Technical Progress Report September 15,1996 to November 14,1996

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1 NATU'RAL CONVECTON HEAT EXCHANGERS FOR SOLAR WATER HEATNG SYSTEMS Technical Progress Report September 5,996 to November 4,996 Jane H. Davidson Department of Mechanical Engineering University of Minnesota Minneapolis, MN Prepared for the United States Department of Energy Under Contract No. DE-FG36-94G00030

2 .._ This report was prepared as an account of work sponsored by an agency of the United States Government. Neither the United States Government nor any agency thereof, nor any of their employees, makes any warranty, express or implied, or assumes any legal liability or responsibility for the accuracy, completeness, or usefulness of any information, apparatus, product, or proctss disclosed, or represents that its use would not infringe privately owned rights. Reference herein to any specific commercial product, process, or service by trade name, trademark, manufacturer, or otherwise does not necessarily constitute or imply its endorsement, m m mendation. or favoring by the United States Government or any agency thereof. The views and opinions of authors expressed herein do not necessarily state or reflect those of the United States Government or any agency thereof.

3 DSCLAMER Portions of this document may be illegible electronic image products. mages are produced from the best available original document.

4 NATURAL CONVECTON HEAT EXCHANGERS FOR SOLAR WATER HEATNG SYSTEMS DE-FG36-94GOl.0030 Jane H. Davidson University of Minnesota Reporting Period: Project ]Personnel: September 5,996 - November 4,996 Jane H. Davidson, Faculty; Scott Dahl, Ph.D. candidate Project Objectives: The goals of this project are ) to develop guidelines for the design and use of thermosyphon side-arm heat exchangers in solar domestic water heating systems and 2) to establish appropriate modeling and testing criteria for evaluating the performance of systems using this type of heat exchanger. The tasks for the project are as follows. ) Develop a model of the thermal performance of thermosyphon heat exchangers in solar water heating applications. A test protocol will be developed which minimizes the number of tests required to adequately account for mixed convection effects. The TRNSYS component model will be fully integrated in a system component model and will use data acquired with the specified test protocol. 2) Conduct a fundamental study to establish friction and heat transfer correlations for conditions and geometries typical of thermosyphon heat exchangers in solar systems. Data will be obtained as a function of a buoyancy parameter based on Grashof and Reynolds numbers. The experimental domain will encompass the ranges expected in solar water heating systems. Summary of Results Experimental study of tube-in-shell thermosyphon heat exchangers proves that these devices operate in the mixed convection regime on the thermosyphon side. Consequently, modeling and testing procedures must not assume that thermal performance is a forced convection phenomenon. We have developed testing and modeling procedures that rely on a correlation of UA with the relevant mixed convection dimensionless parameters (Re, Gr, Pr). A TRNSYS system model has been written based on these procedures. A fundamental study of pressure drop and heat transfer in side-arm heat exchangers is underway. Current Activities Task The colaborative effort with Steven Long of Florida Solar Energy Center to develop and evaluate a TRNSYS model of side-arm heat exchangers has been complete. Tow new TRNSYS modules have been developed. A paper describing this work is attached. Task 2 Modification of the experimental facility to install the new tube-in-shell heat exchanger is in progress. Efforts have focused on construction and instrumentation of the heat exchanger as well as modification of the computer data acquisition software. Future,4ctivities Experimental study with constant flux boundary conditions will begin in December.

