Process Heat Transfer Hof Master

This workbook contains a compilation of edited, and formatted valuable and practical "Tips" on HeatExchangers that have been published and offered to the engineering public by Mr. Dale Gulley, an experienced and recognized authority on Heat Exchanger Design, Fabrication, and Consulting. For some years , I have endeavored to collect as many of Dale's valued advice and "tips" as I possibly could. By doing so, I have gained further insight and knowledge by reading and applying his tips and know-how. Dale is not only an outstanding and recognized heat transfer expert, but he has been a contributing and positive member of The Tubular Exchanger Manufacturers' Association for many years, advocating the useful and positive efforts this organzatn has done for the engineerng profession world-wide. In the past 50 years I have arrived at many conclusions and results in dealing with the design, specification, fabrication, and operation of heat exchangers that are identical with Dale's Tips. My experience also coincides with that of a lot of my past and present engineering colleagues. My field experience has proven Dale's advice and Tips to be not only credible - but also valuable in applying heat transfer to process operations. I have put my effort into this compilation in order to make use of this valuable engineering know-how as a basis for Experienced Based Learning when dealing with heat exchangers. Through out this compilation, my personal notes on some of the Tips can be seen off the printed area of the worksheet and to the right-hand side. I have used this method to record my own experience related to the topic and to add empirical support and reinforcement to what Dale describes. Please note that I have used the following spreadsheet and workbook techniques to assist in employing the ideas and recommendations expounded by Dale: • The bulk of the Tips are organized in the same manner as they are found in Dale's Website. I have made use of Exel's Hyperlink feature to facilitate the quick and accurate access to any of the topics that are listed and grouped in the Table of Contents. Once you locate a subject or topic that you want to read or persue in the Table of Contents, all you have to do is click on the subject and the hyperlink will take you directly to the selected Tip. • I have made every effort to convert Dale's original presentation of recommended calculations and equations to a format that allows the reader to immediately employ his/her basic data to make the indicated calculation using Excel's basic spreadsheet feature. The reader can type in the basic data in the YELLOW-filled cells and the resultant calculated answer will be generated in BOLD RED numbers. This allows the reader to do several "what-if" calculations quickly to get an idea of the perceived effect on the heat exchanger. • The various groups of the Tips subject matter are also hyper-linked and a reader can go directly to one of the groups of Tips directly from the Table of Contents. These Tips are compiled and freely distributed with Dale Gulley's permission and approval. I would ask all engineers who are helped and assisted by this contribution to call or email Dale with thanks and gratitude for his contribution to heat exchange. Dale is active in heat exchange design, software, and process engineering out of Tulsa, Oklahoma. Needless to say, his organization can be of great help in a heat exchanger application. A Workbook Collection of Dale Gulley's Heat Exchanger Tips Art Montemayor - 05 April 2011 Exchangers that have been published and offered to the engineering public by Mr. Dale Gulley, an experienced Dale is not only an outstanding and recognized heat transfer expert, but he has been a contributing and positive fabrication, and operation of heat exchangers that are identical with Dale's Tips. My experience also coincides worksheet and to the right-hand side. I have used this method to record my own experience related to the topic engineers who are helped and assisted by this contribution to call or email Dale with thanks and gratitude for his out of Tulsa, Oklahoma. Needless to say, his organization can be of great help in a heat exchanger application. 4/26/2011 - Added Bs factor the second term in the denominator for the equation Chris Haslego for the seal bar calculation on sheet "Calculations". Cheresources.com Admin 6/28/2011 - Added three new tips from Gulleyassociates.com with permission. Chris Haslego Boiling Cheresources.com Admin Estimate - critical heat flux for propane chillers. Calculations Estimate - optimum flow velocity for gas inside tubes. Construction Longitudinal baffle heat conduction cures. 11/10/2011 - Added five new tips from Gulleyassociates.com with permission Chris Haslego Boiling Cheresources.com Admin Kettle Reboilers - Supports or Baffles Construction Design Temperatures of Carbon Steel and Low Alloy Tubes and Tubesheets Design Temperatures of Nonferrous Tubes and Tubesheets Misc. Fouling factors for water(hr-ft2-F/Btu) Fouling Factors for Liquid Hydrocarbons(hr-ft2-F/Btu) 4/5/2012 - Added five new tips from Gulleyassociates.com with permission Chris Haslego Boiling Cheresources.com Admin Vertical Thermosyphon-Calculate Pressure Drop at The Outlet Nozzle Vertical Thermosyphon-Design for a Smaller Liquid Preheat Zone Calculations Estimate - Hydrocarbon Gas Heat Transfer Coefficient in Shell Side Tube Bundle Vibration Best Design Feature to Prevent Bundle Vibration Misc. Viscous Flow - Use More Pressure Drop Than Usual Chris Haslego Cheresources.com Admin Chris Haslego Cheresources.com Admin Chris Haslego Cheresources.com Admin Design Temperatures of Carbon Steel and Low Alloy Tubes and Tubesheets Design Temperatures of Nonferrous Tubes and Tubesheets Chris Haslego Cheresources.com Admin Vertical Thermosyphon-Calculate Pressure Drop at The Outlet Nozzle Vertical Thermosyphon-Design for a Smaller Liquid Preheat Zone Estimate - Hydrocarbon Gas Heat Transfer Coefficient in Shell Side 2911 E. 77 Pl., Tulsa, OK 74136 · P.O. Box 700295, Tulsa, OK. 74170-0295 Phone: (918) 744-0100 Air Coolers: 1. Air flow accessories - don't overlook these when calculating fan HP 2. Box header design - limit of process temperature change 3. Connecting bundles of existing coolers for a new service 4. Fan drive changes that increase capacity of existing cooler 5. Fan drive noise - suggestions on how to reduce 6. Maximum motor HP for a fan 7. Maximum tube wall temperature for wrap-on fins 8. Optimum number of tube rows 9. Overall heat transfer rate estimate for hydrocarbons 10. When do bare tubes become more efficient than fin tubes? 11. When To limit number of tube passes in air coolers 12. When to use wind coolers Boiling: 1. Avoid mist flow boiling inside tubes 2. Kettle reboiler - liquid carryover problem solutions 3. Kettle reboiler - shell nozzle arrangement problem 4. Kettle reboiler - shell vapor outlet nozzle location 5. Kettle reboiler - sizing shell vapor space 6. Kettle reboiler - undersized shell effects 7. Estimate - pool boiling heat transfer coefficient for hydrocarbons 8. Large boiling temperature difference problems 9. Lowest limit of boiling temperature difference 10. Vertical thermosyphon - choking two phase flow with small outlet nozzle 11. Vertical thermosyphon - minimum recirculation rate 12. Vertical thermosyphon - check for liquid preheat zone 13. Vertical thermosyphon - who sets recirculation rate 14. Vertical Thermosyphon-Calculate Pressure Drop at The Outlet Nozzle 15. Vertical Thermosyphon-Design for a Smaller Liquid Preheat Zone Calculations 1. What diameter to use to start design of a coil 2. Estimate - gas heat transfer coefficient inside tubes 3. Estimate - hydrocarbon heat transfer coefficient in tubes 4. Estimate - latent heat of hydrocarbons 5. Estimate - liquid thermal conductivity of light hydrocarbons 6. Estimate - overall heat transfer coefficient in shell & tube 7. Estimate - tube length that lowers tube pressure drop 8. How to calculate excess surface and overdesign surface 9. Use superficial velocities to calculate best heat transfer flow pattern 10. L/D equation for heat Transfer coefficient inside tubing 11. LMTD correction factor charts for TEMA G and J type shells 12. Low LMTD correction factor for divided flow 13. What is the lowest LMTD correction to use in shell & tube 14. Minimum flow area for shell side inlet nozzle 15. How to calculate performance of heat exchangers with plugged tubes 16. How to increase heat transfer for low Reynolds numbers 17. Calculate when to use seal bars on a bundle to increase heat transfer 18. Calculate S & T bundle diameter from number of tubes 19. Equation for calculating tube count in shell & tube 20. Check for hot tube wall temperature of cooling water 21. Sometimes larger tubes are better than small ones 22. Weighted MTD 23. Estimate - optimum flow velocity for gas inside tubes. 24. Estimate - Hydrocarbon Gas Heat Transfer Coefficient in Shell Side Condensing: 1. Avoid small baffle cuts in S & T condensers 2. Estimate - Condensing heat transfer coefficient for hydrocarbons inside tubing 3. Maximum heat transfer rate inside tubes for total condensation 4. Quick estimate for reflux condenser LMTD in air cooler 5. Reflux (Knockback) condenser comments 6. Steam condenser types 7. Sulfur condenser - design within tube velocity limits 8. Warning about small temperature pinch points in condensers 9. When to slope single tube pass tubes in condensing service 10. Zone those condensers! 11. Estimate - critical heat flux for propane chillers. Construction: 1. Benefits of using rotated square pitch in shell & tube 2. Caution when using a longitudinal baffle in the shell side 3. Using turbulators for tube side laminar flow 4. Discussion of types of triple segmental baffles in shell & tube 5. Check entrance and exit space for shell nozzles 6. Horizontal vs vertical baffle cut in shell & tube 7. Is expansion joint required in the shell of a fixed tube sheet? 8. Increasing capacity of existing shell & tube exchangers 9. Locating vents on the shell side of vertical exchangers 10. Optimun gasket location for flanges 11. Reinforcing rods as tube inserts to increase heat transfer 12. Shell side impingement protection 13. Special shell & tube heat exchanger type (NTIW) 14. When to consider by-pass strips in shell & tube bundle 15. What is too large of temperature change in 2 tube passes ? 16. When to rotate square tube pitch in shell & tube exchanger 17. Longitudinal baffle heat conduction cures. Heat Recovery: 1. Deciding on what fin spacing to use 2. Estimate of nozzle size for HRSG 3. Face area estimate for HRSG units 4. Maximum exhaust gas temperaure for steel fin tubes 5. When to use bare tubes in waste heat boilers Materials: 1. Cooling water flowing inside 304SS U-tubes Pressure Drop: 1. Allowing for fouling in pressure drop calculations 2. Allowable pressure drop suggestions 3. Allowable shell side pressure drop if a multi-leaf(a.k.a. lamaflex) long baffle is used 4. Better baffle window pressure drop equation 5. Designing for better use of tube pressure drop 6. Effect of 1st tube rows on shell nozzle pressure drop 7. Pressure drop on kettle side 8. Reducing high shell side pressure drop in fixed tube sheet exchangers 9. Use impingement rods instead of plate to lower shell press. drop 10. What design pressure drop to use for heavy liquids inside tubes 11. Maximum velocity inside tubes 12. Calculate shell nozzle pressure drop 13. Improve shell side pressure drop calculations Tube Bundle Vibration: 1. Features of a new S & T bundle that replaces bundle that vibrated 2. Vibration cure when designing shell & tube bundles 3. Conditions likely to cause shell & tube bundle vibration 4. Cures for vibration in existing bundle 5. Best Design Feature to Prevent Bundle Vibration Miscellaneous: 1. Allocation of streams in shell & tube 2. Articles published by Dale Gulley 3. Avoid these fluids when using lowfin tubing 4. Best heat transfer flow pattern 5. Check liquid thermal conductivity at high reduced temperatures 6. Check piping connections when there is under-performance 7. Evaluating an exchanger for a new service 8. Check heat release curve data for skipping over dewpoints and bubblepoints 9. When will exchangers with low-fins be more economical than exchangers with bare tubes? 10. Problems with excess heat exchanger surface 11. Purchasing warning for shell & tube exchangers 12. What is the minimum velocity inside tubing for slurries? 13. Suggestions for low-fins and potential S & T bundle vibration 14. Choose shell & tube or multi-tube heat exchangers 15. Thermal design problem with shell side long baffle 16. Trouble shooting article in Hydrocarbon Processing 17. Under-surfaced S&T quote 18. When to add shell in Series 19. When to consider a long baffle in the shell 20. Which stream goes inside the tubes of gas/gas exchangers? 21. Weighted MTD 22. Why did performance decline in a TEMA type F,G or H type shell? 23. Zone those condensers 24. Viscous Flow - Use More Pressure Drop Than Usual Note: Input data into YELLOW cells and receive output in BOLD RED Air flow accessories - don't overlook louvers and screens when calculating fan HP Air static pressure loss is used to calculate the horsepower required for fans used in process air coolers. Charts and equations in the literature are usually for the tube bundle only. Frequently, air coolers have accessories like louvers and fan guards. They may also have hail, bug, or lint screens. Don't overlook the accessory pressure drop because they can increase the static pressure as much as 25%. Box header design - limit of process temperature change March, 1998 In the design of an Air cooled heat exchanger, avoid imposing too large a temperature change in the box headers. Too much temperature drop between the inlet and outlet tube passes can cause leakage where the tubes meet the tubesheet. If the temperature change of the tube side stream is over approximately 400 o F, then use a split header design. This allows a hot top section to slide past a cooler bottom section. Connecting Bundles of Existing Coolers for a new Service April, 1998 When re-using air cooled exchangers in a new service, don't overlook connecting the bundles in a series-parallel arrangement. New air coolers nearly always have the bundle connected in parallel. Arrange the bundles for more series type flow to increase the tube side velocity and get higher heat transfer rates. For example, an air cooler with six bundles could be arranged with four bundles in parallel, connected to two bundles in series. The two series bundles would handle the coldest part of the heat load where higher velocity is needed the most. Increase Capacity of Existing Air Cooler with Fan Drive Changes October, 1997 If you need to increase the capacity of an air cooler, don't junk it for a new one until you have exhausted the possibilities on changing the fan and the fan motor. The least expensive change is to increase the fan blade angle if it will not overload the motor. But check to make sure the blade angle is not already at the maximum. The next best change in terms of cost is to increase the fan speed by changing the drive ratio between the fan and the motor. If these changes are not enough you could increase the motor size or change the fan for one with more blades. Suggestions to Reduce Fan Drive Noise The most effective solution is to reduce the fan speed by changing the drive ratio between the fan and the motor. Other suggestions are to reduce the fan blade angle or change to a fan with more blades. Maximum Motor HP for a Fan Adding more HP to a fan will only work up to a point. The fan efficiency reaches a peak. Then increasing the HP will produce no more air. An estimate for this HP is: Max HP = 46.4 HP Fan Diam = 7.00 feet This is for fan diameters greater than 3.5 ft. 17 + 8.4 (Fan Diam - 3.5) = Temperature Limit of Wrap-On Fins for Aircoolers June, 2000 Above a certain temperature, it will be too hot for wrap-on fins. Due to thermal expansion, the aluminum fins will lose good contact with the tubing. In this case an integral type fin tube should be used. The summer time air outlet temperature is a very rough approximation. To be more exact, the tube wall temperature needs to be calculated for the hottest tube row. Then: T wall = Ta + (Th 1 - Ta) x Ro x Uc = 459 o F Where T wall = temperature of tube wall Ta = air outlet temperature = 200 o F Th 1 = temperature inside tube = 488 o F R o = thermal resistance of air = 0.12 hr-ft 2 - o F/Btu U C = clean overall heat transfer coefficient = 7.5 Btu/hr-ft 2 - o F Example: Steam is condensing at 488 o F. Assume that the U C is 7.5 and R o is 0.12. If the air outlet temperature is 200 o F, then: T wall = 200 + (488 - 200) x 0.12 x 7.5 T wall = 459 o F As you can see, the problem is more severe at high heat transfer rates. Not even the aircooled manufacturers agree exactly what this maximum tube wall temperature should be. The ASME code for allowable stress of aluminum has a maximum temperature of 400 o F. I believe this is the upper limit. Then the above example is operating too hot for wrap-on fins. Optimun Number of Tube Rows The optimum number of tube rows is a function of the maximum acceptable temperature rise of the air side. There are three limitations and the smallest air rise of the three should be used. The limitations are: 1 Limit the LMTD correction factor to a minimum of 0.9 for one tube pass - maximum air outlet temperature to be the same as the process side outlet temperature. 2 Minimum temperature difference at the hot end to be 8 to 10 o F. 3 Maximum air outlet temperature to be 300 o F if tension wound fins are used. Hydrocarbon U Estimate (Air-Coolers) February, 2002 In the preliminary design or checking of process air-coolers you need an estimate of the overall heat transfer coefficient (U). An estimate that is based on fin surface can be made from the following: Where OP is the operating pressure in PSIA When do Bare Tubes become More Efficient Than Fin Tubes? If the inside heat transfer coefficient beomes too low, fin tubes can become inefficient. This can be the case in heavy oil coolers. If it is expected that the heat transfer coefficient is below approximately 20 Btu/hr-ft 2 - o F, investigate both bare and fin tubes. When To Limit Tube Passes in an Aircooler November, 1999 For tube side streams that have a high heat transfer coefficient, it is probably not advantageous to use more than two tube passes. This would be for condensing streams like ammonia and steam. This could also be true for high thermal conductivity liquid streams if the LMTD is high. The velocity on these type of streams will have a minor effect on the overall heat transfer coefficient in the typical aircooler. The major thermal resistance is the air side heat transfer coefficient. Air Cooler Using Wind December, 2000 Where cooling water is not available and the outlet temperature is not critical, an air cooler can be built that depends only on the wind for cooling. It will have the best performance when the tubes have high fins and the tubes are perpendicular to the wind direction. In areas where the wind does not have a prevailing direction, arrange the tubes in a bird cage type pattern. Then there is cooling no matter which way the wind blows. If there is a prevailing wind direction, use an air cooler bundle that sets on a stand that faces the wind. Gases Overall Heat Transfer Coefficient Rt = 0.29 x Sqrt (100/OP) + 0.145 Fluid in Tube side Where: viscosity is less than 3 cP. Rt = 0.165 x Sqrt (avg. tube viscosity) + 0.145 U = 1/Rt U = 1/Rt Liquids March, 2000 Air static pressure loss is used to calculate the horsepower required for fans used in process air coolers. Charts and equations in the literature are usually for the tube bundle only. Frequently, air coolers have accessories like louvers and fan guards. They may also have hail, bug, or lint screens. Don't overlook the accessory pressure drop because In the design of an Air cooled heat exchanger, avoid imposing too large a temperature change in the box headers. tubesheet. If the temperature change of the tube side stream is over approximately 400 o F, then use a split header arrangement. New air coolers nearly always have the bundle connected in parallel. Arrange the bundles for more series type flow to increase the tube side velocity and get higher heat transfer rates. For example, an air cooler with possibilities on changing the fan and the fan motor. The least expensive change is to increase the fan blade angle if it will not overload the motor. But check to make sure the blade angle is not already at the maximum. The next best Adding more HP to a fan will only work up to a point. The fan efficiency reaches a peak. Then increasing the HP Above a certain temperature, it will be too hot for wrap-on fins. Due to thermal expansion, the aluminum fins will lose good contact with the tubing. In this case an integral type fin tube should be used. The summer time air outlet temperature is a very rough approximation. To be more exact, the tube wall temperature needs to be calculated for As you can see, the problem is more severe at high heat transfer rates. Not even the aircooled manufacturers agree exactly what this maximum tube wall temperature should be. The ASME code for allowable stress of aluminum has a maximum temperature of 400 o F. I believe this is the upper limit. Then the above example is operating too hot for The optimum number of tube rows is a function of the maximum acceptable temperature rise of the air side. There two tube passes. This would be for condensing streams like ammonia and steam. This could also be true for high thermal conductivity liquid streams if the LMTD is high. The velocity on these type of streams will have a minor arrange the tubes in a bird cage type pattern. Then there is cooling no matter which way the wind blows. If there Mist Flow Boiling Inside Tubes November, 2001 This is a flow pattern to avoid in heat transfer. The mist flow region is dependent upon velocity, % vapor and stratification effects. In this type of flow the tube wall is mostly dry and the liquid droplets are carried along in a vapor core. Therefore the heat transfer is much lower because the much higher thermal conductivity of the liquid is in very little contact with the tube wall. The higher the % vaporization, the lower the velocity needs to be to avoid mist flow. For example in a vertical tube where the vaporization is 50 % and the vapor density is 1.0 lb/cu ft, the velocity needs to be