol 40, pp 1365–1378, 1997 Tafti, D K., and Zhang, X., Geometry Effects on Flow Transition in Multilouvered Fins—Onset, Propagation and Characteristic Frequencies, International Journal of Heat and Mass Transfer, vol 44, pp 4195–4210, 2001 Tafti, D K., Wang, G., and Lin, W., Flow Transition in a Multilouvered Fin Array, International Journal of Heat and Mass Transfer, vol 43, pp 901–919, 2000 Tafti, D K., Zhang, L W., and Wang, G., Time-Dependent Calculation Procedure for Fully Developed and Developing Flow and Heat Transfer in Louvered Fin Geometries, Numerical Heat Transfer Part A: Applications, vol 35, pp 225–249, 1999 Wang, C C., Lee, C J., Chang, C T., and Lin, S P., Heat Transfer and Friction Correlation for Compact Louvered Fin-and-Tube Heat Exchangers, International Journal of Heat and Mass Transfer, vol 42, pp 1945–1956, 1999 Malapure, V P., Mitra, S K., and Bhattacharya, A., Numerical Investigation of Fluid Flow and Heat Transfer Over Louvered Fins in Compact Heat Exchangers, International Journal of Thermal Sciences, vol 46, pp 199–211, 2007 T’Joen, C., Jacobi, A M., and De Paepe, M., Flow Visualization in Inclined Louvered Fins, Experimental and Thermal Fluid Science, vol 33, pp 664–674, 2009 T’Joen, C., Willockx, A., Steeman, H J., and De Paepe, M., Thermo-Hydraulic Characteristics of Inclined Louvered Fins, Proceedings of the 6th International Conference on Enhanced, Compact and Ultra-Compact Heat Exchangers, 16–21 September, Potsdam, Germany, 2007 Taylor, J R., An Introduction to Error Analysis, 2nd ed, University Science Books, Sausalito, CA, 1997 Tafti, D K., and Cui, J., Fin-Tube Junction Effects on Flow and Heat Transfer in Flat Tube Multilouvered Heat Exchangers, International Journal of Heat and Mass Transfer, vol 46, pp 2027–2038, 2003 Fluent, Inc., Fluent User’s Guide Version 6.2, Fluent, Inc., Lebanon, NH, 2005 T’Joen, C., Thermo-Hydraulic Study of Inclined Louvered Fins, Ph.D thesis, Ghent University-UGent, Belgium, 2008, http://hdl.handle.net/1854/LU-528875 Wartick, N W., and Jacobi, A M., An Experimental Study of Low-Reynolds Number Flow and Heat Transfer in an Array of Louvers at a Non-Zero Angle of Attack, Air Conditioning and Refrigeration Center CR-27, 2000 Available online at: http://acrc.mechse.illinois.edu heat transfer engineering [34] T’Joen, C., Steeman, H J., Willockx, A., and De Paepe, M., Determination of Heat Transfer and Friction Characteristics of an Adapted Inclined Louvered Fin, Experimental Thermal and Fluid Science, vol 30, pp 319–327, 2006 Christophe T’Joen is a postdoctoral researcher at the Department of Flow, Heat, and Combustion Mechanics, Ghent University, Belgium He received his master’s degree in electromechanical engineering from Ghent University in 2004 and his Ph.D in 2008 Currently his main research topic is interrupted fin designs for compact heat exchangers For his master’s dissertation he was honored with the Marcel Herman prize Huisseune Henk is a Ph.D student at the Department of Flow, Heat, and Combustion Mechanics, Ghent University, Belgium He received his master’s degree from Ghent University, Belgium, in 2007 For his master’s dissertation, he was honored with the Umicore Award Currently, he is doing research on compact heat exchangers, focused on tube–fin junction effects and vortex generators Arnout Willockx is a Ph.D student at the Department of Flow, Heat, and Combustion Mechanics, Ghent University, Belgium He received his master’s degree in mechanical engineering from Ghent University in 2003 His research is mainly focused on a study of fin effectiveness for different fin forms Hugo Cani`ere is a Ph.D student at the Department of Flow, Heat, and Combustion Mechanics, Ghent University, Belgium In 2005, he received his master’s degree in mechanical engineering from Ghent University Currently he is working on experimental two-phase flow and heat transfer of evaporating refrigerants Michel De Paepe is a professor of applied thermodynamics at Ghent University, Department of Flow, Heat, and Combustion Mechanics, Gent, Belgium, leading the Applied Thermodynamics and Heat Transfer Group He obtained his Ph.D from Ghent University in 1999 on the topic of the thermodynamics of steam-injected gas turbines and his master’s degree in mechanical engineering from the same university in 1995 For both his master’s dissertation and his Ph.D he was honored with the WEL Energy prize His research interests are energy performance of thermal systems in general A first focus is on thermodynamics simulation of fuel cells, cogeneration plants, and gas turbines A second research field is modeling of heat, air, and moisture transfer in buildings and HVAC installations Finally, his work is also dedicated to complex heat transfer phenomena in heat exchangers and two-phase flow He is the author of more than 50 research papers in journals and conference proceedings vol 32 no 2011