5 TESTNG, MODELNG AND RATNG SOLAR WATER HEATERS WTH SDE-ARM THERMOSYPHON HEAT EXCHANGERS Jane H. Davidson and Scott D. Dah Department of Mechanical Engineering University of Minnesota Church St. S.E. Minneapolis, MN Steven Long Florida Solar Energy Center 679 Clearlake Rd. Cocoa, FL ABSTRACT ndirect solar water heating systems traditionally use mechanically-pumped heat exchangers. To reduce capital and maintenance CDS~Sas well as electrical demand, m a n u f a c m are considering heat exchangers that rely on thennosyphon flow rather than a pump in the water loop. A. testing protocol and a new TRNSYS model of these systems are presented. The heat exchangermodel accounts for the mixed convection heat transfer, and the thermosyphon loop model accuratt:ly predicts pressure drop. Comparison of the model to expeiriment shows excellent agreement. Daily and annual ratings are presented. An indirect sollat water heater with a thennosyphon heat exchanger is shown in Fig.. Analogous to conventional system, the "side-arm" heat exchanger is external to the tank,but flow in the water loop is buoyancy-driven rather than mechanically-pumped. Water flow rate varies as temperatures hi the collector and tank change. Because these heat exchangers operate in the mixed convection regime, UA is not a unique: function of system flow rates. Thus, testing and modeling quire differentprocedures than those used for systems with pumped flows. We present a testing protocol, and two TRNSYS subroutines that u:se the test data to predict system performance. Ratings for an hypothetical system with a heat exchanger tested at the University of Minnesota are computed under the Solar Rating and CertiFcation Corporation OG300 rating day (SRCC, 995) and for the Sacramento Municipal Utility District's (SMUD) rating year. Performance of this system is compared tci that of a pumped system. Fig. : Solar water heater with thermosyphon heat exchanger 2. MODELNG The TRNSYS model relies on empirical correlations: shear pressure drop across the heat exchanger as a function of thermosyphon flow rate, and UA as a function of Prandtl, Reynolds and Grashof numbers. Future versions of the model will account for the temperaturedependence of pressure drop. A flow chart of the TFWSYS model is shown in F'ig. 2. The major modification in this model as compared to the empirical model of Fraser ef al. (993) developed for WATSUN and adapted for TRNSYS by Avina et al. (995) is the heat transfer process in the heat exchanger. The new TRNSYS modules are Type 63: General Thermosyphon Loop and Type 72: Thermosyphon Heat Exchanger. 2. TvDe 63:General Thermoswhon Loop Type 63 calculates the instantaneous steady-state flow rate for a thermosyphon water loop. The loop is divided into components (e.g. heat exchanger, tank, piping). Hydrostatic

pressure (Le. tlemperature distribution) is input from the component models. Shear pressure drop may be specified by standard correlations or measured values. Using a -D rn.

6 pressure (Le. tlemperature distribution) is input from the component models. Shear pressure drop may be specified by standard correlations or measured values. Using a -D rn.*l, flow rate is determined by equating frictional losses to hydrostatic pressure difference: The right-handl-sideof Eq. () is the driving force for the thennosyphon flow. Frictional losses include those across the heat excbnger, pipes and fittings. Pressure drop correlationsused for pipes and fittings are given in the Appendix. Frictional losses in the tank are neglected. For a closed loop with N segments, Eq. () is expressed as: Temperature distribution in the tank is modeled with TRNSYS Type 4. The temperature distribution and hydrostatic pressure in the heat exchanger are determinedin the new Type 72. The flow rate is obtained by numerical solution of Eq. (2) using the Van Wijngaarden-Dekker-Brent method (Press et ai., 992). 2.2 Type 72: Thennosvphon Heat Exchanger Fluid properties are evaluated at (T,,,+TJZ. The Grashof number is based on inlet temperatures. Outlet temperatures are determined by converting UA to effectiveness using the NTU method. We added the expression for a two-pass tubein-shell heat exchanger with a single pass on the shell side (ncropera and DeWitt, 98). Once outlet temperaturesare determined, temperature distribution in the heat exchanger is determined from a user defined temperature profile and the hydrostatic pressure is computed. f no temperature profile shape is specified, a linear profile is assumed. One problem that arises is predicting the startup flow when the hydrostatic driving force is zero yet the collector-side fluid is warmer than the water in the heat exchanger. n this circumstance, we arbitrarily set the flow rate in the heat exchanger to 0. g/hr. 3. 3EsmKi Tests must be conducted at the desired collector flow mte with the antifreezemixture intended for use. The U of M test facility, shown in Fig. 3, consists of a simulated collector loop in which 50/50 ethylene glycol is heated with an electric boiler, and a water loop that includes the heat exchanger, water storageheating tank and a cold water supply. The 3 kw in-line electric boiler in the "collector" loop maintains a user selected constant outlet temperatwe (Fo.5 "C). Collector flow rate is controlled with a globe valve. The cold water supply is controlled to & *C. Type 72 models the heat transfer using a user specifed relationship for and Dahl, 996), UA P ~ P F Q P G ~ ' ~. (3) Fig. 3: University of Minnesota test facility. Calculate hydrastatic head in loop Solve for mc(eq. 2) using Brenrs method ; # Load # +,,* using succassive iteration (used with old TANSYS sober ) iterateon Fig. 2 Flow chart of TRNSYS model. The 246 liter water storage tank (49.2 cm high and 56.9 cm in diameter) has a 4500W electric element located 82 cm from the bottom. Tank UA was measured. The heat exchanger is a double-wall, two-pass tube-in-shell. The tube is helically ribbed on the interior, knurled on the exterior, and has a heat transfer area of 0.2 m2. Cool water enters the shell where the heat exchanger is connected to the 2