below approximately 80 ft/sec. If the vaporization is 75 %, the maximum velocity is approximately 30 ft/sec. This comes from the Fair equation. In a horizontal tube where there can be stratification, these maximum velocities are much lower. If the mist flow region cannot be avoided, then twisted tape turbulators can be used to increase the heat transfer. They will throw the liquid in the vapor core toward the tube wall. Kettle Reboiler - Location of Vapor Outlet Nozzles When it is necessary to have dry vapor leaving the kettle side, the location of the nozzles is important. The inlet nozzle should not be located directly under the vapor outlet. This probably results in some liquid carryover. When there is a single vapor outlet, it is usually centered over the bundle with the inlet nozzle located some distance away. There have been cases where someone other than the thermal designer changed the location of this vapor nozzle without the thermal designers OK. In one case the vapor outlet was moved to the back of the kettle resulting in appreciable liquid carryover Kettle Reboiler - Problem Shell Nozzle Arrangement Sometimes you see kettle reboilers where the inlet nozzle is directly under the outlet vapor nozzle. This arrangement creates extra turbulence under the vapor nozzle which affects the amount of liquid entrainment in the outlet vapor. It is safer to use the conventional nozzle arrangement where the inlet is some lateral distance away unless a demister pad is used. Another problem with the vertical nozzle arrangement is when the kettle bundle is relatively long and there is a single pair of nozzles. Then there is no good flow distribution. The boiling zones near the ends of the bundle will have lower fluid circulation rates and lower heat transfer. Kettle Reboiler - Location of Vapor Outlet Nozzles October, 2000 When it is necessary to have dry vapor leaving the kettle side, the location of the nozzles is important. The inlet nozzle should not be located directly under the vapor outlet. This probably results in some liquid carryover. When there is a single vapor outlet, it is usually centered over the bundle with the inlet nozzle located some distance away. There have been cases where someone other than the thermal designer changed the location of this vapor nozzle without the thermal designers OK. In one case the vapor outlet was moved to the back of the kettle resulting in appreciable liquid carryover Sizing the Vapor Space in Kettle Reboilers June, 1998 The size of the kettle is determined by several factors. One factor is to provide enough space to slow the vapor velocity down enough for nearly all the liquid droplets to fall back down by gravity to the boiling surface. The amount of entrainment separation to design for depends on the nature of the vapor destination. A distillation tower with a large disengaging space, low tower efficiency and high reflux rate does not require as much kettle vapor space as normal. Normally, the vapor outlet is centered over the bundle. Then the vapor comes from two different directions as it approaches the outlet nozzle. Only in rare cases are these two vapor streams equal in quantity. A simplification that has been extensively used is to assume the highest vapor flow is 60% of the total. One case where this would cause an undersized vapor space is when there is a much larger temperature difference at one end of the kettle then the other. The minimum height of the vapor space is typically 8 inches. It is higher for high heat flux kettles. Kettle Reboiler - Effect of Undersized Kettle Diameter July, 1997 What effect will an undersized kettle diameter have? The effect will be a decrease in the boiling coefficient. A boiling coefficient depends on a nucleate boiling component and a two-phase component that depends on the recirculation rate. An undersized kettle will not have enough space at the sides of the bundle for good recirculation. Another effect is high entrainment or even a two-phase mixture going back to the tower. Estimate - Pool Boiling Heat Transfer Coefficient for Hydrocarbons Boil h = 22 (Δt) 1.25 = 2,925 Btu/(hr)(ft 2 )( o F) Where Δt = (tube wall temperature - liquid temperature) = 50 o F t = temperature, o F Large Boiling Temperature Differences March, 1999 Large temperature differences in heat exchangers where liquid is vaporized are a warning flag. When the temperature differences reach a certain value, the cooler liquid can no longer reach the heating surface because of a vapor film. This is called film boiling. In this condition, the heat transfer deteriorates because of the lower thermal conductivity of the vapor. If a design analysis shows that the temperature difference is close to causing film boiling, the vaporizer should be started with the boiling side full of relatively cooler liquid. This way, you don't start flashing the liquid. The liquid is slowly heated up to a more stable condition. If the vaporizer is steam heated, the steam pressure should be reduced which will reduce the temperature difference. With steam heating, take a close look at the design if the LMTD is over 90 o F. This is close to the critical temperature difference where film boiling will start. Lower Limit of Boiling Film Temperature Difference February, 1997 A reboiler or chiller is best designed so that it doesn't have the lower heat transfer mode of natural convection. The dividing line between natural convection and boiling depends on the type of tubing used. If steel bare tubes are used, the lower limit of temperature difference between the tube wall and the boiling fluid is approximately 5 o F. We have designed hydrocarbon chillers down to the temperature difference of 2 o F using low-finned tubes. Special enhanced tube surfaces can be used for even lower temperature differences than 2 o F. Choking a Vertical Thermosyphon December, 1999 Choking down on the channel outlet nozzle and piping reduces the circulation rate through a heat exchanger. Since the tubeside heat transfer rate depends on velocity, the heat transfer is lower at reduced recirculation rates. A rule of thumb says that the inside flow area of the channel outlet nozzle and piping should be the same as the flow area inside the tubing. The Shell Oil Company, in an experimental study, showed that a ratio of 0.7 in nozzle flow area/tube flow area reduced the heat flux by 10%. A ratio of 0.4 cut the heat flux almost in half. An approximate equation for the amount of heat flux reduction is: Reduction = 3.06X -1.63X 2 - 0.43 = 53.32% Where X = area ratio= 0.40 Minimum Recirculation Rate in Thermosyphon Reboilers When does a recirculation rate become too low (high % vaporization)? When this happens, the tube wall is no longer wet and the heat transfer diminishes. The guidelines in the literature show the lowest permissible recirculation rates give from 25 to 40% vaporization for hydrocarbons. It has been observed that this threshold is when the outlet two-phase density (volume basis) is below 1.0 lb/cu-ft. Nearly all thermosyphons have outlet densities above this value. Vertical Thermosyphon - Check for Liquid Preheat Zone February, 2001 When designing vertical thermosyphon reboilers with boiling at low operating process fluid pressures, check for the presence of a liquid preheat zone. Back pressure raises the boiling point at the interface of liquid preheat zone and subcooled boiling. This boiling point rise creates a liquid zone with relatively low heat transfer and it reduces the temperature driving force (MTD). If the operating pressure is below approximately 25 PSIA, there should be a liquid preheat zone. The lower the operating pressure, the more likely there is liquid preheat. If there is no liquid preheat, there may be an input error. Vertical Thermosyphon Recirculation Rate December, 1997 In the design of vertical thermosyphons, the recirculation rate should be set by the process engineer if there will be anything unusual about the connecting piping. The recirculation rate is especially sensitive to the size and configuration of the outlet piping. If the recirculation rate is left for the thermal designer to set, they will have to make piping assumptions that may be violated later in the actual installation. Estimate - Critical Heat Flux For Propane Chillers A simple equation is presented for a kettle reboiler. It is conservative for very small bundles. The crital heat flux depends on the geometry of the bundle. The following estimate is based on 3/4 inch tubes on 15/16 inch pitch. It is actually good for any tube diameter with a tube pitch/tube diameter ratio of 1.25 and triangular tube pitch. A boiling temperature of -30 F. is assumed for the propane. CHF = 32500 = 12844 Btu/h ft2 Ds (0.25) CHF = crital heat flux in Btu/(hr)(ft)2 Ds = shell bundle diameter in inches = 41 Example What is the critical heat flux for a 41 inch diameter bundle? CHF = 32500 (41) 0.25 CHF = 12,850 Kettle Reboilers - Support or Baffles? For kettle reboilers use segmental baffles instead of full supports if shell fouling factor is greater Than 0.002(hr-ft 2 -F/Btu) Vertical Thermosyphon-Calculate Pressure Drop at The Outlet Nozzle A rule of thumb is that the pressure drop at the outlet nozzle should not be greater than 30% of the total static head. There is another tip in this boiling section about choking the flow with a small outlet nozzle. The inside flow area of the outlet nozzle should be the same or greater than the total flow area inside the tubing. For a channel with a side outlet the pressure drop is composed of a turning loss and a contraction loss The following equations calculate the pressure drop at the outlet. The pressure drop for expansion into the channel is not included here but is with the tube pressure drop. Ktr = ___1______ = 0.445576 (used for pressure drop calc) Ds 0.3 (If Ktr less than 0.40, use 0.40) Kc =0.5 (1 - (No/Ds)2) = 0.28077 KT = Ktr + Kc = 0.726346 ΔPn = KT = 0.000108 x Vn 2 x ρtp = 0.16944 Where: Ds = Top channel ID (inches) = 14.8 Ktr = pressure loss coefficient for turning loss = 0.445576 (calculated) Kc = pressure loss coefficient for contraction into nozzle = 0.28077 KT = total pressure loss coefficient = 0.726346 No = Outlet nozzle ID (inches) = 9.8 Vn = velocity thru nozzle (ft/sec) = 120 ρtp = two-phase density (lb/ft3) = 0.15 ΔPn = pressure drop thru channel and outlet nozzle (Psi) = 0.17 Vertical Thermosyphon-Design for a Smaller Liquid Preheat Zone At low operating pressures there will be a sensible heat liquid zone with relatively low heat transfer. This is caused by the fact that a small pressure change will cause a large increase in the boiling point. There has been a case where 90% of the tube length was in the sub-cooled phase. What can you change that will decrease the size of the liquid preheat zone and increase the overall heat transfer? One answer is to evaluate the piping system above the top tubesheet. In order to make an evaluation check the pressure drop at the outlet. There is on this section of the website equations to calculate the pressure drop of a nozzle that is at right angle to the top channel. Most vertical thermosyphons have the outlet nozzle at right angles to the top channel. There may be a simple change of enlarging the outlet nozzle that would be the cure. But there needs to be a check to make sure the nozzle and connecting piping are not so large that there is liquid slip. If enlarging the right angle nozzle and piping is not the answer then there are other configerations that will use less outlet pressure drop. Next the pressure drop of using a B type channel with a long radius ell could be tried. If this doesn't do it, try a mitered channel design. Another solution to the problem is to investigate inserts such as swisted tape, wire matrix , or helically coiled. This is a flow pattern to avoid in heat transfer. The mist flow region is dependent upon velocity, % vapor and stratification effects. In this type of flow the tube wall is mostly dry and the liquid droplets are carried along in a vapor core. Therefore the heat transfer is much lower because the much higher thermal conductivity of the liquid is in very little contact with the tube wall. The higher the % vaporization, the lower the velocity needs to be to avoid mist flow. For example in a vertical tube where the vaporization is 50 % and the vapor density is 1.0 lb/cu ft, the velocity needs to be below approximately 80 ft/sec. If the vaporization is 75 %, the maximum velocity is approximately 30 ft/sec. This comes from the Fair equation. In a horizontal tube where there can be stratification, these maximum velocities are much lower. If the mist flow region cannot be avoided, then twisted tape turbulators can be used to increase the heat Art's Note: When it is necessary to have dry vapor leaving the kettle side, the location of the nozzles is important. The inlet nozzle I agree. I have also found that locating the inlet liquid nozzle directly under the vapor outlet is not good. should not be located directly under the vapor outlet. This probably results in some liquid carryover. When there is In Amine BKU reboilers I found that locating the inlet rich amine liquid as close to the U-tube bundle tubesheet a single vapor outlet, it is usually centered over the bundle with the inlet nozzle located some distance away. There gave the best, consistant results in obtaining good solution stripping. This gives the heating medium have been cases where someone other than the thermal designer changed the location of this vapor nozzle without the thermal designers OK. In one case the vapor outlet was moved to the back of the kettle resulting in appreciable Sometimes you see kettle reboilers where the inlet nozzle is directly under the outlet vapor nozzle. This arrangement creates extra turbulence under the vapor nozzle which affects the amount of liquid entrainment in the outlet vapor. It is safer to use the conventional nozzle arrangement where the inlet is some lateral distance away unless a demister Another problem with the vertical nozzle arrangement is when the kettle bundle is relatively long and there is a single pair of nozzles. Then there is no good flow distribution. The boiling zones near the ends of the bundle will When it is necessary to have dry vapor leaving the kettle side, the location of the nozzles is important. The inlet nozzle should not be located directly under the vapor outlet. This probably results in some liquid carryover. When there is a single vapor outlet, it is usually centered over the bundle with the inlet nozzle located some distance away. There have been cases where someone other than the thermal designer changed the location of this vapor nozzle without the thermal designers OK. In one case the vapor outlet was moved to the back of the kettle resulting in The size of the kettle is determined by several factors. One factor is to provide enough space to slow the vapor velocity down enough for nearly all the liquid droplets to fall back down by gravity to the boiling surface. The amount of entrainment separation to design for depends on the nature of the vapor destination. A distillation tower with a large disengaging space, low tower efficiency and high reflux rate does not require as much kettle vapor space as normal. Normally, the vapor outlet is centered over the bundle. Then the vapor comes from two different directions as it approaches the outlet nozzle. Only in rare cases are these two vapor streams equal in quantity. A simplification that has been extensively used is to assume the highest vapor flow is 60% of the total. One case where this would cause an undersized vapor space is when there is a much larger temperature difference at one end of the kettle then the other. The minimum height of the vapor space is typically 8 inches. It is higher for high heat What effect will an undersized kettle diameter have? The effect will be a decrease in the boiling coefficient. A boiling coefficient depends on a nucleate boiling component and a two-phase component that depends on the recirculation rate. An undersized kettle will not have enough space at the sides of the bundle for good recirculation. Another Large temperature differences in heat exchangers where liquid is vaporized are a warning flag. When the temperature differences reach a certain value, the cooler liquid can no longer reach the heating surface because of a vapor film. This is called film boiling. In this condition, the heat transfer deteriorates because of the lower thermal conductivity of the vapor. If a design analysis shows that the temperature difference is close to causing film boiling, the vaporizer should be started with the boiling side full of relatively cooler liquid. This way, you don't start flashing the liquid. The liquid is slowly heated up to a more stable condition. If the vaporizer is steam heated, the steam pressure should be reduced which will reduce the temperature difference. With steam heating, take a close look at the design A reboiler or chiller is best designed so that it doesn't have the lower heat transfer mode of natural convection. The dividing line between natural convection and boiling depends on the type of tubing used. If steel bare tubes are used, the lower limit of temperature difference between the tube wall and the boiling fluid is approximately 5 o F. We have designed hydrocarbon chillers down to the temperature difference of 2 o F using low-finned tubes. Special enhanced Choking down on the channel outlet nozzle and piping reduces the circulation rate through a heat exchanger. Since the tubeside heat transfer rate depends on velocity, the heat transfer is lower at reduced recirculation rates. A rule of thumb says that the inside flow area of the channel outlet nozzle and piping should be the same as the flow area inside the tubing. The Shell Oil Company, in an experimental study, showed that a ratio of 0.7 in nozzle flow When does a recirculation rate become too low (high % vaporization)? When this happens, the tube wall is no longer wet and the heat transfer diminishes. The guidelines in the literature show the lowest permissible recirculation rates give from 25 to 40% vaporization for hydrocarbons. It has been observed that this threshold is when the outlet two-phase density (volume basis) is below 1.0 lb/cu-ft. Nearly all thermosyphons have outlet densities above When designing vertical thermosyphon reboilers with boiling at low operating process fluid pressures, check for the presence of a liquid preheat zone. Back pressure raises the boiling point at the interface of liquid preheat zone and subcooled boiling. This boiling point rise creates a liquid zone with relatively low heat transfer and it reduces the temperature driving force (MTD). If the operating pressure is below approximately 25 PSIA, there should be a liquid preheat zone. The lower the operating pressure, the more likely there is liquid preheat. If there is no liquid In the design of vertical thermosyphons, the recirculation rate should be set by the process engineer if there will be anything unusual about the connecting piping. The recirculation rate is especially sensitive to the size and configuration of the outlet piping. If the recirculation rate is left for the thermal designer to set, they will have to make piping The crital heat flux depends on the geometry of the bundle. The following estimate is based on 3/4 inch tubes on 15/16 inch pitch. It is actually good for any tube diameter with a tube pitch/tube diameter ratio of 1.25 and triangular tube pitch. For kettle reboilers use segmental baffles instead of full supports if shell fouling factor is greater Than 0.002(hr-ft 2 -F/Btu) A rule of thumb is that the pressure drop at the outlet nozzle should not be greater than 30% of the total static head. There is another tip in this boiling section about choking the flow with a small outlet nozzle. The inside flow area of the outlet nozzle should be the same or greater than the total flow area inside the tubing. For a channel with a side outlet the pressure drop is composed of a turning loss and a contraction loss The following equations calculate the pressure drop at the outlet. The pressure drop for expansion into the channel is not included here but is with the tube pressure drop. (used for pressure drop calc) (If Ktr less than 0.40, use 0.40) At low operating pressures there will be a sensible heat liquid zone with relatively low heat transfer. This is caused by the fact that a small pressure change will cause a large increase in the boiling point. There has been a case where 90% of the tube length was in the sub-cooled phase. What can you change that will decrease the size of the liquid preheat zone and One answer is to evaluate the piping system above the top tubesheet. In order to make an evaluation check the pressure drop at the outlet. There is on this section of the website equations to calculate the pressure drop of a nozzle that is at right angle to the top channel. Most vertical thermosyphons have the outlet nozzle at right angles to the top channel. There may be a simple change of enlarging the outlet nozzle that would be the cure. But there needs to be a check to make sure the nozzle and connecting piping are not so large that there is liquid slip. If enlarging the right angle nozzle and piping is not the answer then there are other configerations that will use less outlet pressure drop. Next the pressure drop of using a B type channel with I agree. I have also found that locating the inlet liquid nozzle directly under the vapor outlet is not good. In Amine BKU reboilers I found that locating the inlet rich amine liquid as close to the U-tube bundle tubesheet gave the best, consistant results in obtaining good solution stripping. This gives the heating medium Note: Input data into YELLOW cells and receive output in BOLD RED What Coil Diameter to Use to Start Design October, 2002 When starting to design a coil or other single continuous tube heat exchanger, the diameter is unknown. An example of this is an economizer in a heat recovery system. In this case it is desirable to have a single flow path rather than using parallel paths where headers are required. The following gives guidelines for liquids on a diameter selection: Size 1” tube 1 ¼” pipe 1 ½” pipe 2” pipe 3” pipe 4” pipe Estimate Gas Heat Transfer Rate for Hydrocarbons February, 1998 If you need to estimate a gas heat transfer rate or see if a program is getting a reasonable gas rate, use the following: h = 75 x (Op. pressure/100) 1/2 = 75 Btu/hr-ft 2 - o F Generally more accurate Or, h = 1.4W 0.8 = 66 Btu/hr-ft 2 - o F Generally understated Operating pressure = 100.0 Psia. W = 123.00 lb/tube/hr This is for inside the tubes. The rate will be lower for the shell side or if there is more than one exchanger. Estimate Hydrocarbon Heat Transfer Coefficient In Tubes Use the following equation to estimate the heat transfer coefficient when liquid is flowing inside 3/4 inch tubing: Hio = 87 Btu/Ft 2 -hr- o F Where: Viscosity = 3.0 cP. This is limited to a maximum viscosity of 3 cP Estimate - Latent Heat of Hydrocarbons An equation from the Bureau of Standards Miscellaneous Publication No. 97 can be used when the Specific Gravity is greater than 0.67 and less than 0.934. It is: (111 - 0.09T)/SG60 = 113 Btu/lb Lat heat = Unit Capacity flow rate 3,000-5,000 # / tube / hr 5,000-10,000 # / tube / hr 10,000-17,000 # / pipe / hr 150 / sqrt(avg. viscosity) = 17,000-35,000 # / pipe / hr 35,000-70,000 # / pipe / hr 70,000-130,000 # / pipe / hr Where: Lat heat = The fluid's Latent Heat in Btu/lb T = The fluid temperature in o F = 100 SG60 = The fluid's Specific gravity @ 60 o F = 0.9000 (0.67<SG<0.934) For hydrocarbons below a Specific Gravity of 0.67 and pressures below 50 psia, use: Lat heat = 172 - 0.195 T Liquid Thermal Conductivity for Light Hydrocarbons July, 1999 You can make an estimate for the liquid thermal conductivity of light hydrocarbons if you know their specific heat. It is good for propane and heavier. K = 0.025 / (specific heat) 1.5 = 0.0427 0.7000 But/lb °F Estimate Overall Heat Transfer Rate (U) in S & T December, 2001 In the preliminary design of shell and tube heat exchangers, you need an estimate of the overall heat transfer coefficient (U). Process simulator programs give you a UA from which you can estimate the surface if you have a U value. An estimate for a hydrocarbon U value can be made from the following: Rt = Fouling + Sqrt(avg. tube viscosity)/150 +((avg. shell viscosity) 0.27 )/140 = Where, Avg. tube viscosity = 2.0 cP Fouling = U = 1/Rt = Avg. shell viscosity = 3.0 cP Fouling is the total for both sides. The above is limited to a maximum viscosity of 3 cP for the tube side. There is no limit on the shell viscosity. This is also limited to bare tube surface with no internal turbulation devices. Estimated Tube Length That Lowers Tube Pressure Drop September, 2001 When the calculated tube side pressure drop exceeds the allowable, there are several design options. One option is to design with shorter tubing when the number of tube passes is one. To estimate the new tube length, use the following equation: New Lg = 15.9 feet Where Lg = Existing tube length = 20 feet Allow.Ap = Allowable Tube pressure drop = 1.00 psi Calc. Ap = Calculated Tube pressure drop = 2.00 psi The final tube length needs to be slightly longer than calculated because the calculated surface will be larger due to a lower tube velocity that gives a lower heat transfer. How to Calculate Excess Surface and Over-design Surface Specific Heat = Lg (Allowed Ap/Calc. Ap) 1/3 = 0.0005 100 x (A actual –A calculated ) / A calculated = 31.58% Where A actual = actual heat transfer surface = 250.0 A calculated = surface calculated from design overall heat transfer coefficient = 190.0 To calculate over-design surface use the clean overall heat transfer coefficient for A calculated . Use Superficial velocities to Calculate Best Heat Transfer Flow Pattern The best heat transfer occurs when there is an annular flow pattern. Then there is a relatively thin liquid film and little vapor in contact with the heat transfer surface. How do you tell if the flow is annular? It will be when the superficial gas velocity is above the following value: If the superficial liquid velocity is below 0.30 ft/s: Vg Max = 41.3 ft/sec. where VL = 0.25 ft/sec. (less than 0.30 ft/s) If the superficial liquid velocity is above 0.30 ft/s: Vg Max = 57.2 ft/sec. where VL = 1.00 ft/sec. (more than 0.30 ft/s) L/D Equation For Heat Transfer Coefficient Inside Tubing For Reynolds numbers below 10,000 there is an L/D effect on the heat transfer coefficient inside tubing. If you use the full tube length for L, you may be too conservative. There will be turbulation at the tube entrance before laminar flow is fully developed. The turbulent length needs to be subtracted from the full tube length. Use the following for tube sizes 1.0 inch or less. L = Tube Length - 0.0027 D i Re 11 feet Where L = variable to use in L/D expression, ft Tube Length = length of tube, ft = 20 feet D i = tube I.D., in = 0.650 inches Re = Reynolds number = 5,000 LMTD Correction Factor Charts for TEMA G and J Shell Types There are LMTD correction factor charts in TEMA for a single type G shell and two in series of type J shells. For charts of more shells in series, refer to the enclosed Dale Gulley-generated charts in this Workbook. Excess surface = 72 – 148 VL +100 VL 2 = 28.1 + 28 VL + 1.12 VL 2 = the superficial liquid velocity = the superficial liquid velocity = Divided Flow LMTD Correction November, 1996 Something to watch out for is the LMTD correction for Divided Flow Shell & Tube Exchangers. Divided flow (shell type J) does not have the same correction as the usual flow pattern (shell type E). We have seen several instances lately where a thermal design program made this correction factor mistake. True, there is very little difference at correction factors above 0.90. However, there is a difference at lower values. For example: Correction F n 0.805 0.775 0.765 0.65 Contact us if you do not have LMTD correction factor charts for divided flow. TEMA has one chart for a single shell but it gives high values for the above examples and it is hard to read in this range. Refer to the enclosed Dale Gulley charts for up to 4 shells in series that are found in this Workbook. Lowest Limit of LMTD Correction Factor What is the lowest LMTD correction factor to be used? Here is what several literature sources say: Heat Exchanger Design Handbook (HEDH): “F should be kept above 0.75 to 0.80” Perry's Chemical Engineers' Handbook: “Values of F less than 0.80 (0.75 at the very lowest) are generally unacceptable because the exchanger configuration chosen is inefficient ...” In over 50 years of experience, a correction factor of 0.75 is the lowest we have seen a thermal designer use. Although there was one case where an operating shell-and-tube heat exchanger reflected a lower LMTD correction factor than 0.75. Another way of looking at the correction factor is to never use a temperature cross of more than 5 degrees F in a single multi-tube pass shell. Minimum Flow Area For Shell Side Inlet Nozzle For single phase liquids and no impingement plate: (Flow (lb/hr) x 0.04) / (38.73 Sq.Root(ρ)) = 0.261 in 2 Shell Side Flow Rate = 2,000 lb/hr Shell Side fluid density = 62.50 lb/ft 3 For boiling liquids and no impingement plate: (Flow (lb/hr) x 0.04) / (22.36 Sq.Root(ρ)) = 0.453 in 2 Shell-type Flow Minimum area = Minimum area = Shell type "J" Equal outlet temperatures Shell type "E" Shell type "J" Cold outlet 5 o F higher than hot outlet Shell type "E" How to Calculate the Performance of Heat Exchangers with Plugged Tubes 1. Using the actual overall heat transfer coefficient (U), calculate the heat transfer resistances that excludes the tube side resistance: R other = 1/U -1/h io 2. Calculate new h io and new surface using usable number of tubes 3. Calculate new U U new = 1/(1/h io + R other ) How to Calculate the Performance of Heat Exchangers With Plugged Tubes 4. Calculate a new heat load from new surface and a new U How to Increase Heat Transfer for Low Reynolds Numbers September, 1999 If pressure drop is available and if the tube side Reynolds number is less than 5,000 and more than 1,000, you can probably increase the heat transfer considerably by increasing the number of tube passes and using shorter tubes. This will not only increase the tube velocity but there will be a lower L/D correction. Both of these factors will increase the heat transfer. Calculate When to Use Seal Bars on a Bundle to Increase Heat Transfer One of the fluid by-pass streams that lower the shell-side heat transfer is the stream that flows around the bundle. To evaluate, calculate a heat transfer variable named FSBP. It is the ratio of the by-pass to the cross flow area. The by-pass area is normally: FSBP = Bs (Ds - OTL) = 0.054 Bs(Ds-OTL)+Bs(OTL-Do)(P-Do)/P Where Ds = inside diameter of shell = 23.00 in. OTL = The Outer Tube Limit, or outer diameter of the tube bundle = 22.75 in. 18.00 in. Do = tube OD = 1.00 in. P = tube spacing = 1.25 in. Typical value is 1.25 x tube OD If FSBP is more that 0.15, then seal bars are needed. Calculate Tube Bundle Diameter January, 2000 Following are equations for one tube pass bundle diameter when the tube count is known or desired: For tubes with 30 Deg. Pitch: DS = 1.052 x pitch x SQRT(count) + tube O.D. = 17.384 inches Bs = Baffle Spacing = For tubes with 90 Deg. Pitch: DS = 1.13 x pitch x SQRT(count) + tube O.D. = 18.617 inches Where: Tube OD = 0.750 inches DS = Bundle diameter Count = 250 Number of tubes Pitch = Tube spacing = 1.000 inches Tube Count Calculation for S & T August, 2002 If you don't have a tube count table for a shell and tube exchanger, the tube count can be calculated. The following equation is good for any size tube on any tube pitch. It is primarily for situations where there is not a need for allowance for bundle entrance and exit area. Count = F [0.7854 x TC 2 - (PLw + Do - P) (TC x Npl)] / P 2 = 372 Tubes on Square Pitch Where: = 428 Tubes on Triangular Pitch Do = Tube O.D. = 0.750 inches F = 1.00 for square pitch F = 1.15 for triangle pitch Npl = Number of tube pass lanes (1 for two pass) = 2 PLw = Tube pass lane width (typical is 0.625 inches) = 0.625 inches P = Tube pitch = 1.000 inches TC = (Bundle diameter - tube O.D.) = 22.250 inches For tube pass lane width for square rotated tube pitch use (1.414P – Do). The decrease in the number of tubes due to bundle entrance and exit area could be allowed for by using a larger PLw. Tube Wall Temperature for Cooling Water January, 1999 When designing heat exchangers where hot process streams are cooled with cooling water, check the tube wall temperature. Hewitt says that where calcium carbonate may deposit, heat transfer surface temperatures above 140 o F should be avoided. Corrosion effects should also be considered at hot tube wall temperatures. As a rough rule of thumb, make this check if the inlet process temperature is above 200 o F for light hydrocarbon liquids and 300 - 400 o F for heavy hydrocarbons. Consider using Air coolers to bring the process fluid temperature down before it enters the water-cooled exchanger. Sometimes Larger Tubes are Better October, 1998 There is an exception to the rule that a shell and tube heat exchanger service using 3/4 inch tubes will be cheaper than one using 1-inch tubes. This is when the tubeside has a much lower heat transfer coefficient than the outside of the tubes and the following conditions are present: The flow will be in laminar flow if two (2) tube passes are used. If four (4) tube passes are used, the tubes in the 3/4 inch selection will have to be significantly shorter than allowed in order to meet pressure drop. On the other hand, the 1-inch tube design uses the full allowable tube length. Weighted MTD If there is more than a slight curvature in the heat release curve, things get more complicated. Then a step-wise method using local temperatures and local heat transfer coefficients are used to calculate the heat exchanger area. The question is what do you report as the MTD and the correction factor? There is a reference in TEMA in the temperature relations section T-3.2 that refers to a weighted MTD. The article mentioned was published by Dale Gulley in the June 1966 issue of Hydrocarbon Processing. The article shows how to calculate a weighted MTD and its correction factor if one is required. Estimate - Optimum Flow Velocity for Gas Inside Tubes Since the design of heat exchangers is a trial and error solution, a good starting point is desired. Usually the design starts with an estimated overall heat transfer coefficient. If you don't know a good starting value for this coefficient the equations presented here give this starting point with simple equations. In the design of heat exchangers using up the maximum allowable pressure drops gives the highest heat transfer for single phase fluids. The equations below estimates the tube velocity(W)for a gas that will meet the maximum allowable pressure drop. From W you can calculate the tube count or heat transfer coefficient. For a given tube length the following equation gives the optimum tube velocity for turbulent flow. Gases will be in turbulent flow more than 99% of the time. If your calculated tubeside velocity is below what the following equation calculates, you need more tube travel where tube travel is in the form of number of tube passes or total tube length(s) for countercurrent flow. These equations can be used for two phase flow as long as the two phase viscosity is less than 0.015 cp, For 3/4 inch tubes with 0.06 tube wall W = 1600(ΔPρ/L) 0.555 For 1.0 inch tubes with 0.06 tube wall W = 3500(ΔPρ/L) 0.555 Where: L = total tube lengths in ft. (Add [8 x tube ID in inches] ft for turning losses for each tube pass) W = lb/hr/tube ΔP = allowable pressure drop inside tubes in psi (deduct 15% for nozzle pressure drops) ρ = density in lb/cu.ft. L = 21 ft For 3/4" tubes, W = 1,497 Tube count = ΔP = 7 psig For 1" tubes, W = 3,274 Tube count = ρ = 2.66 lb/ft 3 Mass flow = 195000 lb/h Example Use 3/4 inch tubes and 16 foot tubes. The maximum allowable pressure drop inside the tubes is 7 psi (after nozzle deduction) and the gas density is 2.66 lb/cu.ft. The tube side flow is 195,000 lb/hr. What should be the starting tube count? Solution W = 1600(7 x 2.66/(16+5)) 0.555 W = 1497 lb/hr/tube Tube count = 195,000/1497 = 130 For a shell-and-tube heat exchanger, calculate the shell diameter when given the tube count here: Calculate S & T diameter from number of tubes Estimate - Hydrocarbon Gas Heat Transfer Coefficient in Shell Side Its difficult to estimate a gas heat transfer coefficient in the shell side because of the many variables. The following will give you a value within 25%. Ho = 430.Cp(ΔP/L x ρ)1/3 = 17 Btu/h ft2 °F where Cp = specific heat (Btu/lb-F) = 0.15 L = tube length (ft) = 10 ΔP = shell side pressure drop (Psi) = 2 (subtract nozzle losses) ρ = density of gas (lb/ft3) = 0.085 When starting to design a coil or other single continuous tube heat exchanger, the diameter is unknown. An example of this is an economizer in a heat recovery system. In this case it is desirable to have a single flow path rather than using parallel paths where headers are required. The following gives guidelines for liquids on a diameter selection: If you need to estimate a gas heat transfer rate or see if a program is getting a reasonable gas rate, use the following: Generally more accurate Use the following equation to estimate the heat transfer coefficient when liquid is flowing inside 3/4 inch tubing: An equation from the Bureau of Standards Miscellaneous Publication No. 97 can be used when the Specific Gravity You can make an estimate for the liquid thermal conductivity of light hydrocarbons if you know their specific heat. In the preliminary design of shell and tube heat exchangers, you need an estimate of the overall heat transfer coefficient (U). Process simulator programs give you a UA from which you can estimate the surface if you have a U value. 0.020 51 Fouling is the total for both sides. The above is limited to a maximum viscosity of 3 cP for the tube side. There is no When the calculated tube side pressure drop exceeds the allowable, there are several design options. One option is to design with shorter tubing when the number of tube passes is one. To estimate the new tube length, use the The final tube length needs to be slightly longer than calculated because the calculated surface will be larger due to The best heat transfer occurs when there is an annular flow pattern. Then there is a relatively thin liquid film and little vapor in contact with the heat transfer surface. How do you tell if the flow is annular? It will be when the For Reynolds numbers below 10,000 there is an L/D effect on the heat transfer coefficient inside tubing. If you use the full tube length for L, you may be too conservative. There will be turbulation at the tube entrance before laminar flow is fully developed. The turbulent length needs to be subtracted from the full tube length. Use the following for tube There are LMTD correction factor charts in TEMA for a single type G shell and two in series of type J shells. Something to watch out for is the LMTD correction for Divided Flow Shell & Tube Exchangers. Divided flow (shell type J) does not have the same correction as the usual flow pattern (shell type E). We have seen several instances lately where a thermal design program made this correction factor mistake. True, there is very little Contact us if you do not have LMTD correction factor charts for divided flow. TEMA has one chart for a single shell but it gives high values for the above examples and it is hard to read in this range. Refer to the enclosed Dale In over 50 years of experience, a correction factor of 0.75 is the lowest we have seen a thermal designer use. Although there was one case where an operating shell-and-tube heat exchanger reflected a lower LMTD correction factor than 0.75. Another way of looking at the correction factor is to never use a temperature cross of more than Calculate the expected performance of an exchanger that has had to have some tubes plugged by doing the following: 1. Using the actual overall heat transfer coefficient (U), calculate the heat transfer resistances that excludes the 1. You know the original overall heat transfer coefficient for the un-plugged exchanger and the number of tubes plugged. 2. Therefore, you know the original heat transfer area, the original hi and ho, the original tubeside velocity and the original duty and terminal temperatures. 3. You want to know what will be the new duty capacity and terminal temperatures with the unit operating with plugged tubes. How to Calculate the Performance of Heat Exchangers With Plugged Tubes After a heat exchanger goes into operation it may develope leaks in the tube walls. The following procedure calculates the new heat load and new overall heat transfer coefficient. 1. Using the actual overall heat transfer coefficient (U). calculate the heat transfer resistances that exclude the tubeside resistance If pressure drop is available and if the tube side Reynolds number is less than 5,000 and more than 1,000, you R other = 1/U -1/h io can probably increase the heat transfer considerably by increasing the number of tube passes and using shorter tubes. 2. Calculate new h io and new surface using usable number of tubes This will not only increase the tube velocity but there will be a lower L/D correction. Both of these factors will 3. Calculate new U U new = 1/(1/h io + R other ) 4. Calculate new heat load from new surface and new U One of the fluid by-pass streams that lower the shell-side heat transfer is the stream that flows around the bundle. To evaluate, calculate a heat transfer variable named FSBP. It is the ratio of the by-pass to the cross flow area. If you don't have a tube count table for a shell and tube exchanger, the tube count can be calculated. The following equation is good for any size tube on any tube pitch. It is primarily for situations where there is not a need for Tubes on Square Pitch Tubes on Triangular Pitch For tube pass lane width for square rotated tube pitch use (1.414P – Do). The decrease in the number of tubes When designing heat exchangers where hot process streams are cooled with cooling water, check the tube wall temperature. Hewitt says that where calcium carbonate may deposit, heat transfer surface temperatures above 140 o F should be avoided. Corrosion effects should also be considered at hot tube wall temperatures. As a rough rule of thumb, make this check if the inlet process temperature is above 200 o F for light hydrocarbon liquids and 300 - 400 o F for heavy hydrocarbons. Consider using Air coolers to bring the process fluid temperature down There is an exception to the rule that a shell and tube heat exchanger service using 3/4 inch tubes will be cheaper than one using 1-inch tubes. This is when the tubeside has a much lower heat transfer coefficient than the outside If four (4) tube passes are used, the tubes in the 3/4 inch selection will have to be significantly shorter than allowed in order to meet pressure drop. On the other hand, the 1-inch tube design uses the full allowable tube length. If there is more than a slight curvature in the heat release curve, things get more complicated. Then a step-wise method using local temperatures and local heat transfer coefficients are used to calculate the heat exchanger area. The question is what do you report as the MTD and the correction factor? There is a reference in TEMA in the temperature relations section T-3.2 that refers to a weighted MTD. The article mentioned was published by Dale Gulley in the June 1966 issue of Hydrocarbon Processing. The article shows how to calculate a weighted Usually the design starts with an estimated overall heat transfer coefficient. If you don't know a good starting value for this In the design of heat exchangers using up the maximum allowable pressure drops gives the highest heat transfer for single phase fluids. The equations below estimates the tube velocity(W)for a gas that will meet the maximum allowable pressure drop. From W you can calculate the tube count or heat transfer coefficient. For a given tube length the following equation gives the optimum tube velocity for turbulent flow. Gases will be in turbulent flow more than 99% of the time. If your calculated tubeside velocity is below what the following equation calculates, you need more tube travel where tube travel is in the form of number of tube passes or total tube length(s) for countercurrent flow. These equations can be used for two phase flow as long as the two phase viscosity is less than 0.015 cp, 130 60 Use 3/4 inch tubes and 16 foot tubes. The maximum allowable pressure drop inside the tubes is 7 psi (after nozzle deduction) and the Its difficult to estimate a gas heat transfer coefficient in the shell side because of the many variables. The following will give you a value within 25%. Calculate the expected performance of an exchanger that has had to have some tubes plugged by doing the following: 1. You know the original overall heat transfer coefficient for the un-plugged exchanger and the 2. Therefore, you know the original heat transfer area, the original hi and ho, the original tubeside velocity 3. You want to know what will be the new duty capacity and terminal temperatures with the After a heat exchanger goes into operation it may develope leaks in the tube walls. The following procedure calculates the new heat load and new overall heat transfer coefficient. 