drain valve, 7.6 cm from the bottom of the tank. Heated water enters the tank at the temperature/pressurerelief valve port, 33 cm above the bottom of the tank.

7 drain valve, 7.6 cm from the bottom of the tank. Heated water enters the tank at the temperature/pressurerelief valve port, 33 cm above the bottom of the tank. Hot anti-freeze enters the heat exchanger at the top, flows through the twopass tube and ireturns to the boiler. Measured valu~esinclude collector-sideforced flow rate, temperatures at the inlet and outlet of the heat exchanger, temperature differences across the heat exchanger, vertical temperature distribution in the storage tank,and shear pressure drop on the thermosyphon si& of the heat exchanger. Temperatures are measured using type T thermocouples,(m.5 "C). Five junction thermopiles (?(0.5% of reading "C)) are used to measure temperature dilferences across the heat exchanger. 4. COMPARSON OF MODET, AND E X P W T The TRNSYS model was evaluated by comparing simulated and experimental values of thermosyphon flow rate, heat exchanger temperaturesand delivered energy for the described test. Modeling was performed at the Florida Solar Energy Center and experimentswere conducted at the University of Minnesota. As shown in Figs. 4 and 5, the predicted values are nearly identical to measured values (and well within experimentalerror). No adjustments to pressure drop were necessary to reach this agreement. 8 Y Collector-sidemass flow rate is measured with a Coriolis flow meter (*(OS% of the reading kg/s)). Thermosyphon mass flow rate is determined from an energy balance across the heat exchanger (estimated error is &4% except at very low flow rates when the error can be as high as eo%). The measured value of UA is determined using, P m a lo00 $ ep2 z P,500 o.oo0 Time, minutes UA- (&p),.,(th,i Fig. 4 Measured and predicted thennosyphon flow rate and energy during the two part test. -Th,o) ATLMTD 00 (5) Although shear pressure drop on the water side of the heat exchanger may be affected by heating, measurements obtained using forced flow without heating and with the heat exchanger oriented horizontally may not cause large errors. The difficulty of measuring shear pressure drop in heated vertical sections during simulated operation is discussed in Dah and Daviclson (996). Pressure drop measurements are made with a differential electronic aansducer (23 Pa). The test consists of two parts. During the first part, the tank is kept at '25 C while the temperature of the collector fluid is raised 5 C every 30 minutes from 25 to 95 C. Approximately 5minutes is required to reach steady-state. Thermosyphon flow rate increases as the driving force increasesdue to increasesin the collector outlet temperature. During the second part of the test, the tank is allowed to heat up and stratify. Collector outlet temperature is held at 95 C. As the hydrostatic pressure in the tank decreases, thennosyphon flow rate decreases Time. minutes Hg. 5: Measured and predicted fluid temperatures in the heat exchanger during the two part test. The slight difference in cold water inlet temperatures is probably due to imprecision in tank UA values and mixing in the storage tank.the selection of nodes is based on water flow rate following the recommendation of Kleinbach et al. (993). The use of more nodes is not warranted based on the uncertainty of how the tank performs during conditions not encountered during testing. 3