1. Using the actual overall heat transfer coefficient (U). calculate the heat transfer resistances that exclude the tubeside resistance Avoid Small Baffle Cuts in S & T Condensers July, 2001 There will be a theoretical liquid level when there is condensation in a heat exchanger. The condensing heat transfer coefficient decreases as its' liquid film increases. For best heat transfer the liquid level should be low as possible. Small baffle cuts in a shell and tube exchanger will hold a higher liquid level than large cuts. Use a separated flow model equation system to determine the theoretical liquid level. Unless you want subcooling, do not use a baffle cut that would hold a liquid level higher than the theoretical one. Estimate - Condensing Heat Transfer Coefficient for Hydrocarbons Inside Tubing Cond h = 828 Where Cond h = W = Condensing fluid in tubes = 750.00 lbs/hr/tube Maximum Condensing Rate Inside Tubes August, 2001 Following is a close estimate of the maximum heat transfer rate for total condensation. It is based on the maximum condensing rate for the average hydrocarbon to be 750 BTU/hr-ft 2 -F. It is good for other types of chemical compounds. Hi = 3,193 Btu/hr Where, K liq = liquid thermal conductivity of the condensate =0.350 Btu/hr-ft- o F For example this equation yields a maximum heat transfer rate for steam to be 3,600. Quick Estimate for Reflux Condenser LMTD in Air-cooler This type of service has steam condensing out from a non-condensable gas which is mostly CO 2 . The condensing curve has a hump which will give a LMTD higher than one calculated from a straight line condensing plot. An equation that makes a quick estimate for the LMTD is: Standard LMTD x Factor In the case of outlet process temperatures below 153.5 o F, Then LMTD Factor = 1.4 -0.0092 (T -110) Where T = outlet temperature and air inlet temperature is 100 o F. Reflux (“Knock back”) Condenser June, 2001 Do not design this like the usual vertically condensing heat exchanger where both gas and liquid flow in one direction. In this type of condenser, the coldest condensate will be in contact with the entering hot vapor (in the bottom section). Nearly everything about this type condenser is different. It is both difficult to design and difficult to control. The flow patterns, pressure drop and heat transfer calculations are different. Be sure the heat transfer calculations are zoned. Btu/(hr)(ft 2 )( o F) (4.15) W 0.8 = Inside condensing heat transfer coefficient 750 (K liq / 0.07) 0.9 = Types of Steam Condensers Small steam condensers use shell-and-tube heat exchangers while large steam condensers use surface condensers. A conventional “E” type shell is used when the steam condensing temperature is above approximately 120 o F. For lower temperatures, a “X” type shell can be used. A point is reached where the size or operating pressure requires a surface condenser. Sulfur Condenser - Tube Velocity Limits For good operation of a sulfur condenser the design velocities inside the tubes should be within certain limits. The velocity range is between 1.5 and 6.0 lb/sq ft-sec. Below this range there will be slugging. Above this range sulfur fogging will occur.. Small Temperature Pinch Points in Condensers November, 1998 Be extra careful when condensers are designed with a small pinch point. A pinch point is the smallest temperature difference on a temperature vs. heat content plot that shows both streams. If the actual pressure is less than the process design operating pressure, there can be a significant loss of heat transfer. This is especially true of fluids that have a relative flat vapor pressure plot like ammonia or propane. For example: If an ammonia condenser is designed for 247 PSIA operating pressure and the actual pressure is 5 PSI less and the pinch point is 8 o F, there can be a 16% drop in heat transfer. When to Slope Single Tube Pass Tubes in Condensing Service January, 2002 At low vapor velocities, it has been proven that even a slight downward slope of tubes gives a significant increase in heat transfer in the case of tube-side condensation. But this does not mean the larger the slope the higher the heat transfer. The benefit of sloping stops at an angle of approximately 10 o . A common case of a condenser needing to have the tubes sloped is when they are operating near atmospheric pressure and there is one tube pass. An example of this is a sulfur condenser. It has a low pressure drop usually less than 0.5 psi. They typically are designed with a slope of 1/8 inch per foot of tubing. Zone Those Condensers The heat transfer and pressure drop of a condenser usually should be zoned. A typical heat exchanger that condenses 100% of the vapor will go through 2 or 3 different flow pattern zones before the flow becomes a liquid. There is better accuracy if the flow patterns are determined and their individualistic equations are used. There will be a theoretical liquid level when there is condensation in a heat exchanger. The condensing heat transfer coefficient decreases as its' liquid film increases. For best heat transfer the liquid level should be low as possible. Small baffle cuts in a shell and tube exchanger will hold a higher liquid level than large cuts. Use a separated flow model equation system to determine the theoretical liquid level. Unless you want subcooling, do not use a baffle cut Estimate - Condensing Heat Transfer Coefficient for Hydrocarbons Inside Tubing Following is a close estimate of the maximum heat transfer rate for total condensation. It is based on the maximum condensing rate for the average hydrocarbon to be 750 BTU/hr-ft 2 -F. It is good for other types of chemical compounds. This type of service has steam condensing out from a non-condensable gas which is mostly CO 2 . The condensing curve has a hump which will give a LMTD higher than one calculated from a straight line condensing plot. An Do not design this like the usual vertically condensing heat exchanger where both gas and liquid flow in one direction. In this type of condenser, the coldest condensate will be in contact with the entering hot vapor (in the bottom section). Nearly everything about this type condenser is different. It is both difficult to design and difficult to control. The flow patterns, pressure drop and heat transfer calculations are different. Be sure the heat transfer calculations are zoned. Small steam condensers use shell-and-tube heat exchangers while large steam condensers use surface condensers. A conventional “E” type shell is used when the steam condensing temperature is above approximately 120 o F. For lower temperatures, a “X” type shell can be used. A point is reached where the size or operating pressure requires For good operation of a sulfur condenser the design velocities inside the tubes should be within certain limits. The velocity range is between 1.5 and 6.0 lb/sq ft-sec. Below this range there will be slugging. Above this range sulfur Be extra careful when condensers are designed with a small pinch point. A pinch point is the smallest temperature difference on a temperature vs. heat content plot that shows both streams. If the actual pressure is less than the process design operating pressure, there can be a significant loss of heat transfer. This is especially true of fluids that have a relative flat vapor pressure plot like ammonia or propane. For example: If an ammonia condenser is designed for 247 PSIA operating pressure and the actual pressure is 5 PSI less and the pinch point is 8 o F, there January, 2002 At low vapor velocities, it has been proven that even a slight downward slope of tubes gives a significant increase in heat transfer in the case of tube-side condensation. But this does not mean the larger the slope the higher the heat transfer. The benefit of sloping stops at an angle of approximately 10 o . A common case of a condenser needing to have the tubes sloped is when they are operating near atmospheric pressure and there is one tube pass. An example of this is a sulfur condenser. It has a low pressure drop usually less than 0.5 psi. They typically are designed with The heat transfer and pressure drop of a condenser usually should be zoned. A typical heat exchanger that condenses 100% of the vapor will go through 2 or 3 different flow pattern zones before the flow becomes a liquid. There is Rotated Square Tube Pitch Some heat exchanger specifications for shell and tube heat exchangers mention square pitch but do not specifically mention rotated square pitch. Engineers with little thermal design experience who are trying to strictly adhere to the specifications may reject this type of tube pitch. The benefits for this type of tube pitch sometimes get lost because of this. Rotated square pitch gives better mixing of the shell fluid and better heat transfer for the heavier fluids. Frequently the shell size can be reduced when there will be heavier liquids on the shell side and the designer uses rotated square pitch. Caution When Using a Longitudinal Baffle in the Shell Side The following are potential problems when considering the use of a longitudinal baffle in a new S & T heat exchanger: 1. The largest temperature drop across the long baffle is more than 250 o F. Then the thermal efficiency is lost due to conduction across the long baffle. Check and make sure this has been taken into consideration; 2. If the long baffle is not welded to the shell, the pressure drop across the long baffle is more than 7 to 8 psi. This will also lose thermal efficiency. The seal on the long baffle should be tested in the shop after fabrication. Using Turbulators for Tube Side Laminar Flow If the flow inside the tubes of a heat exchanger is in laminar or viscous flow, take a look at enhancing the heat transfer. One simple and inexpensive device is the twisted-tape insert. Using twisted-tape inserts for laminar flow in new heat exchangers results in cost savings and smaller heat exchangers. Twisted-tape inserts can be used in existing heat exchangers to make a significant increase in capacity. The amount of increase in heat exchanged depends on whether the increase in pressure drop can be tolerated. If there is no pressure drop limitation, there can be as much as a 50% increase in capacity. Here are the recommended guide lines for using twisted tape inserts: 1 Pressure drop in the tube side without inserts is less than 3 to 4 PSI. 2 Minimum fluid viscosity of 2 centipoise unless there is a very low velocity 3 Use a minimum tube diameter of 5/8” for .001 fouling. Use a minimum of 1” diameter for 0.0015 fouling. It is not recommendable to use turbulators in a service that has a fouling factor greater than 0.0015. These guidelines for tube diameter are due to fouling being more of a problem with turbulators in small tubes. Triple Segmental Baffles November, 1997 There is more than one kind of triple segmental baffles in the shell side of heat exchangers. Be sure you know which kind if you are checking a design that uses them. There is the kind you see in TEMA where there are three different groups in a set. The total number of baffle pieces is six. There is the kind that is like producing two double segmental streams in parallel. There are two groups in a set and a total of five baffle pieces. Another kind has only three pieces in a group and each piece has a different shape. Entrance and Exit Space for Shell Nozzles January, 2001 There have been cases where not enough space was under the shell nozzles. This can be critical for applications like a horizontal thermosyphon or other pressure drop sensitive applications. Check the distance from the nozzle I.D. to the nearest tube row or impingement plate. If there is an impingement plate this distance should be ¼ or more of the nozzle I.D. If there is no impingement plate this distance should be 1/6 or more of the nozzle I.D. If pressure drop is not a consideration and TEMA requirements are met and vibration is not a problem then the above calculated distance could be reduced. This criterion naturally doesn't apply to shells with distributor belts or where the nozzle is beyond the back of a U-tube bundle. For information on calculating shell nozzle pressure drops, refer to “Calculate Shell Nozzle Pressure Drop” in the calculation Tab of this Workbook. Horizontal vs. Vertical Baffle Cut in S & T Exchangers May, 2001 In shell and tube heat exchangers it is safer from a thermal design standpoint to use vertical baffle cuts but horizontal cuts have an advantage in certain situations. Horizontal cuts are best if the shell side stream is clean and single phase. There will be less of the shell side stream bypassing through the tube pass lanes. Since in a multi-tube pass exchanger there will be more horizontal tube pass lanes than vertical pass lanes, you need to flow perpendicular to these pass lanes for minimum by-passing of the shell stream. This means horizontal cut. Where you do not want to use horizontal cut is when there is either condensing or where there is the possibility of foreign material being in the flowing stream. It is suggested to use a maximum fouling factor of 0.002 for horizontal baffle cut. It may be possible to use horizontal cut in certain boiling applications. Is an Expansion Joint Required in the Shell? December, 1998 A fixed tube sheet exchanger does not have provision for expansion of the tubing when there is a difference in metal temperature between the shell and tubing. When this temperature difference reaches a certain point, an expansion joint in the shell is required to relieve the stress. It takes a much lower metal temperature difference when the tube metal temperature is hotter than the shell metal temperature to require an expansion joint. Typically, an all steel exchanger can take a maximum of approximately 40 o F metal temperature difference when the tube side is the hottest. When the shell side is the hottest, the maximum is typically 150 o F. Usually if an expansion joint is required, it is because the maximum allowable tube Compressive stress has been exceeded. According to the TEMA procedure for evaluating this stress, the compressive stress is a strong function of the unsupported tube span. This is normally twice the baffle spacing. Increasing Capacity of Existing Shell & Tube Exchangers March, 1997 To increase heat transfer check out using low fins or other special tubing. When an increase in capacity will cause excessive pressure drop, you may not have to junk the heat exchangers. Investigate the relatively inexpensive modification of reducing the number of tube passes. Other possibilities are arranging multiple exchangers in parallel. Locating Vents on the Shell Side of Vertical Exchangers July, 1998 Proper venting of equipment is not always given the consideration it deserves. One place where venting is especially a problem is underneath the tubesheet of a vertical exchanger. The problem is that there will always be a space above the vent connection to trap gases or vapors. Besides the poor heat transfer in this region, this can cause corrosion problems. It is important to get the vent connection as close to the tubesheet as possible. Using multiple connections that are smaller is one solution. Another solution is to fabricate the upper tubesheet with a small vent tunnel inside. Flange Gasket Location May, 1999 There is an optimum diameter of the gasket for flanges. It is when the total Operating moment of the flange under pressure is equal to the gasket seating moment. For low-pressure flanges, the diameter should be as close to the bolt circle as possible. For high-pressure flanges, the diameter should be as close to the flange I.D. as possible. In this case, low pressure is considered to be below 300 psi. High pressure is considered to be approximately 750 psi and higher. Using Rods for Tube Inserts to Increase Heat Transfer August, 1998 Use concrete reinforcing rods inserted inside the tubes to increase the heat transfer and tube velocity. It is a quick and economical solution. This is usually done only in clean services. A typical case is using 3/8" rods inside a 3/4" x 14 BWG tube. The tube side heat transfer coefficient is increased by a factor of 1.7. However, you have to be able to stand the increase in pressure drop. It goes up by a factor of 9.5. Another example is a 1.0" x 16 BWG avg. wall tube where the heat transfer goes up by a factor of 1.17 and the pressure drop by a factor of 3.5. Shell Side Impingement Protection There may be tube vibration or erosion if the shell-side fluid velocity is above a maximum value. These values can be found in TEMA section RCB-4.61 & 4.62. In the eighth edition the maximum values can be found on page 35. The most common impingement protection is a plate baffle that is slightly above the tube bundle. But this type of protection has some drawbacks. It has a relatively higher pressure drop than most other methods and the tubes on the first several rows tend to vibrate. Other types of impingement protection are: 1. Plate within a nozzle enlarger 2. Solid rods instead of tubes for the first 2 or 3 rows. 3. Snap-on tube protectors on top of the tubes in the first 2 or 3 rows 4. Small angle iron types setting on top of the tubes in the first 2 or 3 rows 5. Vapor belt Special S & T Exchanger Type (NTIW) September, 1998 A shell & tube heat exchanger with normal segmental baffles has tubes that miss every other baffle. This can lead to long unsupported tube lengths for some applications. A long tube span has a low natural frequency and is prone to vibration. One solution is to design a “no tubes in window” (NTIW) exchanger. This design has no tubes in the baffle cut out. By using intermediate supports between baffles, the natural frequency of the tubes can be raised considerably to resist vibration. When to Consider By-pass Strips in S & T Bundle Use a by-pass strip if tubes are removed under a nozzle. Removing tubes leaves an open area where the shell fluid can flow either over or under the bundle. Consider by-pass strips if the bundle to shell clearance is more than 3/4 inches and the shell fluid is mostly sensible heat transfer. Especially consider by-pass strips if the shell liquid is a hydrocarbon with an average viscosity greater than 1 centipoise and the tube fluid has a high heat transfer coefficient (example water). In this case, a 5 to 10% increase in heat duty can be achieved by installing by-pass strips. What is too Large a Temperature Change in 2 Tube Passes? December, 1996 Warning! Large tube side temperature change. A big difference between the inlet and outlet temperature of the tube side causes leakage and bypass problems. The worst case is a shell and tube exchanger with two (2) tube passes where a gasket is used to seal between the passes. A careful analysis should be made if the temperature difference across the pass plate is more than 300 oF. For a channel type that has a welded in pass plate, make an analysis if the temperature difference is more than 450 oF. If this temperature difference causes an over stressed condition, possible cures are: - Add a unit in series so each unit has a smaller temperature difference; - Use one tube pass if the penalty isn't too great; - For air coolers, use a split headers design. Rotated Square Tube Pitch February, 1999 Some heat exchanger specifications for shell and tube heat exchangers mention square pitch but do not specifically mention rotated square pitch. Engineers with little thermal design experience who are trying to strictly adhere to the specifications may reject this type of tube pitch. The benefits for this type of tube pitch sometimes get lost because of this. Rotated square pitch gives better mixing of the shell fluid and better heat transfer for the heavier fluids. Frequently the shell size can be reduced when there will be heavier liquids on the shell side and the designer uses rotated square pitch. Longitudinal Baffle Heat Conduction Cures With a longitudinal baffle and a long temperature range there can be a problem with heat conduction through the longitudinal baffle. There will be a loss of thermal efficiency due to the heat conduction. The longitudinal baffle can be fabricated in one of two ways. 1. Leaving an small enclosed air gap between two longitudinal baffles. 2. Spray an insulating material like Ryton on the longitudinal baffle. Design Temperatures of Carbon Steel and Low Alloy Tubes and Tubesheets Use the higher of the shell-side and tube-side design temperatures up to 650 F. At higher design temperatures use the arithmetic average of the 2 design temperatures. Design Temperatures of Nonferrous Tubes and Tubesheets Water in the shell-side Use the arithmetic average of the shell-side and tube-side design temperatures. Water in the tube-side Use the higher of the tube-side design temperature or tube-side outlet temperature + 1/3 of the LMTD. Some heat exchanger specifications for shell and tube heat exchangers mention square pitch but do not specifically mention rotated square pitch. Engineers with little thermal design experience who are trying to strictly adhere to the specifications may reject this type of tube pitch. The benefits for this type of tube pitch sometimes get lost because of this. Rotated square pitch gives better mixing of the shell fluid and better heat transfer for the heavier fluids. Frequently the shell size can be reduced when there will be heavier liquids on the shell side and the designer uses The following are potential problems when considering the use of a longitudinal baffle in a new S & T heat exchanger: The largest temperature drop across the long baffle is more than 250 o F. Then the thermal efficiency is lost due to conduction across the long baffle. Check and make sure this has been taken into consideration; If the long baffle is not welded to the shell, the pressure drop across the long baffle is more than 7 to 8 psi. This will also lose thermal efficiency. The seal on the long baffle should be tested in the shop after fabrication. If the flow inside the tubes of a heat exchanger is in laminar or viscous flow, take a look at enhancing the heat transfer. One simple and inexpensive device is the twisted-tape insert. Using twisted-tape inserts for laminar flow in new heat exchangers results in cost savings and smaller heat exchangers. Twisted-tape inserts can be used in existing heat exchangers to make a significant increase in capacity. The amount of increase in heat exchanged depends on whether the increase in pressure drop can be tolerated. If there is no pressure drop limitation, there can be as much as a 50% Use a minimum tube diameter of 5/8” for .001 fouling. Use a minimum of 1” diameter for These guidelines for tube diameter are due to fouling being more of a problem with turbulators in small tubes. There is more than one kind of triple segmental baffles in the shell side of heat exchangers. Be sure you know which kind if you are checking a design that uses them. There is the kind you see in TEMA where there are three different groups in a set. The total number of baffle pieces is six. There is the kind that is like producing two double segmental streams in parallel. There are two groups in a set and a total of five baffle pieces. Another kind has only three pieces There have been cases where not enough space was under the shell nozzles. This can be critical for applications like a horizontal thermosyphon or other pressure drop sensitive applications. Check the distance from the nozzle I.D. to the nearest tube row or impingement plate. If there is an impingement plate this distance should be ¼ or more of the nozzle I.D. If there is no impingement plate this distance should be 1/6 or more of the nozzle I.D. If pressure drop is not a consideration and TEMA requirements are met and vibration is not a problem then the above calculated distance could be reduced. This criterion naturally doesn't apply to shells with distributor belts or where the nozzle is beyond For information on calculating shell nozzle pressure drops, refer to “Calculate Shell Nozzle Pressure Drop” in In shell and tube heat exchangers it is safer from a thermal design standpoint to use vertical baffle cuts but horizontal cuts have an advantage in certain situations. Horizontal cuts are best if the shell side stream is clean and single phase. There will be less of the shell side stream bypassing through the tube pass lanes. Since in a multi-tube pass exchanger there will be more horizontal tube pass lanes than vertical pass lanes, you need to flow perpendicular to these pass lanes for minimum by-passing of the shell stream. This means horizontal cut. Where you do not want to use horizontal cut is when there is either condensing or where there is the possibility of foreign material being in the flowing stream. It is suggested to use a maximum fouling factor of 0.002 for horizontal baffle cut. It may be possible to use horizontal A fixed tube sheet exchanger does not have provision for expansion of the tubing when there is a difference in metal temperature between the shell and tubing. When this temperature difference reaches a certain point, an expansion joint in the shell is required to relieve the stress. It takes a much lower metal temperature difference when the tube metal temperature is hotter than the shell metal temperature to require an expansion joint. Typically, an all steel exchanger can take a maximum of approximately 40 o F metal temperature difference when the tube side is the hottest. When the shell side is the hottest, the maximum is typically 150 o F. Usually if an expansion joint is required, it is because the maximum allowable tube Compressive stress has been exceeded. According to the TEMA procedure for evaluating this stress, the compressive stress is a strong function of the unsupported tube span. This is normally To increase heat transfer check out using low fins or other special tubing. When an increase in capacity will cause excessive pressure drop, you may not have to junk the heat exchangers. Investigate the relatively inexpensive modification of reducing the number of tube passes. Other possibilities are arranging multiple exchangers in parallel. Proper venting of equipment is not always given the consideration it deserves. One place where venting is especially a problem is underneath the tubesheet of a vertical exchanger. The problem is that there will always be a space above the vent connection to trap gases or vapors. Besides the poor heat transfer in this region, this can cause corrosion problems. It is important to get the vent connection as close to the tubesheet as possible. Using multiple connections that are smaller is one solution. Another solution is to fabricate the upper tubesheet with a small vent tunnel inside. There is an optimum diameter of the gasket for flanges. It is when the total Operating moment of the flange under pressure is equal to the gasket seating moment. For low-pressure flanges, the diameter should be as close to the bolt circle as possible. For high-pressure flanges, the diameter should be as close to the flange I.D. as possible. In this case, low pressure is considered to be below 300 psi. High pressure is considered to be approximately 750 psi Use concrete reinforcing rods inserted inside the tubes to increase the heat transfer and tube velocity. It is a quick and economical solution. This is usually done only in clean services. A typical case is using 3/8" rods inside a 3/4" x 14 BWG tube. The tube side heat transfer coefficient is increased by a factor of 1.7. However, you have to be able to stand the increase in pressure drop. It goes up by a factor of 9.5. Another example is a 1.0" x 16 BWG avg. wall tube where the heat transfer goes up by a factor of 1.17 and the pressure drop by a factor of 3.5. There may be tube vibration or erosion if the shell-side fluid velocity is above a maximum value. These values can be found in TEMA section RCB-4.61 & 4.62. In the eighth edition the maximum values can be found on page 35. The most common impingement protection is a plate baffle that is slightly above the tube bundle. But this type of protection has some drawbacks. It has a relatively higher pressure drop than most other methods and the tubes on A shell & tube heat exchanger with normal segmental baffles has tubes that miss every other baffle. This can lead to long unsupported tube lengths for some applications. A long tube span has a low natural frequency and is prone to vibration. One solution is to design a “no tubes in window” (NTIW) exchanger. This design has no tubes in the baffle cut out. By using intermediate supports between baffles, the natural frequency of the tubes can be raised Use a by-pass strip if tubes are removed under a nozzle. Removing tubes leaves an open area where the shell fluid Consider by-pass strips if the bundle to shell clearance is more than 3/4 inches and the shell fluid is mostly sensible Especially consider by-pass strips if the shell liquid is a hydrocarbon with an average viscosity greater than 1 centipoise and the tube fluid has a high heat transfer coefficient (example water). In this case, a 5 to 10% increase Warning! Large tube side temperature change. A big difference between the inlet and outlet temperature of the tube side causes leakage and bypass problems. The worst case is a shell and tube exchanger with two (2) tube passes where a gasket is used to seal between the passes. A careful analysis should be made if the temperature difference across the pass plate is more than 300 oF. For a channel type that has a welded in pass plate, make an analysis if the temperature difference is more than 450 oF. If this temperature difference causes an over stressed Some heat exchanger specifications for shell and tube heat exchangers mention square pitch but do not specifically mention rotated square pitch. Engineers with little thermal design experience who are trying to strictly adhere to the specifications may reject this type of tube pitch. The benefits for this type of tube pitch sometimes get lost because of this. Rotated square pitch gives better mixing of the shell fluid and better heat transfer for the heavier fluids. Frequently the shell size can be reduced when there will be heavier liquids on the shell side and the designer uses With a longitudinal baffle and a long temperature range there can be a problem with heat conduction through the Use the higher of the tube-side design temperature or tube-side outlet temperature + 1/3 of the LMTD. Choosing Fin Spacing June, 2002 In waste heat applications, the fin spacing depends not only on the heat transfer but the cleanliness of the exhaust gas. If the gas is fouled from soot or other fine particulates, use a maximum of 5 fins per inch. For very dirty gases the fin spacing can be as low as 2 fins per inch. Usually there will be soot if fuels heavier than diesel fuel are fired. The designer needs to know the source of the waste heat gas so that he can make a decision on what fin spacing to use. HRSG Nozzle Size April, 2002 For an estimate of the nozzle size entering and leaving a HRSG unit use: D = 4.43 inches Where: D = diameter of nozzle Flow = Gas flow = 1,000 lbs/hr This is based on a total of 0.8 inches of water Face Area for HRSG Units April, 2001 The starting point in the design of a heat-recovery steam generator (HRSG) is the face area. This will determine the preliminary duct dimensions and starting face areas of any economizers and superheaters. 0.40 ft 2 Where: Flow = 1,000 lbs/hr Where face area is in square feet. This is based on using 2 inch O.D. tubing with 1 inch high fins. The tubing is arranged on 4 1/8 inch triangular pitch. Maximum Exhaust Gas Temperature for Steel Fin Tubes Here is an approximation of the maximum exhaust temperature for steel fin tubes when generating steam. Otherwise, the fins would need to be the more expensive 409 SS material. This is based on the typical 2 inch O.D. tubing with 1 inch fins and 6 to 7 fins/inch. MaxTg = 1,021 o F Where MaxTg = maximum gas temperature Btemp = water boiling temperature = 300 o F. When to Use Bare Tubes in Waste Heat Boilers Use bare tubes if the bundle is quite small or the gas temperature is greater than 1,350 to 1,400 o F. 1,090 - 0.23 Btemp = 0.14 x (flow) 1/2 = Face area = Exhaust Gas flow = (Flow / 2,500) = In waste heat applications, the fin spacing depends not only on the heat transfer but the cleanliness of the exhaust gas. If the gas is fouled from soot or other fine particulates, use a maximum of 5 fins per inch. For very dirty gases the fin spacing can be as low as 2 fins per inch. Usually there will be soot if fuels heavier than diesel fuel are fired. The designer needs to know the source of the waste heat gas so that he can make a decision on what fin spacing to use. The starting point in the design of a heat-recovery steam generator (HRSG) is the face area. This will determine This is based on using 2 inch O.D. tubing with 1 inch high fins. The tubing is arranged on 4 1/8 inch triangular pitch. Otherwise, the fins would need to be the more expensive 409 SS material. This is based on the typical 2 inch O.D. Cooling Water Flowing Inside 304SS U-tubes June, 1999 Normally it is OK to use 304SS when cooling water with low chloride content is flowing inside U-tubes. But if for some reason the operating pressure drops to saturation there can be corrosion problems. The tube vibration that results from the flashing of steam amplifies the stress that causes stress corrosion cracking. Normally it is OK to use 304SS when cooling water with low chloride content is flowing inside U-tubes. But if for some reason the operating pressure drops to saturation there can be corrosion problems. The tube vibration that Calculating Fouled Pressure Drop August, 1999 There are various ways to account for fouling when calculating pressure drop. One way would be to add a small amount to the tube diameter. This has a complex effect that is not linear in nature. A simpler method is to add 10% for each 0.001 increase in fouling factor. Then multiply this factor by the clean pressure drop. You would use a pressure drop factor of 1.2 for a fouling factor of 0.002. Allowable Pressure Drop Suggestions March, 2002 If you are at a loss as to what allowable pressure drop to specify, here are some suggestions: Gas Liquid Change of phase Boiling: Condensing operating pressure Allowable Shell Side Pressure Drop if a Multi-leaf (a.k.a. Lamaflex) Long Baffle is Used Four thin (0.008”) stainless strips are normally used to seal the sides of the long baffle. Because of their flexibility, they are not able to withstand large shell side pressure drops. It is best to limit the pressure drop to 5 psi with 7.5 psi being the maximum. Better Baffle Window Pressure Drop Equation A new baffle window pressure drop equation has been published in the June 2004 issue of Hydrocarbon Processing. The name of the article is “More Accurate Exchanger Shell-Side Pressure Drop Calculations”. The article can be found on this page with the subject “Heat Exchanger Articles Published by Dale Gulley”. The equation improves the accuracy of the shell side pressure drop. Refer to the article for more detail. The equation has the following form: K P = Pressure loss coefficient for velocity head equation fi = Friction factor for ideal tube bundle C 1 = Constant based on the type of tube layout For 30 deg. Triangular, 2.2; For 90 deg. Square, 3.64; For 45 deg. sq. rotated, 2.29; For 60 deg. Triangular, 1.79 estimated. N cw = Effective number of tube rows crossed in baffle window Less than atmospheric Atmospheric to 25 psi 50 to 150 psi 25 to 50 psi 0.5 1 2 3.0 to 5.0 150 psi + 3 1.0 to 5.0 psi for less than 10 % vapor Fluid and Condition Allowable Pressure Drop, psi 3 to 5 8 to 10 0.5 to 1.0 psi for greater than 10 % vapor D = Distortion factor for ideal fluid stream. It varies with baffle cut. Refer article elsewhere on this site for equation. (Baffle cuts from 24% to 29% (fractional) have a distortion factor of 1.0) Sl = Total of leakage areas (in 2 ) Sw = Net flow area in baffle window (in 2 ) EXAMPLE: This is taken from the first experimental case in “A Reappraisal of Shellside Flow in Heat Exchangers HTD-Vol. 36”. Average flow of 990,000 lb/hr with a density of 62.4 lb/ft3 is flowing through a 13.25 ID nozzle. The shell ID is 23.25 in. and the OTL is 22.375 in. The tube OD is 0.75 in. on a tube pitch of 0.9375 in. with 30 degree layout. There are 7 baffles and 26% baffle cut. The following are taken from a tip in this section named “Improve Shell Side Pressure Drop Calculations” fi = 0.1025 Ncw = 5.96 Sl = 11.0 Sw = 44.47 C1 for a 30 degree layout is 2.2 D = 1 since the fractional baffle cut is 26% Kp = 0.1025 ( (2.2 x5.96) -2(11/44.47) 2 ) ) Kp = 1.33 Gw = (990000 x 0.04)/44.47 = 890.5 (#/sqft-sec) ΔPw = Kp x 0.000108 x Gw 2 /ρ ΔPw = 1.33 x 0.000108 x (890.5) 2 x 7/62.4 ΔPw = 12.78 Designing Better Use of Tube Pressure Drop October, 1999 When the calculated pressure drop inside the tubes is under-utilized, the estimated pressure drop with increased number of tube passes is: AP x (NPASS/OPASS) 3 = 12.0 psi Where AP = Previous Pressure drop 1.5 psi NPASS = New number of tube passes = 4 OPASS = Old number of tube passes = 2 This would be a good estimate if advantage is not taken of the increase in heat transfer. Since the increased number of tube passes gives a higher velocity and increases the calculated heat transfer coefficient, the number of tubes to be used will decrease. The use of fewer tubes increases the new pressure drop. For a better estimate of the new pressure drop, add 25% if the heat transfer is all sensible heat. Effect of 1 st Tube Rows on Shell Nozzle Pressure Drop Usually when shell-and-tube heat exchangers are designed, the tube layout is made so that the shell entrance area is New tube AP = approximately equal to the shell nozzle flow area. The average distance to the 1st tube row is Dn/4 where Dn is the inside diameter of the shell nozzle. In this case the pressure loss coefficient is 1.0 for the pressure drop calculation for the shell nozzle entrance. If the shell nozzles are greater than 2” and some tubes are not omitted from the tube layout, the nozzle entrance pressure drop can be significantly higher than the normal calculation based on the nozzle flow area. In a case of a 6” shell nozzle and where no tubes were omitted in a BEM type heat exchanger, the pressure drop was 3 times higher than that calculated with just the nozzle flow area. For more information, you can refer to the tip “Calculate Shell Nozzle Pressure Drop” in this Workbook. Kettle Pressure Drop April, 1999 Usually you will see the allowable pressure drop on the specification sheet for the shell side of a kettle reboiler to be stated as “nil”. This is close to being true only for the bundle. The inlet and outlet kettle nozzles will have a definite pressure drop. It is best to locate the inlet nozzle on the side of the kettle and above the bundle. This keeps the pressure drop down because there are no tubes in the vicinity to provide a restriction. Fixed Tube Sheet Exchanger and High Shell Side Pressure Drop July, 2000 When there is a design problem meeting the allowable shell side pressure drop, reverse the stream sides. Since it is a fixed tube sheet exchanger, the unit can be designed with one (1) tube pass. Other types of heat exchangers can be designed with a single tube pass but they can have more operating problems. The pressure drop can be further reduced by using axial nozzles that are on the exchanger centerline. This eliminates large turning pressure drop losses. Impingement Rods January, 1997 When shell pressure drop is critical and impingement protection is required, use rods or tube protectors in top rows instead of a plate. These create less pressure drop and better distribution than an impingement plate. An impingement plate causes an abrupt 90 degree turn of the shell stream which causes extra pressure drop. Specifying Pressure Drop for Heavy Liquids Inside Tubes Frequently process engineers specify 5 or 10 psi for allowable pressure drop inside heat exchanger tubing. For heavy liquids that have fouling characteristics, this is usually not enough. There are cases where the fouling excludes using turbulators and using more than the customary tube pressure drop is cost effective. This is especially true if there is a relatively higher heat transfer coefficient on the outside of the tubing. The following example illustrates how allowable pressure drop can have a big effect on the surface calculation. A propane chiller was cooling a gas treating liquid that had an average viscosity of 7.5 cP. The effect on the calculated surface was as follows: You can see that using 25 psi pressure drop reduced the surface by nearly one-half. This would result in a price reduction for the heat exchanger of approximately 40%. This savings offset the cost of the pumping power. Maximum Velocity Inside Tubes An estimate for maximum tube velocity inside steel tubes 50 1,419 Allowable tube pressure drop, psi 25 2,104 Exchanger surface ft 2 5 4,012 V max = 10.2 ft/sec Where V max = maximum fluid velocity Density = 62.00 lb/cu ft. Calculate Shell Nozzle Pressure Drop Shell nozzle pressure drop calculation methods are difficult to find in the open literature. The nozzle pressure drops are difficult to predict accurately. There is a complex flow pattern of a tube matrix, bundle bypassing, and recirculation. Because of this, it is possible to have pressure loss coefficients greater than the customary 1.5 velocity heads for sharp edge expansion/contraction edges. If the bundle entrance area is equal to or greater than the inlet nozzle flow area, use a pressure loss coefficient of 1.0. If the bundle exit area is equal to or greater than the exit nozzle area, us a pressure loss coefficient of 0.58. There are indications that it should be larger. The following procedure is for the situation where the nozzle flow area is greater than the entrance or exit area and the bundles do not have an impingement plate. If there is an impingement plate, there will have to be added a turning loss to the calculation below. If the two shell side nozzles are not the same size, calculate the inlet pressure drop and take 2/3 of it and make a separate calculated pressure drop for the outlet and take 1/3 of it. Shell Entrance or Exit Area: 1. Calculate the bundle bypass area Sb = π x Dn x h 2. Calculate the slot area Aslot = 0.7854Dn 2 (Pt -Dt)/(F2 x Pt) 3. Calculate the shell entrance and exit area.(As) As = Sb + Aslot (refer TEMA RGP-RCB-4.621 & 4.622) 4. Calculate ratio of Sb to total area FR = Sb/As 5. Kn = 0.65 +2.14 (FR -0.4) (minimum Kn = 0.8, maximum = 1.8) 6. ΔPn = Kn x .000108Vs 2 x density (ΔPn = total of both nozzles) where ΔPn = Total nozzle pressure drop (lb/ft 2 ) Dn = Nozzle ID in. Ds = Shell ID in. Dt = Tube outside diameter in. F2 = 0.707 for 45 degree pitch, all others use 1.0 h = 0.5(Ds-OTL) in. Kn = Pressure loss coefficient OTL= Outer tube limit diameter in. Pt = Tube center to center pitch in. Vs = velocity in the entrance/exit area (ft/sec) EXAMPLE 990,000 lb/hr with a density of 62.4 lb/ft 3 is flowing through a 13.25 in. ID nozzle. The shell ID is 23.25 in. and the OTL is 22.375 in. The tube OD is 0.75 in. on a tube pitch of 0.9375 in. with 30 degree layout. 