8 5. DALYS R:CC OG300 RATNG 70 The SRCC OG300 simulations (see SRCC, 995) are for an hypothetical system with the side-arm heat exchanger and for the same system with a pump added to the water loop. The system spcxifications are in Table. Neither system exists nor have: the systems been designed for optimum perfonnance. 60 0E 50 3 iij 40 L. al e z Efficiency UA (WC) Flow Rate Storage Pump Storage Tank 3m2 q W/m2.0C(AT/) W/m2-oC2fAT2m 0.03 kg/s kgls 25 W m3, 6 nodes Load PUMPED 43,300 kj 2,403 kj 2,897 kj Hour of Day Fig. 6: Predicted temperatures in thennosyphon system for OG300 rating. % 0.008t.. $ Y,,,..,.,,...,...,...,..,,,,., 6 OG300 ratings are given in Table 2. Temperatures across the thennosyphon heat exchanger are plotted in Fig. 6. Thennosyphon flow rate is plotted in Fig. 7. Temperatures across the pumped heat exchanger are shown in Fig. 8. The lower UA of the thennosyphon heat exchanger results in higher collector operating temperatures and lower energy transfer. ncreased stratification of the tank in the thermosyphon system results from modeling the tank with more nodes. Parasitic power is reduced by 5%by the elimination of the 25W pump. Solar delivery is reduced by 28%. A larger heat exchanger would improve the performance. TABLE2. OG300 RATNGS 0 a Pro,43Re0.'3GroJo Variable None m3 0 nodes 8 g s s 0.00 ; OF...! Hour of Day Fig. 7: Predicted thennosyphon mass flow rate for OG300 rating THERMOSYPHON 43,300 kj 5,356 kj 27,946 kj.5 Net Solar Auxiliary Solar Energy Factor*,834 kj Parasitic EnerjTy 2,62 kj * Energy delivered/(auxiliary+pat-asitic) s2 8 i;i Hour of Day Fig. 8: Predicted Temperatures in the pumped system for OG300 rating. 4

9 * 6. ANNUALSMULATlON T UA V SRCC ratings of performance under the SMUD specified year are compmd in Table 3. Thermal performance of the thermosyphon system is 9%lower. Parasitic energy is reduced by GlEd&ms P A E PUMPED THERMOSYPHON.rl P V P SubscriDts ParasiticEneigy 769 MJ 637MJ C h HX 0 S 7. SUMM A new TRNSJ S model for an indirect solar water heating system that uses a side-arm thermosyphon heat exchanger has been developed. Comparison of predicted temperatures, flow rates and energyto measured values validates the new thermosyphon loop and heat exchanger subroutines. The model correctly accounts for mixed convection heat transfer in the heat exchanger and pressure drop across the thermosyphon loop..although a comparison of a conventional pumped system and the thermosyphon system is included, the merits of the thermosyphon system should not be based on this comparison since neither system is optimized. The model can be used in future efforts to optimize system design. 8. NOMENCLATURE S X X ~ ~heat, ~ C J/(kg- C) hydraulic diameter, m Darcy friction factor acceleration of gravity, m/s2 Grashof number, gpd3(tbj-t$/v2 vlertical rise, m loss coefficient for minor losses eiermal conductivity,w/(m2.0c) length of flow, m log-mean-tempmedifference, C mass flow rate,kg/s number of segments used in TRNSYS pressure, Pa p i m e t e r constants, P, P2, P3,P4, Eq. (3) Prandtl number, c@ heat transfer, W Reynolds number, pvd/p tank temperature, C overall heat transfer coefficient-ma,w/ C average velocity, m/s vertical coordinate, m coefficientof expansion, PC m m pipe roughness, m efficiency viscosity, Pas kinematic viscosity, m2/s fluid density, kg/m3 cold hot heatexcbanger inlet outlet shear waterstoragetank 9. ACKNOWLEDGMENTS The support of this work by the United States Department of Energy, Micro-Motion nc., Heliodyne, nc. and the University of Minnesota is appreciated. 0. REFERENCES () Avina, J., Beckman, W. A., and Klein, S. A., 995, Simulation of a Natural Convection Heat Exchanger Solar Domestic Hot Water System, Proceedings, Solar 95, Annual Conference of the American Solar Energy Society, Minneapolis, MN, pp (2) Churchill, S. W., 977, Friction Factor Equations Spans all Fluid Flow Regimes, Chem Eng., Vol., pp (3) Davidson, J. H. and Dahl, S. D., 996, ssues of predicting Performance of ThermosyphonHeat Exchangers, Proceedings, ASME ntemtional Solar Energy Conference, San Antonio, TX, pp, -8. (4) Fraser, K. F., Hollands, K. G. T., and Brunger, A. P., 993, Validation of a Model for Natural Convection Heat Exchangers in SDHW Systems, Proceedings, Hurmony with Nature, SES Solar World Congress, Budapest, Hungary, pp (5) Hooper, W.B., 98, The Two k Method Predicts Head Losses in Pipe Fittings, Chem. Eng., pp (6) ncropera and DeWitt, 98, Fundamentals of Heat and Mass Transfer, Wiley &Sons. (7) Kleinbach, E.M., Beckman, W.A., and Klein, S.A., 993, Performance Study of One-DimensionalModels for 5