80 / sqrt(density) = fluid density = Calculate Sb h = 0.50(23.25-22.375)= 0.4375 Sb = π x 13.25 x 0.4375 = 18.23 Calculate Aslot Aslot = 0.7854(13.25 2 ) (0.9375-0.75)/(1.00 x .9375)= 27.58 Calculate total area As As = Sb + Aslot = 18.23 + 27.58 = 45.81 Calculate FR FR = 18.23/45.81 = 0.4 Calculate Kn Kn = 0.65 +2.14(0.4 -0.4) = 0.65 (use minimum 0.8) Calculate nozzle pressure drop Vs = (990000 x 0.04)/(45.81 x 62.4)= 13.85 ΔPn = 0.8 x 0.000108 x 13.85 2 x 62.4 = 1.03 psi Comment - Using 1.5 total pressure loss coefficient and the nozzle flow area gives only 0.21 PSI Improve Shell Side Pressure Drop Calculations The shell side pressure drop calculation can be improved by better equations for the baffle window and the nozzle pressure drops. Both of these methods can be found elsewhere on this web page. The baffle window pressure drop in the open literature is a function only of the number of tubes crossed and the velocity in the window. It does not take into account a friction factor, type of tube pattern or fluid eddies. When there are no tubes removed under the shell nozzles and the nozzles are large, using the nozzle flow area can result in wrong pressure drop calculations. This is taken from the first experimental case in “A Reappraisal of Shell side Flow in Heat Exchangers HTD-Vol. 36”. Average flow of 990,000 lb/hr with a density of 62.4 lb/ft 3 is flowing through a 13.25 ID nozzle. The shell ID is 23.25 in. and the OTL is 22.375 in. The effective tube length is 11.729 ft. The tube OD is 0.75 in. on a tube pitch of 0.9375 in. with 30 degree layout. There are 7 baffles and 26% baffle cut From the following the cross flow pressure drop is calculated: Bs = 17.6 in fi = 0.1025 - Ideal tube bank correlation ( J. Taborek) Nc = 13.75 Rb = 0.536 Re = 40,249 Rl = 0.615 ΔPc = 6.41 psi ΔPshell = ΔPc + ΔPw + ΔPn From other tips: ΔPw = 12.78 ΔPn = 1.03 ΔPshell = 6.41 +12.78 +1.03 = 20.2 psi Experimental = 20.3 psi One way would be to add a small amount to the tube diameter. This has a complex effect that is not linear in nature. A simpler method is to add 10% for each 0.001 increase in fouling factor. Then multiply this factor by the clean Allowable Shell Side Pressure Drop if a Multi-leaf (a.k.a. Lamaflex) Long Baffle is Used Four thin (0.008”) stainless strips are normally used to seal the sides of the long baffle. Because of their flexibility, they are not able to withstand large shell side pressure drops. It is best to limit the pressure drop to 5 psi with A new baffle window pressure drop equation has been published in the June 2004 issue of Hydrocarbon Processing. The name of the article is “More Accurate Exchanger Shell-Side Pressure Drop Calculations”. The article can be found on this page with the subject “Heat Exchanger Articles Published by Dale Gulley”. The equation improves the accuracy of the shell side pressure drop. Refer to the article for more detail. The equation has the following form: Distortion factor for ideal fluid stream. It varies with baffle cut. Refer article elsewhere on this site for This is taken from the first experimental case in “A Reappraisal of Shellside Flow in Heat Exchangers HTD-Vol. 36”. Average flow of 990,000 lb/hr with a density of 62.4 lb/ft3 is flowing through a 13.25 ID nozzle. The shell ID is 23.25 in. and the OTL is 22.375 in. The tube OD is 0.75 in. on a tube pitch of 0.9375 in. with 30 degree layout. The following are taken from a tip in this section named “Improve Shell Side Pressure Drop Calculations” When the calculated pressure drop inside the tubes is under-utilized, the estimated pressure drop with increased This would be a good estimate if advantage is not taken of the increase in heat transfer. Since the increased number of tube passes gives a higher velocity and increases the calculated heat transfer coefficient, the number of tubes to be used will decrease. The use of fewer tubes increases the new pressure drop. For a better estimate of the new Usually when shell-and-tube heat exchangers are designed, the tube layout is made so that the shell entrance area is approximately equal to the shell nozzle flow area. The average distance to the 1st tube row is Dn/4 where Dn is the inside diameter of the shell nozzle. In this case the pressure loss coefficient is 1.0 for the pressure drop calculation If the shell nozzles are greater than 2” and some tubes are not omitted from the tube layout, the nozzle entrance pressure drop can be significantly higher than the normal calculation based on the nozzle flow area. In a case of a 6” shell nozzle and where no tubes were omitted in a BEM type heat exchanger, the pressure drop was 3 times higher than that calculated with just the nozzle flow area. For more information, you can refer to the tip “Calculate Usually you will see the allowable pressure drop on the specification sheet for the shell side of a kettle reboiler to be stated as “nil”. This is close to being true only for the bundle. The inlet and outlet kettle nozzles will have a definite pressure drop. It is best to locate the inlet nozzle on the side of the kettle and above the bundle. This When there is a design problem meeting the allowable shell side pressure drop, reverse the stream sides. Since it is a fixed tube sheet exchanger, the unit can be designed with one (1) tube pass. Other types of heat exchangers can be designed with a single tube pass but they can have more operating problems. The pressure drop can be further reduced by using axial nozzles that are on the exchanger centerline. This eliminates large turning pressure drop losses. When shell pressure drop is critical and impingement protection is required, use rods or tube protectors in top rows instead of a plate. These create less pressure drop and better distribution than an impingement plate. An impingement plate causes an abrupt 90 degree turn of the shell stream which causes extra pressure drop. Frequently process engineers specify 5 or 10 psi for allowable pressure drop inside heat exchanger tubing. For heavy liquids that have fouling characteristics, this is usually not enough. There are cases where the fouling excludes using turbulators and using more than the customary tube pressure drop is cost effective. This is especially true if there is a relatively higher heat transfer coefficient on the outside of the tubing. The following example illustrates how allowable pressure drop can have a big effect on the surface calculation. A propane chiller was cooling a gas treating liquid You can see that using 25 psi pressure drop reduced the surface by nearly one-half. This would result in a price reduction for the heat exchanger of approximately 40%. This savings offset the cost of the pumping power. Shell nozzle pressure drop calculation methods are difficult to find in the open literature. The nozzle pressure drops are difficult to predict accurately. There is a complex flow pattern of a tube matrix, bundle bypassing, and recirculation. Because of this, it is possible to have pressure loss coefficients greater than the customary 1.5 velocity heads for If the bundle entrance area is equal to or greater than the inlet nozzle flow area, use a pressure loss coefficient of 1.0. If the bundle exit area is equal to or greater than the exit nozzle area, us a pressure loss coefficient of 0.58. There are indications that it should be larger. The following procedure is for the situation where the nozzle flow area is greater than the entrance or exit area and the bundles do not have an impingement plate. If there is an impingement plate, there will have to be added a turning loss to the calculation below. If the two shell side nozzles are not the same size, calculate the inlet pressure drop and take 2/3 of it and make a separate calculated pressure drop for the outlet and 990,000 lb/hr with a density of 62.4 lb/ft 3 is flowing through a 13.25 in. ID nozzle. The shell ID is 23.25 in. and The shell side pressure drop calculation can be improved by better equations for the baffle window and the nozzle The baffle window pressure drop in the open literature is a function only of the number of tubes crossed and the velocity in the window. It does not take into account a friction factor, type of tube pattern or fluid eddies. When there are no tubes removed under the shell nozzles and the nozzles are large, using the nozzle flow area can This is taken from the first experimental case in “A Reappraisal of Shell side Flow in Heat Exchangers HTD-Vol. 36”. Average flow of 990,000 lb/hr with a density of 62.4 lb/ft 3 is flowing through a 13.25 ID nozzle. The shell ID is 23.25 in. and the OTL is 22.375 in. The effective tube length is 11.729 ft. The tube OD is 0.75 in. on a tube pitch Features of a New S & T bundle to Replace Bundle That Vibrated 1. If possible, design for lower cross flow velocity with special baffles. 2. Make sure that impingement plate is very secure. 3. Use a tube/baffle clearance of 1/64. 4. Use thicker baffles. 5. Use closer baffle/shell clearance. 6. Use thicker tubes. 7. If tubes are low fins, have the tubing bare where it goes through the baffles. Vibration Cure When Designing Shell & Tube Bundles May, 2000 The cure depends upon whether it is flow induced or acoustical type vibration. Both types can be cured by using a lower cross flow velocity across the bundle. To do this, use double or triple segmental baffles. This not only lowers the velocity but the closer resulting baffle spacing increases the natural frequency of the bundle. Another possibility is to use a “No Tubes in Baffle Window” design. Then you can use as many baffle supports as necessary with very little effect on shell pressure drop. If the vibration is the acoustical type, use either 30 degree triangular pitch or square rotated pitch. The former is the best. Another cure is to use a de-resonating baffle. In a few cases, putting the problem stream inside the tubes would be better. Conditions Likely to Cause Shell & Tube Bundle Vibration May, 1997 Bundle vibration can cause leaks due to tubes being cut at the baffle holes or tubes being loosened at the tubesheet joint. There are services that are more likely to cause bundle vibration than others are. The most likely service to cause vibration is a single-phase gas operating at a pressure of 100 to 300 PSI. This is especially true if the baffle spacing is greater than 18 inches and single segmental type. Another service that sometimes causes bundle vibration is water in the shell side. Water has a relatively higher momentum than other most fluids. Therefore, if extra precautions on bundle design are not taken, a vibration problem can develop later when the exchanger goes into operation. Cures for Vibration in Existing Bundle September, 1997 Most flow-induced vibration occurs with the tubes that pass through the baffle window of the inlet zone. The unsupported lengths in the end zones are normally longer than those in the rest of the bundle. For 3/4 inch tubes, the unsupported length can be 4 to 5 feet. The cure for removable bundles, where the vibration is not severe, is to stiffen the bundle. This can be done by inserting metal slats or rods between the tubes under the nozzles. Normally this only needs to be done with the first few tube rows. Another solution is to add a shell nozzle opposite the inlet so as to cut the inlet fluid velocity in half. For non-removable bundles, this is the best solution. Adding a distributor belt on the shell would be a very good solution but it is expensive. If a U-tube bundle has a vibration problem in the bend area, metal slates or rods can be inserted between the tubes. If a slight decrease in heat transfer is not a problem, encircle the U-bends with a band or heavy wire and squeeze the tubes together. Best Design Feature to Prevent Bundle Vibration In designing a shell-and-tube heat exchanger, use a 30o triangular tube pitch if possible. This will lower the vortex shedding frequency which is a direct function of something called a Strouhal number. The Strouhal number is a constant composed of the vortex shedding frequency, shell side velocity and tube OD. The 30° triangular tube pitch has a significantly lower Strouhal number than other tube pitch types. Using Barrington as a source, for 3/4 inch tubes on 30o triangular tube pitch the Strouhal number is 0.21. But for 60o rotated triangular tube pitch the Strouhal number is 0.81. The cure depends upon whether it is flow induced or acoustical type vibration. Both types can be cured by using a lower cross flow velocity across the bundle. To do this, use double or triple segmental baffles. This not only lowers the velocity but the closer resulting baffle spacing increases the natural frequency of the bundle. Another possibility is to use a “No Tubes in Baffle Window” design. Then you can use as many baffle supports as necessary If the vibration is the acoustical type, use either 30 degree triangular pitch or square rotated pitch. The former is the best. Another cure is to use a de-resonating baffle. In a few cases, putting the problem stream inside the tubes would be better. Bundle vibration can cause leaks due to tubes being cut at the baffle holes or tubes being loosened at the tubesheet joint. There are services that are more likely to cause bundle vibration than others are. The most likely service to cause vibration is a single-phase gas operating at a pressure of 100 to 300 PSI. This is especially true if the baffle spacing is greater than 18 inches and single segmental type. Another service that sometimes causes bundle vibration is water in the shell side. Water has a relatively higher momentum than other most fluids. Therefore, if extra precautions on bundle design are not taken, a vibration problem can develop later when the exchanger goes into operation. Most flow-induced vibration occurs with the tubes that pass through the baffle window of the inlet zone. The unsupported lengths in the end zones are normally longer than those in the rest of the bundle. For 3/4 inch tubes, the unsupported length can be 4 to 5 feet. The cure for removable bundles, where the vibration is not severe, is to stiffen the bundle. This can be done by inserting metal slats or rods between the tubes under the nozzles. Normally this only needs to be done with the first few tube rows. Another solution is to add a shell nozzle opposite the inlet so as to cut the inlet fluid velocity in half. For non-removable bundles, this is the best solution. Adding a distributor belt If a U-tube bundle has a vibration problem in the bend area, metal slates or rods can be inserted between the tubes. If a slight decrease in heat transfer is not a problem, encircle the U-bends with a band or heavy wire and squeeze In designing a shell-and-tube heat exchanger, use a 30o triangular tube pitch if possible. This will lower the vortex shedding frequency which is a direct function of something called a Strouhal number. The Strouhal number is a constant composed The 30° triangular tube pitch has a significantly lower Strouhal number than other tube pitch types. Using Barrington as a source, for 3/4 inch tubes on 30o triangular tube pitch the Strouhal number is 0.21. But for 60o rotated triangular tube pitch the Allocation of Streams in Shell & Tube May, 1998 For those exchangers that need countercurrent flow, the stream with the highest pressure drop is usually best put in the tube side. This is true unless the design pressure is so high for the shell side that there would be material problems. High pressure drop instead of high design pressure is opposite of conventional thinking. If there are gas streams on both sides with mol. weights about the same and a small temperature difference, put the stream in the tubes with the highest value of the following: (#//hr)(#/hr)/op. pressure Otherwise, calculate the little more difficult term Vel x Vel x Density term for each side and put the stream with the highest value in the tubes. Heat Exchanger Articles Published by Dale Gulley 1. "More Accurate Exchanger Shell-and-Tube Pressure Drop Calculations", Hydrocarbon Processing, June 2004 2. "Troubleshooting Shell-and-Tube Heat Exchangers", Hydrocarbon Processing, September 1996 3. "Computers help Design Tubesheets", The Oil & Gas Journal, May 20,1974 4. "Computer Programs aid Design Work", The Oil & Gas Journal, Jan. 13,1969 5. "How to Calculate Weighted MTD's", Petroleum Refiner, July 1966 6. "How to Figure True Temperature Difference in Shell-and-Tube Exchangers", The Oil & Gas Journal, Sept. 14, 1964 7. "Make This Correction Factor Chart to Find Divided Flow Exchanger MTD", Petro/Chem Engineer, July 1962 8. "Use Computers to Select Exchangers", Petroleum Refiner, July 1960 Copies of the articles are available in .pdf format Avoid These Fluids When Using Low fin Tubing When a fluid has a high surface tension, the fluid doesn't readily flow from the gap between the tube fins. This adds resistance and lowers the heat transfer. The types of fluids that are to be avoided are those whose surface tension is above 30 to 40 dynes/cm. This includes such fluids as condensing steam, aqueous solutions with a high % of water, amines and glycols. Use Superficial velocities to Calculate Best Heat Transfer Flow Pattern The best heat transfer occurs when there is an annular flow pattern. Then there is a relatively thin liquid film and little vapor in contact with the heat transfer surface. How do you tell if the flow is annular? It will be when the superficial gas velocity is above the following value: VgMax = 1,832.0 where VL = the superficial liquid velocity = 5.0 ft/sec. Check Liquid Thermal Conductivity at High Reduced Temperatures November, 2000 There have been instances where process simulators have given results where the liquid thermal conductivity was nearly the same as the vapor thermal conductivity when the reduced temperature was still significantly lower than the critical temperature. Examine carefully the liquid thermal conductivity when its reduced temperature is above approximately 0.70. You may be able to justify a higher conductivity value and thus a higher heat transfer coefficient by using an independent and reliable correlation for the calculation. 72 -148VL +100 VL 2 = Check Piping Connections When There is Under-performance July, 2002 When a heat exchanger is installed and it is not achieving the desired heat duty, the first thing to check is the piping. Is the piping connected to the correct sides? It may be piped-up backwards. The worst case is when the shell side has a viscosity more than approximately 3 cP and there is no extra heat transfer enhancement inside the tubing. This could cause the fluid, when piped to the tube side, to be in laminar flow with its low heat transfer coefficient. Evaluating a Shell & Tube Exchanger For a New Service September, 2000 The best information to have for a shell and tube heat exchanger is a specification sheet and a full set of drawings. If both are not available, it is better to have the drawings. This is because they are more accurate on the mechanical details and they have tube layout details and seal bar information that the specification sheet does not have. What are most often missing on older heat exchangers are the bundle drawings. In this case, you need the original specification sheet. Then you can use its data and simulate the shell side heat transfer and pressure drop by running a thermal design Program to get a baffle configuration. Then this is used with the new process data to evaluate the new service. This procedure will not be as accurate as having the exact baffling but it is the best you can do if this is all you have to work with. Check Heat Release Curves for Skipping Over Dewpoints & Bubblepoints Frequently process engineers specify tabular heat release data that skips over dew points and bubble points. If equal increments of heat load or temperatures are used, chances are that the dew points and bubble points will be missed. It is important that the heat content at dew points or bubble points be shown. When Will Exchangers With Low-fins be More Economical Than Exchangers With Bare Tubes? 1. If the shell size is a least 2 sizes smaller (pipe size). 2. If the shell size is at least 14" O.D. 3. If there are fewer exchangers. when using low-fins 4. When (total shell resistance/total tube resistance) is greater than 0.4 Excess Heat Exchanger Surface Problems September, 2002 Excess surface does not always mean being safe. It can lead to control problems, pulsations, or freezing of condensate. Vaporization services and reboilers can particularly be a problem. Provide a way to control the flow of the heat medium in a new plant. In an existing installation without control, the boiling temperature difference may be so high that there is complete flashing of the liquid into vapor. Then the liquid feed rushes in to replace it which results in pulsations that may give downstream problems. The quickest solution is to either plug the tubes or put an orifice in the outlet vapor line to restrict the flow. Purchasing Shell & Tube Exchangers March, 2001 It is to the benefit of purchasers of shell and tube heat exchangers to not insist on applying their design. If the heat exchanger is to be built to TEMA requirements, it will void the guarantee. The last line of paragraph G5.2 says, “The thermal guarantee shall not be applicable to exchangers where the thermal performance was made by the purchaser”. Minimum Velocity inside Tubing for Slurries The minimum velocity for slurries inside tubes for shell-and-tube is 4 ft/sec. This is for a fine material like a catalyst. For slurries there is a special Reynolds number used for calculating the settling velocity. For more information on slurries, refer to chapter C11 in the piping handbook. Suggestions for Low-Fins and Potential S & T Bundle Vibration May, 2002 Tube bundles are more likely to vibrate if there is not a close clearance between the tubes and baffles. Low-fins are more susceptible to vibration because of the valleys between the fins. Another factor that makes them susceptible is that some low-fins are manufactured with the fin O.D. smaller than the bare ends. Some suggestions if the design software shows that the bundle may vibrate are: 1. Specify the low-fin tubing be bare where it passes through baffling. 2. Specify a tight tube hole tolerance. 3. Purchase tubing that has a fin O.D. the same as the bare ends. Shell & Tube or Multi-Tube? June, 1997 It is best to use Multi-tube (Hairpin) Exchangers instead of Shell & Tube when: 1. You require a small surface (less than 400 square feet); 2. There is a temperature cross in the heat transferred in a Shell & Tube; 3. The liquid flows are less than 150,000 lbs/hr; 4. Natural gas flows less than 1,200 X Sq. root(oper. pressure) Thermal Evaluation of Long Baffles August, 1997 The two thermal design problems associated with using two shell passes and a longitudinal baffle in Shell and Tube heat exchangers are: 1 Heat conduction through the baffle. There is a calculation method by Whistler. It is a correction applied to the LMTD. 2 Fluid by-pass around the long baffle. If possible, use an exchanger type where the long baffle is seal welded to the shell in order to avoid bypassing of the shell fluid. This should be done with a full penetration weld. The exchanger types, where the long baffles can be welded in, are Fixed Tube Sheets or U-Tubes. If U-tubes, the number of tube passes must be a multiple of four. Then the bundles can be removed. Other designs use multi-leaf long baffles for two shell passes. Since these cannot make a perfect seal, the amount of shell fluid bypassing the bundle must be calculated. Trouble-Shooting Article October, 1996 To find out more about heat exchangers, see Dale Gulley's article in the 1996 September issue of Hydrocarbon Processing. The title is “Troubleshooting Shell-and-Tube Heat Exchangers”. It gives helpful information on diagnosing problems. Undersurfaced S&T Quote When a vendor’s heat exchanger quote is under-surfaced, the following should be asked: 1. Are there seal strips? If so, how many? 2. What tube hole clearance was used in the baffles When to Add Shells in Series August, 2000 Usually you should design for the least number of shells for an item. However, there are times when it is more economical to add a shell in series to the minimum configuration. This will be when there is a relatively low flow in the shell side and the shell stream has the lowest heat transfer coefficient. This happens when the baffle spacing is close to the minimum. The minimum for TEMA is (Shell I.D./5). Then adding a shell in series gives a higher velocity and a much better heat transfer because of the smaller flow area in the smaller required exchangers. When to Consider a Long Baffle in the Shell October, 2001 The cost curve for a shell and tube heat exchanger decreases with increasing surface. The curve flattens at about 6,000 square feet of bare surface. If the first selection has multiple shells that are not countercurrent flow and each shell has less than 6,000 square feet, consider using a long baffle for cost savings. This is especially true if the exchanger is of a type where the long baffle can be welded to the shell (less likely to bypass fluid). Which Stream Goes Inside Tubes for Gas/Gas Exchangers? In a counter-current flow heat exchanger, the steam with the highest factor as calculated below goes inside the tubes: Factor = (flow) 2 / density You can also use the following factor if both gases’ molecular weight and temperature are about the same on both sides: Factor = (flow) 2 / pressure Why Did the Performance Decline in a TEMA F, G, or H Type Shell? Has performance declined after the bundle has been pulled and later installed back in the shell? If the longitudinal long baffle is sealed on the sides with leaf seals, they are probably the problem. These thin flexible strips should be positioned so that they form a concave pattern and flex upward. Then, when the shell fluid puts pressure on the leaves, they will press harder against the sides of the shell. If there is too much pressure - or if the bundle is installed upside down - the leaves will flex downward, and the shell fluid will bypass the bundle. Another possibility is that the leaf seals were damaged when the bundle was out of the shell. Fouling factors for water(hr-ft2-F/Btu) 0.0005steam,steam condensate,engine jacket water 0.0010boiler feed water 0.0015clean water,moutain water,etc. 0.0020normal cooling tower water For cooling water when velocity is 3 -8 ft/sec Fouling = 0.025/V 1.67 Where V =ft/sec Fouling Factors for Liquid Hydrocarbons(hr-ft2-F/Btu) 0.0010If sp. gravity At 60F less than 0.80, lube oil and heating oils 0.0020If sp. gravity At 60F 0.80 -0.87 0.0030If sp. gravity At 60F 0.87 -1.00 0.0050Heavy fuel oils Viscous Flow - Use More Pressure Drop Than Usual High viscosity fluids can have a problem achieving the design heat transfer. The fluids are usually petroleum based and have an API of 20 or less. Low pressure drops can cause maldistribution of the tubeside flow which in turn reduces the heat transfer. That is why you can see allowable pressure drops 2 or 3 times higher than usual. There is a method by A.C. Mueller for calculating this minimum allowable pressure drop. Another thing that can help is to use more tube passes and shorter tubes than normal. Also the fluid could be placed in the shell side if cleanig isn't a problem. For those exchangers that need countercurrent flow, the stream with the highest pressure drop is usually best put in the tube side. This is true unless the design pressure is so high for the shell side that there would be material problems. High pressure drop instead of high design pressure is opposite of conventional thinking. If there are gas streams on both sides with mol. weights about the same and a small temperature difference, put the stream in the tubes with the Otherwise, calculate the little more difficult term Vel x Vel x Density term for each side and put the stream with the 1. "More Accurate Exchanger Shell-and-Tube Pressure Drop Calculations", Hydrocarbon Processing, June 2004 6. "How to Figure True Temperature Difference in Shell-and-Tube Exchangers", The Oil & Gas Journal, Sept. 14, 7. "Make This Correction Factor Chart to Find Divided Flow Exchanger MTD", Petro/Chem Engineer, July 1962 When a fluid has a high surface tension, the fluid doesn't readily flow from the gap between the tube fins. This adds resistance and lowers the heat transfer. The types of fluids that are to be avoided are those whose surface tension is above 30 to 40 dynes/cm. This includes such fluids as condensing steam, aqueous solutions with a high % of water, The best heat transfer occurs when there is an annular flow pattern. Then there is a relatively thin liquid film and little vapor in contact with the heat transfer surface. How do you tell if the flow is annular? It will be when the superficial November, 2000 There have been instances where process simulators have given results where the liquid thermal conductivity was nearly the same as the vapor thermal conductivity when the reduced temperature was still significantly lower than the critical temperature. Examine carefully the liquid thermal conductivity when its reduced temperature is above approximately 0.70. You may be able to justify a higher conductivity value and thus a higher heat transfer coefficient When a heat exchanger is installed and it is not achieving the desired heat duty, the first thing to check is the piping. Is the piping connected to the correct sides? It may be piped-up backwards. The worst case is when the shell side has a viscosity more than approximately 3 cP and there is no extra heat transfer enhancement inside the tubing. This could cause the fluid, when piped to the tube side, to be in laminar flow with its low heat transfer coefficient. The best information to have for a shell and tube heat exchanger is a specification sheet and a full set of drawings. If both are not available, it is better to have the drawings. This is because they are more accurate on the mechanical details and they have tube layout details and seal bar information that the specification sheet does not have. What are most often missing on older heat exchangers are the bundle drawings. In this case, you need the original specification sheet. Then you can use its data and simulate the shell side heat transfer and pressure drop by running a thermal design Program to get a baffle configuration. Then this is used with the new process data to evaluate the new service. This procedure will not be as accurate as having the exact baffling but it is the best you can do if this is all you have January, 1998 Frequently process engineers specify tabular heat release data that skips over dew points and bubble points. If equal increments of heat load or temperatures are used, chances are that the dew points and bubble points will be missed. Excess surface does not always mean being safe. It can lead to control problems, pulsations, or freezing of condensate. Vaporization services and reboilers can particularly be a problem. Provide a way to control the flow of the heat medium in a new plant. In an existing installation without control, the boiling temperature difference may be so high that there is complete flashing of the liquid into vapor. Then the liquid feed rushes in to replace it which results in pulsations that may give downstream problems. The quickest solution is to either plug the tubes or put an orifice in It is to the benefit of purchasers of shell and tube heat exchangers to not insist on applying their design. If the heat exchanger is to be built to TEMA requirements, it will void the guarantee. The last line of paragraph G5.2 says, “The thermal guarantee shall not be applicable to exchangers where the thermal performance was made by the purchaser”. The minimum velocity for slurries inside tubes for shell-and-tube is 4 ft/sec. This is for a fine material like a catalyst. For slurries there is a special Reynolds number used for calculating the settling velocity. For more information on May, 2002 Tube bundles are more likely to vibrate if there is not a close clearance between the tubes and baffles. Low-fins are more susceptible to vibration because of the valleys between the fins. Another factor that makes them susceptible is that some low-fins are manufactured with the fin O.D. smaller than the bare ends. Some suggestions if the design The two thermal design problems associated with using two shell passes and a longitudinal baffle in Shell and Tube To find out more about heat exchangers, see Dale Gulley's article in the 1996 September issue of Hydrocarbon Processing. The title is “Troubleshooting Shell-and-Tube Heat Exchangers”. It gives helpful information on Usually you should design for the least number of shells for an item. However, there are times when it is more economical to add a shell in series to the minimum configuration. This will be when there is a relatively low flow in the shell side and the shell stream has the lowest heat transfer coefficient. This happens when the baffle spacing is close to the minimum. The minimum for TEMA is (Shell I.D./5). Then adding a shell in series gives a higher velocity and a much better heat transfer because of the smaller flow area in the smaller required exchangers. The cost curve for a shell and tube heat exchanger decreases with increasing surface. The curve flattens at about 6,000 square feet of bare surface. If the first selection has multiple shells that are not countercurrent flow and each shell has less than 6,000 square feet, consider using a long baffle for cost savings. This is especially true if the exchanger is of a type where the long baffle can be welded to the shell (less likely to bypass fluid). In a counter-current flow heat exchanger, the steam with the highest factor as calculated below goes inside the tubes: You can also use the following factor if both gases’ molecular weight and temperature are about the same on both sides: Has performance declined after the bundle has been pulled and later installed back in the shell? If the longitudinal long baffle is sealed on the sides with leaf seals, they are probably the problem. These thin flexible strips should be positioned so that they form a concave pattern and flex upward. Then, when the shell fluid puts pressure on the leaves, they will press harder against the sides of the shell. If there is too much pressure - or if the bundle is installed upside down - the leaves will flex downward, and the shell fluid will bypass the bundle. Another possibility is that High viscosity fluids can have a problem achieving the design heat transfer. The fluids are usually petroleum based Low pressure drops can cause maldistribution of the tubeside flow which in turn reduces the heat transfer. That is why you can see allowable pressure drops 2 or 3 times higher than usual. There is a method by A.C. Mueller for calculating this minimum allowable pressure drop. Another thing that can help is to use more tube passes and shorter Art Montemayor September 30, 2005 Rev: 0 A E L B F M C G N N H P Channel integral with tubesheet Split Flow Shell Fixed tubesheet; like "C" & removable cover. stationary head. Bundle Channel integral with tubesheet Double split flow Outside, packed floating head & removable cover. Shown: Removable Tube Stationary head. Bonnet (Integral Cover) 2-pass shell with longitudinal Fixed tubesheet; like "B" baffle stationary head. TEMA DESIGNATIONS Front End Stationary Head Shell Type Rear End Stationary Head Channel and removable cover One-pass shell Fixed tubesheet; like "A" Page 282 of 327 FileName: 136851736.xls.ms_office WorkSheet: TEMA Designations Art Montemayor September 30, 2005 Rev: 0 D J S Conventional Front End Heads: A or, B K T Other popular rear end head types employed: U W Kettle type of reboiler Pull-through floating head U-tube bundle design (No Rear Head Required) Packed floating tubesheet with lantern ring device (split-ring) Special, high-pressure closure Divided shell flow Floating head with backing Page 283 of 327 FileName: 136851736.xls.ms_office WorkSheet: TEMA Designations Art Montemayor September 30, 2005 Rev: 0 Some examples of the TEMA designation for Heat Exchangers are shown below: Front bonnet (Intergral Cover), with one-Pass Shell and a Fixed Tubesheet rear Bonnet Fixed tubesheet heat exchanger. This is a very popular version as the heads can be removed to clean the inside of the tubes. The front head piping must be unbolted to allow front head removal; if this is undesirable, then this can be avoided by applying a type A front head. In that case only the cover needs to be removed. It is not possible to mechanically clean the outside surface of the tubes as these are fixed inside the shell. Chemical cleaning can be used in the shell side. Shown is a version with one shell pass and two tube passes. This is probably the least expensive of the shell-and-tube designs. This is the same type of heat exchanger as shown above, except it has only one tube pass Channel with Removable Cover, One Pass Shell, Fixed Tubesheet Bonnet This is almost the same type of heat exchanger as the first BEM. The removable cover allows the inside of the tubes to be inspected and cleaned without unbolting the piping. However, as can be expected, the tradeoff is that this convenient feature makes it more expensive. BEM BEM AEM Page 284 of 327 FileName: 136851736.xls.ms_office WorkSheet: TEMA Designations Art Montemayor September 30, 2005 Rev: 0 The maintenance feature of having a removable tube bundles requires an exchanger as the following: Channel and Removable Cover, One Pass Shell, Floating Head with Backing Device A floating head heat exchanger is excellent for applications where the difference in temperature between the hot and cold fluid causes unacceptable stresses in the axial direction, between the shell and tubes. The floating head can move, i.e. it provides the ability to allow tube expansion in the axial direction. Note that the bundle can not be pulled from the front end. For maintenance both the front and rear end head, including the backing device, must be disassembled. If pulling from the front head is required a type AET should be selected. However, it is wise and prudent to be aware of the inherent trade-offs in this design. Note that the tube-side fluid can leak through the internal floating head cover gasket and mix (or contaminate) the shell-side fluid. It is very difficult -and sometimes impossible to mitigate or compensate for the internal bolts tightening the internal bonnet to remain under constant, steady torque. Hot fluid temperatures make the bolts expand and the result is a reduction in bolt torque and subsequent leaks through the bonnet gasket. Additionally, it is a common and expected occurance for maintenance crews to find the internal bolts badly rusted or corroded to the point where they have to be burned or sawed off in order to extract the "removable" tube bundle. The chemical engineer has other options to apply when requiring mechanical expansion of a heat exchanger tube bundle. Various rear head design also exist that allow for tube bundle expansion. Among these are the popular (and inexpensive) "U" tube bundle design. A "P" and "W" rear head design will also contribute this feature without the hazard of internal mixing (or contamination) of the two fluids. Also, be aware that any TEMA shell and tube design with a removable tube bundle feature has - by nature - a larger shell diameter (& increased cost) due to the need to be able to pull the rear tubesheet the length of the exchanger's shell. A larger diameter shell can sometimes also present problems in a lower Reynolds number (yielding a lower heat transfer) and internal by-passing of the shell fluid around the baffles (this also reduces the effective heat transferred. All these effects eventually lead to a bigger heat exchanger (more area and more tubes) in order to do a heat transfer operation. AES Page 285 of 327 FileName: 136851736.xls.ms_office WorkSheet: TEMA Designations Art Montemayor September 30, 2005 Rev: 0 Longitudinal Baffles - their application and inherent problems The employment of longitudinal baffles in heat exchangers - such as the "F", "G", and "H" shell types - can often resolve both heat transfer and fluid flow problems within the shell and tube exchanger used. Their application can significantly increase the shell-side Reynolds Number and lead to more efficient shell-side heat transfer coefficients with a subsequent increase in heat transfer. Additionally, these type of baffles permit the engineer to incorporate counter-flow heat transfer. True counter-current heat transfer is as efficient a heat transfer configuration as an engineer can obtain. In some heat recovery applications, this is highly sought. By splitting the shell-side flow, some applications can actually have a significant reduction in shell-side pressure drop. This is especially true in partial vacuum process operations where a minimum of pressure drop can be tolerated. However, the application of longitudinal baffles should be always carefully scrutinized and used sparingly. There are, as would be expected, some very important trade-offs involved in the application of longitudinal baffles. Firstly, if a longitudinal baffle is a process necessity, the baffle should be seal-welded against the inner shell wall in order to ensure that there will be no internal, by-pass leakage. This positive step negates the possibility of having a removable tube bundle. Additionally, the welding necessity requires a minimum shell diameter and this winds up being applicable only to relatively large streams. By the basic need to establish effective shell-side flow around a longitudinal baffle, one has to accept the obvious fact that a minimum of shell-side clearances can be tolerated. Once having said and applied these facts, one then has to also accept that the required, small baffle clearances mean extraordinary fabrication techniques and resultant super-human maintenance efforts to extract a removable tube bundle. In far too many actual field cases, it has been found that the removable tube bundle with a longitudinal baffle is a non-practical device. Field results have shown that in most cases the tube bundle has resulted in being destroyed in order to remove it. This extraordinary and desperate maintenance act labels such a design as non-practical. Page 286 of 327 FileName: 136851736.xls.ms_office WorkSheet: TEMA Designations Art Montemayor September 30, 2003 Rev: 0 Source: "Applied Process Design for Chemical and Petrochemical Plants"; Vol. 3; p.24 Ernest E. Ludwig; Gulf Publishing Co.; Houston, TX (1965) Tube O. D. & Pitch 8 10 12 13-1/4 15-1/4 17-1/4 19-1/4 21-1/4 23-1/4 25 27 29 31 33 35 37 3/4" on 15/16" Triang. 33 69 105 135 193 247 307 391 481 553 663 763 881 1,019 1,143 1,269 3/4" on 1" Triang. 33 57 91 117 157 217 277 343 423 493 577 667 765 889 1,007 1,127 3/4" on 1" Square 33 53 85 101 139 183 235 287 355 419 495 587 665 765 865 965 1" on 1-1/4" Triang. 15 33 57 73 103 133 163 205 247 307 361 427 481 551 633 699 1" on 1-1/4" Square 17 33 45 65 83 111 139 179 215 255 303 359 413 477 545 595 3/4" on 15/16" Triang. 32 58 94 124 166 228 300 370 452 528 626 734 846 964 1,088 1,242 3/4" on 1" Triang. 28 56 90 110 154 208 264 326 398 468 556 646 746 858 972 1,088 3/4" on 1" Square 26 48 78 94 126 172 222 280 346 408 486 560 644 746 840 946 1" on 1-1/4" Triang. 16 32 52 62 92 126 162 204 244 292 346 410 462 530 608 688 1" on 1-1/4" Square 12 26 40 56 76 106 136 172 218 248 298 348 402 460 522 584 3/4" on 15/16" Triang. 8 34 64 94 134 180 234 304 398 460 558 648 768 882 1,008 1,126 3/4" on 1" Triang. 8 26 60 72 108 158 212 270 336 406 484 566 674 772 882 1,000 3/4" on 1" Square 12 30 52 72 100 142 188 242 304 362 436 506 586 688 778 884 1" on 1-1/4" Triang. XX 8 26 42 58 84 120 154 192 234 284 340 396 466 532 610 1" on 1-1/4" Square XX 12 22 38 58 76 100 134 180 214 256 304 356 406 464 526 3/4" on 15/16" Triang. XX 48 84 108 154 196 266 332 412 484 576 680 788 904 1,024 1,172 3/4" on 1" Triang. XX 44 72 96 134 180 232 292 360 424 508 596 692 802 912 1,024 3/4" on 1" Square XX 48 72 88 126 142 192 242 308 366 440 510 590 688 778 880 1" on 1-1/4" Triang. XX 24 44 60 78 104 138 176 212 258 308 368 422 486 560 638 1" on 1-1/4" Square XX 24 40 48 74 84 110 142 188 214 260 310 360 414 476 534 3/4" on 15/16" Triang. XX 28 56 84 122 166 218 286 378 438 534 622 740 852 976 1,092 3/4" on 1" Triang. XX 20 52 64 98 146 198 254 318 386 462 542 648 744 852 968 3/4" on 1" Square XX 24 44 64 90 130 174 226 286 342 414 482 560 660 748 852 1" on 1-1/4" Triang. XX 20 36 50 74 110 142 178 218 266 322 376 444 508 584 1" on 1-1/4" Square XX 16 32 50 66 90 122 166 198 238 286 336 384 440 500 3/4" on 15/16" Triang. XX 80 116 174 230 294 372 440 532 632 732 844 964 1,106 3/4" on 1" Triang. XX 66 104 156 202 258 322 388 464 548 640 744 852 964 3/4" on 1" Square XX 54 78 116 158 212 266 324 394 460 536 634 224 818 1" on 1-1/4" Triang. XX 34 56 82 112 150 182 226 274 338 382 442 514 586 1" on 1-1/4" Square XX 44 66 88 116 154 184 226 268 318 368 430 484 3/4" on 15/16" Triang. XX 74 110 156 206 272 358 416 510 596 716 826 944 1,058 3/4" on 1" Triang. XX 56 88 134 184 268 300 366 440 518 626 720 826 940 3/4" on 1" Square XX 56 80 118 160 210 268 322 392 458 534 632 718 820 1" on 1-1/4" Triang. XX 30 42 68 100 130 168 206 252 304 356 426 488 562 1" on 1-1/4" Square XX 42 60 80 110 152 182 224 268 316 362 420 478 3/4" on 15/16" Triang. XX 94 140 198 258 332 398 484 576 682 790 902 1,040 3/4" on 1" Triang. XX 82 124 170 224 286 344 422 496 588 694 798 902 3/4" on 1" Square XX 94 132 174 228 286 352 414 490 576 662 760 1" on 1-1/4" Triang. XX 66 90 120 154 190 240 298 342 400 466 542 1" on 1-1/4" Square XX 74 94 128 150 192 230 280 334 388 438 3/4" on 15/16" Triang. XX 68 102 142 190 254 342 398 490 578 688 796 916 1,032 3/4" on 1" Triang. XX 52 82 122 170 226 286 350 422 498 600 692 796 908 3/4" on 1" Square XX 48 70 106 146 194 254 306 374 438 512 608 692 792 1" on 1-1/4" Triang. XX 24 38 58 90 118 154 190 238 290 340 404 464 540 1" on 1-1/4" Square XX 34 50 70 98 142 170 206 254 300 344 396 456 Notes: 1) The above tube counts have an allowance made for Tie Rods. 2) The Radius of Bend for the U-Tube bundles is equal to (2.5) (Tube O.D.); The actual number of U-tubes is 1/2 of the above figures. E i g h t - P a s s F i x e d T u b e s U T u b e s F o u r - P a s s F i x e d T u b e s U T u b e s S i x - P a s s Heat Exchanger Tube Sheet Layout Count Table Shell I. D., inches O n e - P a s s F i x e d T u b e s F i x e d T u b e s U T u b e s T w o - P a s s F i x e d T u b e s U T u b e s Page 287 of 327 FileName: 136851736.xls.ms_office WorkSheet: Tube Counts Art Montemayor March 12, 2001 Rev: 0 T in, Cold Side (t 1 ) 69 o F T out, Cold Side (t 2 ) 83 o F T in, Hot Side (T 1 ) 169 o F T out, Hot Side (T 2 ) 128 o F Exchanger Heat Duty 3,950 M Btu/hr Overall U, estimated 100 Btu/hr - Ft 2 - o F Number of shell passes 1 Number of tube passes 2 Log Mean Temperature Difference, LMTD 72 o F F Factor (see below) 0.98 Adjusted LMTD 70 o F Heat Transfer Area calculated 562 Ft 2 Design contingency factor 1.25 Over-design allowance 1.00 Heat Transfer Area required 702 Ft 2 450 psig, Saturated Steam Req'd, 5,163 lbs/hr CW Req'd @ 14 deg rise, gpm 564 gpm Calculation of F Factor: P (or S) 0.14 R 2.93 Term 1 0.69 [(RP-1)/(P-1)] (1/N) Px 0.14 Term 2 1.60 (R^2+1) 0.5/(R-1) Term 3 0.38 1.46 Term 4A 13.45 Term 4B 7.26 Term 4 0.62 F 0.98 HEAT EXCHANGER SUMMARY q, U, A, AT m w, c p , t 1 W, C p , T 2 W, C p , T 1 w, c p , t 2 Page 288 of 327 FileName: 136851736.xls.ms_office WorkSheet: HX Design Sheet 1 of 1 Corporation 1 2 3 Mfr Ref. No. * No. Req'd 4 TEMA Size, Type Horiz. Vert. Connected in Series Parallel 5 ft 2 Gross Eff. Shells/Unit One Surface/Shell * ft 2 Gross Eff. 6 Plot Plan No. Other Ref. Dwg No. 7 PERFORMANCE OF ONE UNIT 8 9 10 11 12 Liquid lb/h 13 Steam lb/h 14 Non-Condensables lb/h 15 Fluid Vaporized or Condensed lb/h 16 Steam Condensed lb/h 17 ºF 18 19 Viscosity cP 20 Vapor Molecular Weight 21 Specific Heat Btu/lb·ºF 22 Thermal Conductivity Btu/h·ft·ºF 23 Latent Heat Btu/lb 24 Operating Pressure, Inlet psig 25 Velocity Max. Min. fps 26 Pressure Drop, Clean (Allow./Calc.) psi 27 Fouling Resistance ft 2 ·h·ºF/Btu . 