10 4.?, Stratifed Thermal Storage Tanks, Solar Energy, Vol. 50, NO. 2, pp (8) Langhaar, H.L., 942, Steady Flow in the Transition Length of a Straight Tube, J. Appl. Mech.,Vol. 9. pp (9) Press, W.H., Teukolsky, S.A., Vetterlking, W.T., Flannery, B. P., 992, Numerical Recipes in FORTRAN: The art of Sci*entificComputing, 2nd ed., University of Cambridge Press., pp (0) SRCC,.995, Operating Guidelines and Minimum Standards for Certifying. Solar Water Heating Systems: An Optional S W System Certification and Rating Program, SRCC, Washington, DC. Two methods are used to account for minor losses normally expressed in terms of K values. Correlations developed by Hooper (98) are used for bends and run-through tees.. APPENDX: C a l c m of M TABLE Al: VALUES 0F COEFFCENTS TO BE USED N HOOPER CORRELATON rrwsvsa e drop for Shear pressure drop in the water loop is the sum of the losses across the heat exchanger and the major and minor losses in the connecting piping: K + K2(+ d- ), K - (A@ Red where d is the inside diameter in inches and K, and K2depend on the type of fitting as shown in Table Al. Fitting Type 90 elbow 45 elbow Tee-elbow Kt K For the transitions to and from the tank K values are: AP, is determined empirically. For straight pipes, the friction factor.,f, is a function of the flow regime. n our experiments, the Reynolds number in the heat exchanger never exceeded 400. However, in the connecting pipes, Re may exceed 2300 and f must be estimated for the transition regime. A conelation by Churchill (977) that spans the laminar to turbulent region is used. f8[(l, + Red (A+B)3/2 From pipe to tank: From tank to pipe: K.0 K 6o/Red (A3 (A9 For other minor losses K values tabulated values for turbulent flow are doubled to account for the kinetic energy correction factor for laminar flow and low Reynolds number transition flows. Y2 ] Expansion, Where Contraction: K 0.42[-(d/62)~] (A0) where d, is the diameter of the smaller pipe in both expressions. For other non-standard piping connections, the user can specify R values. Pipe roughness E mm. Friction factor is adjusted for developing flow in the connecting pipes with (Langhaar, 942) 6

p?:uam SYSTEMS Technical Progress Report Jane H. Davidson NAllJRAL CONVECTION HEAT EXCHANGERS FOR SOLAR WATER HEATING February 1,1996 to March 31,1996

p?:uam SYSTEMS Technical Progress Report Jane H. Davidson NAllJRAL CONVECTION HEAT EXCHANGERS FOR SOLAR WATER HEATING February 1,1996 to March 31,1996 NAllJRAL CONVECTION HEAT EXCHANGERS FOR SOLAR WATER HEATING SYSTEMS Technical Progress Report February 1,1996 to March 31,1996 Jane H Davidson Department of Mechanical Engineering University of Minnesota