28 Heat Exchanged Btu/h Log MTD (Uncorrected) ºF Log MTD (Corrected) ºF 29 Transfer Rate, Service Btu/ft 2 ·h·ºF . Btu/ft 2 ·h·ºF . 30 CONSTRUCTION AND MATERIALS 31 SHELL SIDE TUBE SIDE Sketch (Bundle, Nozzle Orientation) 32 Design Pressure psig 33 Test Pressure psig 34 Design Temperature ºF 35 Number of Passes per Shell 36 In 37 Out 38 Intermediate 39 Tubes: Type Number OD in. BWG or in. X Min. Av. Wall 40 Tube Length in. Tube Pitch in. Flow Pattern (circle one) 41 Shell: ID in. OD in. Tube-to-Tubesheet Joint 42 Baffles - Cross: Type Spacing in. % Cut on X Diam. Area 43 Baffles - Long: Perm. Removable Seal Type: Bypass Seal: 44 Inlet Nozzle lb/ft·sec Bundle Entrance lb/ft·sec Bundle Exit lb/ft·sec 45 Expansion Joint? Yes X No Type: Impingement Protection? X Yes No 46 47 Tubes Floating Tubesheet 48 Shell Fixed Tubesheet * 0.125 49 Shell Cover Tube Supports * 0.125 50 Channel Cross Baffles * 0.125 51 Channel Cover Long Baffle * 0.125 52 Fltg Head Cover Gaskets 53 § Stress Relieved (Mark "SR') and/or Radiographed (Mark 'XR') Parts User Spec.: 54 Code Requirements: ASME Sec. VIII, Para. 1 (1992) Stamp? 55 Weights: Shell lb Filled with Water lb Bundle lb 56 Remarks 1. Items marked with an asterisk (*) to be completed by Vendor. 57 58 Rev Date Description By Chk. Appr. Rev Date Description By Chk. Appr. * * 0 For Purchase * Stainless Stl ---- Yes Carbon Steel TEMA Class: Carbon Steel PART MATERIAL § THK, in. C.A., in. Carbon Steel Carbon Steel C.A., in. Stainless Stl 16 BWG min. Carbon Steel * ---- PART MATERIAL § THK, in. µv 2 : * * * * * * 16 0.9375 * Transfer Rate, Clean * * Rolled and Seal Welded * 0.75 * Temperature Density, Specific Gravity * Vapor (In/Out) lb/h Fluid Circulated Total Fluid Entering lb/h Fluid Allocation SHELL SIDE TUBE SIDE Manufacturer * Model * Surface/Unit * SHELL & TUBE HEAT EXCHANGER SPECIFICATION (English Units) Project No. Location Unit P&ID No. Service Lean MEA Solution Cooler Equipment No. P.O. No. Connections Size & 60° 90° 45° 30° R e v . N o . Montemayor PLATE & FRAME HEAT EXCHANGER SPECIFICATION Sheet 1 of 1 (English Units) Corporation 1 2 3 Model 4 5 ft 2 ft 2 6 7 8 Fluid Allocation 9 Fluid Circulated 10 Total Fluid Entering 11 Vapor (In/Out) 12 Liquid 13 Steam 14 Non-Condensables 15 Fluid Vaporized or Condensed 16 Steam Condensed 17 Temperature 18 Density, Specific Gravity 19 Viscosity 20 Vapor Molecular Weight 21 Specific Heat 22 Thermal Conductivity 23 Latent Heat 24 Operating Pressure, Inlet 25 Velocity X Max. Min. 26 Pressure Drop, Clean (Allow./Calc.) 27 Fouling Resistance 28 Heat Exchanged ºF ºF 29 30 31 Sketch (Frame, Nozzle Orientation) 32 33 34 35 36 in. 37 In 38 Out 39 Intermediate 40 41 42 Frame Capacity (Max. No. of Plates) 43 44 45 46 47 48 49 50 51 52 lb lb 53 54 55 Date Description By Chk. Appr. Rev Date Description By Chk. Appr. Rev 0 9-Dec-96 For Inquiry ABC DEF XYZ Remarks 1. Items marked with an asterisk (*) to be completed by Vendor. Filled with Water * Client Spec.: Weights: Empty Frame * End Cover Carbon Steel * 0.03125 Carbon Steel 0.03125 Cleaning: Painting: Code Requirements: ASME Sec. VIII, Para. 1 (1992) Stamp? Yes Carbon Steel Insulation: MATERIAL § THK, in. C.A., in. 0.03125 Carbon Steel § Stress Relieved (Mark "SR') and/or Radiographed (Mark 'XR') Parts OSHA Type Protective Shroud? Yes Material: 0.03125 Plate Gaskets Carbon Steel * 0.03125 Frame 0.03125 Carrying Bar Carbon Steel Heat Conservation Plates Stnless Steel 16 BWG min. 0.03125 Connections Stnless Steel Carbon Steel 0.03125 PART MATERIAL § THK, in. C.A., in. PART None No. of Plates 3" 150# RF 6" 125# FF µv 2 , Inlet/Outlet lb/ft·s 300 ---- ---- Impingement Protection? Yes Corrosion Allowance 0.0625 3" 150# RF 6" 125# FF Number of Passes per Frame Two * Test Pressure psig Design Temperature ºF 300 Code Code CONSTRUCTION AND MATERIALS Allocation HOT SIDE COLD SIDE Design Pressure psig 150 125 fps 8.0 8.0 Transfer Rate, Service * Btu/ft 2 ·h·ºF Transfer Rate, Clean Btu/ft 2 ·h·ºF Btu/h Log MTD (Uncorrected) * ft 2 ·h·ºF/Btu 0.001 0.003 3,097,238 157.0 Log MTD (Corrected) * Btu/h·ft·ºF 0.178 0.160 0.358 0.365 Btu/lb ---- psi 10 Btu/lb·ºF 0.867 ---- psig 75 60 0.843 1.0 0.65 ---- ---- ---- ---- 10 * 1.0 * cP 0.54 13.7 0.76 105 0.907 0.929 0.995 0.992 ºF 235 120 90 ---- lb/h ---- ---- ---- ---- lb/h ---- ---- ---- ---- lb/h ---- ---- ---- ---- lb/h ---- ---- ---- lb/h 31,500 206,483 ---- ---- ---- ---- lb/h 31,500 31,500 206,483 206,483 lb/h Cooling Water Mfr Ref. No. * No. Req'd PERFORMANCE OF ONE UNIT P&ID No. Plot Plan No. Other Ref. Dwg No. Effective Surface/Frame Gross * One Size, Type * - * Frames/Unit One Connected in Cooling Water Exchanger Equipment No. Surface/Unit * * HOT SIDE COLD SIDE Single Manufacturer * P.O. No. Project No. 1234567 Location Unit Service Connections Size & R e v . N o . Montemayor Art Montemayor Overall Heat Transfer Coefficient October 02, 2003 Rev: 0 Source: http://www.the-engineering-page.com/forms/he/typU.html Hot Fluid Cold Fluid Water Water 800 – 1,500 140 - 264 Organic solvents Organic Solvents 100 - 300 17 – 52 Light oils Light oils 100 - 400 17 – 70 Heavy oils Heavy oils 50 - 300 9 – 53 Reduced crude Flashed crude 35 - 150 6 – 26 Regenerated DEA Fouled DEA 450 - 650 79 – 114 Gases (p = atm) Gases (p = atm) 5 - 35 1.0 – 6 Gases (p = 200 bar) Gases (p = 200 bar) 100 - 300 17 – 53 Organic solvents Water 250 - 750 44 – 132 Light oils Water 350 - 700 62 - 123 Heavy oils Water 60 - 300 11 - 53 Reduced crude Water 75 - 200 13 – 35 Gases (p = atm) Water 5 - 35 1.0 – 6 Gases (p = 200 bar) Water 150 - 400 26 – 70 Gases Water 20 - 300 4 – 53 Organic solvents Brine 150 - 500 26 – 88 Water Brine 600 – 1,200 106 – 211 Gases Brine 15 - 250 3 - 44 Steam Water 1,500 – 4,000 264 - 700 Steam Organic solvents 500 – 1,000 88 - 176 Steam Light oils 300 - 900 53 – 159 Steam Heavy oils 60 - 450 11 – 79 Steam Gases 30 - 300 5 – 53 Heat Transfer (hot) Oil Heavy oils 50 - 300 9 – 53 Heat Transfer (hot) Oil Gases 20 - 200 4 - 35 Flue gases Steam 30 - 100 5 - 18 Flue gases Hydrocarbon vapors 30 -100 5 - 18 Aqueous vapors Water 1,000 – 1,500 176 – 264 Organic vapors Water 700 – 1,000 123 – 176 Refinery hydrocarbons Water 400 - 550 70 - 97 Vapors with some non condensables Water 500 - 700 88 – 123 Vacuum condensers Water 200 - 500 35 – 88 Heaters Condensers Coolers Typical Overall Heat Transfer Coefficients Shell and Tube Heat Exchangers Overall “U” W/m 2 -C Btu/hr-ft 2 - o F Heat Exchangers Page 291 of 327 FileName: 136851736.xls.ms_office WorkSheet: Typical "U" Art Montemayor Overall Heat Transfer Coefficient October 02, 2003 Rev: 0 Steam Aqueouos solutions 1,000 – 1,500 176 – 264 Steam Light organics 900 – 1,200 159 – 211 Steam Heavy organics 600 - 900 106 – 159 Heat Transfer (hot) oil Refinery hydrocarbons 250 - 550 44 – 97 300 - 450 53 - 79 300 - 700 53 - 123 50 - 150 9 - 26 50 - 300 9 - 53 300 - 600 53 - 106 Coil Fluid Pool Fluid Steam Dilute aqueous solutions 500 – 1,000 88 – 176 Steam Light oils 200 - 300 35 – 53 Steam Heavy oils 70 - 150 12 – 26 Aqueous solutions Water 200 - 500 35 – 88 Light oils Water 100 - 150 18 – 26 Steam Dilute aqueous solutions 800 – 1,500 140 – 264 Steam Light oils 300 - 500 53 – 88 Steam Heavy oils 200 - 400 35 – 70 Aqueous solutions Water 400 - 700 70 - 123 Light oils Water 200 - 300 35 - 53 Jacket Fluid Vessel Fluid Steam Dilute aqueous solutions 500 - 700 88 - 123 Steam Light organics 250 - 500 44 - 88 Water Dilute aqueous solutions 200 - 500 35 - 88 Water Light organics 200 - 300 35 - 53 Art’s Note: Above U’s were originally given in metric units and the conversion to good, old fashioned US engineering units is based on: 1.0 Btu/hr-ft 2 - o F = 5.678263 Watts/m 2 - o K Immersed coils Natural circulation Agitated Process Fluid (tube side) Water Jacketed vessels Heavy organics Gases Condensing hydrocarbons Light organics Vaporizers Air Cooled Exchangers Page 292 of 327 FileName: 136851736.xls.ms_office WorkSheet: Typical "U" Art Montemayor G KETTLES RECIRCULATION COOLER E-G-43 October 24, 1997 ID = Shell Internal diameter, in. = 23 OD T = Tube external diameter, in. = 0.75 P T = Tubes' Pitch, in. = 0.9375 C' = Clearance between tubes = 0.1875 B = Baffle spacing, in. = 15 N = Number of Shell-side baffles = 11 a S = Shell-side crossflow area, ft 2 = 0.4792 W = Shell-side mass flowrate,lb/h = 325500 G S = Shell-side unit mass flowrate,lb/h-ft2 = 679,304 D e = Equivalent shell diameter, ft = 0.045833 N Re = Shell-side Reynolds Number = 643.3 f = Friction factor for pressure drop, ft 2 /in 2 = 0.00352 c P = Fluid's Heat Capacity, Btu/lb- o F = 0.52 µ = Fluid's Viscosity, cP = 20 µ' = Fluid's Viscosity, lb/ft-h 48.4 k = Fluid's therm. cond., Btu/ft-h- o F = 0.086 N Pr = Fluid's Prandtl Number = 292.7 us = Viscosity ratio, (µ/µ w ) 0.14 = 1 s = Shell fluid's specific gravity = 0.930372 AP = Shell-side pressure drop, psi = 16.8 From "Process Heat Transfer"; D. Kern; McGraw-Hill; 1950; pages 147-148 AP = f G 2 s D s (N+1)/(5.22 x 10 10 )D e s u s Page 293 of 327 Electronic FileName: 136851736.xls.ms_office WorkSheet: Shell-Side Pressure drop Quick & Dirty Tubular Heat Exchanger Rating Sheet Project Project No. Item No. E-G-XX Service By Date/Time 13-Mar-97 15:12 Step 1. Input flows, conditions and properties data for shellside and Step 4. Start configuring the exchanger. Begin with the total calculated tubeside. transfer coefficients to this point (i.e., not including shellside h): Tube Side Shell U start = 235 Btu/h·ft 2 · o F CW Fluid Name Warm Water On that basis, assumed U o = 195 Btu/h·ft 2 · o F 418,000 Flow (M), lb/h 195,000 Then the required transfer A = 2,139 ft 2 88 Temp. in, o F 130 Number of tubes required = 545 102 Temp. out, o F 100 Reset tubes/pass (Step 3), then no. of passes = 4 Av. Density 62.05 µ, lb/ft 3 61.9 Total tube count = 584 Av. Viscosity 0.723 µ, cP 0.590 Tubeside AP (incl. returns) = 8.1 psi OK? Av. Heat Capacity 1 c p , Btu/lb· o F 1 Actual effective transfer area, A = 2,293 ft 2 Heat Exchanged 5,850,015 Q, Btu/h 5,850,000 OK? Av. Thermal Conductivity 0.360 k, Btu/h·ft· o F 0.368 Fouling Resistance 0.002 R, ft 2 · h· o F/Btu 0.0015 Step 5. Select tube arrangement Tube Pitch 0.9375 in. Prandtl No. 4.86 c p µ/k 3.88 and estimate shell diameter Pattern Tri Uncorrected MTD 18.9 o F Shell ID from Tube Count Tables 27 in. Corrected MTD 14.0 o F Select Baffle Spacing 16 in. Number of Baffles = 14 Flow Area across Bundle, a s = 0.600 ft 2 Step 2. Input tubing OD, BWG and Tube OD 0.7500 in. Equivalent Diameter, d e (see table) = 0.55 in. length (can be trial and error). BWG 16 Mass Velocity, G s = 325,000 lb/h·ft 2 Tube ID, d = 0.620 in. Shellside Reynolds No., N Re = 25,258 Tube Length, L = 20 ft. Shellside Friction Factor = 0.00178 Flow area per tube, a t = 0.302 in. 2 Shellside AP = 2.7 psi OK? Effective transfer area per tube = 3.927 ft 2 Outside Transfer Factor, j h = 90.4 Outside Film Coefficient, h o = 1,140 Calculated U o = 195.1 Step 3. Estimate the number of Tubes/pass = 146 Check: % difference, U calc. vs U assum. = 0.0% OK? tubes per tube pass. lb/h per tube = 2,863 U clean = 614.9 Av. velocity, fps = 6.11 OK? Tubeside Reynolds No., N Re = 40,324 Tubeside Friction Factor, f = 0.010 Step 6. Check tubeside velocity and AP, shellside AP. If too high or too low, AP per pass, psi = 1.01 OK? adjust tube length, number of tubes per pass, number of passes, and/or shell Inside Transfer Factor, j h = 113.7 baffle spacing. Remember to reset shell diameter from tube count tables, as Inside Film Coefficient, h i = 1,335 required. Reactor Warm Water System Upgrade Quick & Dirty Tubular Heat Exchanger Rating Sheet Btu/h·ft 2 · o F Btu/h·ft 2 · o F Step 6. Check tubeside velocity and AP, shellside AP. If too high or too low, adjust tube length, number of tubes per pass, number of passes, and/or shell baffle spacing. Remember to reset shell diameter from tube count tables, as Signature Date 18-Feb-04 Checked Date Proj No. Project File Subject Sheet 1 of 10 1 2 A horizontal, 1-2 condenser is required for condensing pure propyl alcohol emanating from the top of a distillation 3 column. Side-to-side, 25% cut segmental baffles will be used. Basic data is as follows: 4 5 Propanol flowrate 60,000 lb/hr 6 Propanol vapors' inlet pressure 15.0 psig 7 Propanol vapors' inlet temperature 244 o F 8 Cooling Water inlet temperature 85 o F 9 Propanol allowable pressure drop 2.00 psi 10 CWS allowable pressure drop 10.00 psi 11 Dirt factor 0.003 12 Condenser tubes' length 8.00 feet 13 Tubes' OD 0.7500 inches 14 Tubes' length 8.00 feet 15 Tubes' gauge 16 BWG 16 Tubes' ID 0.6200 inches 17 Tubes' pitch 0.9375 Triangular, inches 18 Clearance between tubes 0.1875 inches 19 Propanol Latent Heat at 15 psig 285 Btu/lb 20 Propanol Molecular Weight 60.1 21 22 23 24 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 44 A B C D E F G H I J K L CALCULATION SHEET Art Montemayor D. Q. Kern, "Process Heat Transfer"; Mc-Graw Hill; 1950; p. 274 Horizontal n-Propanol Total Condenser T e m p e r a t u r e , o F 85 244 Distance along tubes Vapo Condensate Cooling water Cooling water in Signature Date 18-Feb-04 Checked Date Proj No. Project File CALCULATION SHEET Art Montemayor D. Q. Kern, "Process Heat Transfer"; Mc-Graw Hill; 1950; p. 274 Subject Sheet 2 of 10 45 46 First, make a heat and material balance to establish the heat load and the cooling water required: 47 48 Propanol latent heat for condensation = Btu/hr 49 Cooling water terminal temperature = 120 o F 50 Cooling water required = 488,571 lb/hr = 976 gpm 51 Propanol Water Differ. 52 244 Higher Temperature 120 124 53 244 Lower Temperature 85 159 54 0 35 35 55 56 141 o F 57 58 Since the shell side Propanol vapor is essentially isothermal, the exchanger is in true counterflow. 59 60 T c = The caloric temperature of the hot fluid 61 t c = The caloric temperature of the cold fluid 62 t a = The average temperature of the cold fluid 63 The influence of the tube-wall temperature is included in the condensing film coefficient. 64 65 102.5 o F can be used as the caloric temperaure of the cold fluid 66 67 Execute a trial calculation: 68 a) 100 Btu/hr- o F-ft 2 69 Condensing film coefficients will generally range from 150 to 300. Assuming a film coefficient of 1,000 70 for water, U C will range from 130 to 230 Btu/hr- o F-ft 2. 71 72 1,215 ft 2 73 773 74 75 b) Assume that 4 tube passes are used. The quantity of water is large, but the condenser will have a 76 large number of tubes, making a 2-pass assumption inadvisable. 77 From the tube counts table, 4 tube passes using 3/4" OD tubes on 15/16" triangular pitch , yields a 78 count of 766 tubes in a 31 inch ID shell. 79 c) The corrected U D coefficient, using the 31" shell, is now calculated: 80 1,203 ft 2 81 Q/A AT = 101 Btu/hr- o F-ft 2 82 83 84 85 86 87 88 A B C D E F G H I J K L Quantity of 3/4" OD tubes = Corrected area, A = Corrected U D = Horizontal n-Propanol Total Condenser 17,100,000 Difference Log Mean Temperature Difference = LMTD = Heat transfer area = A = Q/U D AT = The mean t a = Assume that U D = Signature Date 18-Feb-04 Checked Date Proj No. Project File CALCULATION SHEET Art Montemayor D. Q. Kern, "Process Heat Transfer"; Mc-Graw Hill; 1950; p. 274 Subject Sheet 3 of 10 89 90 Calculations for shell side hot fluid, n-Propanol 91 92 Assume a maximum baffle spacing. This will be 32-1/2", 31", and 32-1/2" which is equal to 96" or 2 baffles and 93 3 crosses for the proposed side-to-side flow. Since these are the minimum baffles that can be used, this should 94 yield the lowest attainable shell-side pressure drop in this configuration. 95 96 1.33 ft 2 (Eq. 7.1; p.138) 97 where, 98 ID = Shell inside diameter, inches 99 C' = Clearance between tubes, inches 100 B = Baffle spacing, inches 101 P T = Tube pitch, inches 102 103 44,953 lb/hr-ft 2 104 105 89.6 lb/hr-linear ft 106 where, 107 L = Tube length, feet 108 N t = Tube quantity effective for condensation 109 110 200 Btu/hr-ft 2 - o F 111 112 h iO = 1,075 Btu/hr-ft 2 - o F (Refer to line # 165) 113 where, 114 h iO = The inside film (water) heat transfer coefficient refered to the tube OD, Btu/hr-ft 2 - o F 115 h i = The inside (water) film heat transfer coefficient = 1,300 Btu/hr-ft 2 - o F (From fig. 25 ) 116 117 125 o F (Eq. 5.31; p. 98) 118 where, 119 T v = Average temperature of hot fluid (vapor), o F 120 121 184 o F 122 123 0.095 Btu/hr-ft 2 - o F/ft (From Table 4) 124 125 0.80 (From Table 6) 126 127 0.62 cP (From Fig. 14) 128 129 130 131 132 133 A B C D E F G H I J K L Specific Gravity of shell side film = s f = Viscosity of shell side film = µ f = Shell side film thermal conductivity = k f = Horizontal n-Propanol Total Condenser The shell-side or bundle crossflow area = a S =(ID) (C') (B)/(P T * 144) = The shell-side mass velocity = G s = W / a S = The condensate loading on the horizontal tubes = G'' = W/L*N t 2/3 = Assume the value of the average condensing film coefficient = h O = h i (ID/OD) = Tube wall temperature = t W = t a + [h O /(h iO + h O )] (T v - t a ) = Shell side film temperature = t f = (T v + t w )/2 = Signature Date 18-Feb-04 Checked Date Proj No. Project File CALCULATION SHEET Art Montemayor D. Q. Kern, "Process Heat Transfer"; Mc-Graw Hill; 1950; p. 274 Subject Sheet 4 of 10 134 135 h' ( µ f 2 / k f 3 µ f 2 g) 1/3 = 1.5 (4 G''/µ f ) -1/3 (Equation 12.42; p. 266) 136 where, 137 h' = Average condensing film coefficient, Btu/hr-ft 2 - o F 138 µ f = = 1.5004 lb/ft-hr 139 k f = film coefficient thermal conductivity, Btu/hr-ft 2 - o F/ft 140 µ f = film coefficient density, lb/ft 3 = 49.92 141 g = Acceleration of gravity, ft/hr 2 = 142 G'' = Condensate loading for horizontal tubes, lb/hr-ft 143 144 Average shell side condensing film coefficient = 178 Btu/hr-ft 2 - o F 145 146 Calculations for tube side cold fluid, Water 147 148 0.3020 in 2 (From condenser tube table) 149 150 0.402 ft 2 151 where, 152 N T = Number of tubes effective for condensation 153 a' t = Flow area per tube, in 2 154 n = Number of tube passes 155 156 w / a t = lb/hr-ft 2 157 158 5.41 ft/sec 159 160 102.5 o F: 161 162 0.72 cP = 1.74 lb/ft-hr 163 0.0517 ft 164 165 D G t /µ = 36,073 166 167 1,300 Btu/hr-ft 2 - o F (From Fig. 25) 168 169 h iO = 1,075 Btu/hr-ft 2 - o F 170 where, 171 h iO = The inside film (water) heat transfer coefficient refered to the tube OD, Btu/hr-ft 2 - o F 172 173 Based on h' = 172 instead of the assumed 200, a new value of tw and tf could be obtained to give a more exact value 174 of h' based on the fluid properties at a value of tf more nearly correct. However, it is not necessary in this example 175 because the condensate properties will not change materially. 176 177 178 A B C D E F G H I J K L Reynolds Number (for pressure drop only) = 1,216,508 Average water velocity in the tube side = V = G t / (3,600*µ) = Tube side water heat transfer film coefficient = h i = h i (ID/OD) = Horizontal n-Propanol Total Condenser film coefficient absolute viscosity 4.18E+08 Flow area of a 3/4" OD x 16 BWG tube = Flow area per tube = N T a' t / 144 n = Water mass velocity in the tube side = G t = At the average water temperature, t a , of Water viscosity = µ = Tubes' ID = Signature Date 18-Feb-04 Checked Date Proj No. Project File CALCULATION SHEET Art Montemayor D. Q. Kern, "Process Heat Transfer"; Mc-Graw Hill; 1950; p. 274 Subject Sheet 5 of 10 179 180 Calculations for shell side pressure drop 181 182 244 o F 183 Propanol vapor viscosity = 0.010 cP = 0.0242 lb/ft-hr (From fig. 15) 184 185 Shell-side equivalent diameter (D e ): 186 The hydraulic radius employed for correlating shell-side coefficients for bundles having baffles is not the true hydraulic 187 radius. The direction of flow in the shell is partly along an d partly at right angles to the long axes of the bundle's tubes. 188 The flow area at right angles to the long axes is variable from tube row to tube row. A hydraulic radius based upon 189 the flow area across any one row could not distinguish between square and triangular pitch. In order to obtain a simple 190 correlation combining both the size and closeness of the tubes and their type of pitch, excellent agreement is 191 obtained if the hydraulic radius is calculated along (instead of across) the long axes of the tubes. 192 193 D e = (4 * free area)/(wetted area) = [(4) (0.5 *P T * 0.86 * P T - 0.5 * t * d 2 /40] / (0.5 * t *d) 194 = 0.55 inches = 0.0458 ft (From fig. 28) 195 196 85,139 197 198 0.00141 ft 2 /in 2 (From fig. 29) 199 200 3 201 202 Assume that the propanol vapor follows the ideal gas law at the low pressure. 203 204 0.236 lb/ft 3 205 206 0.00378 207 208 2.58 ft 209 210 1.2 psi (Eq. 12.47; p.273) 211 212 213 Calculations for tube side pressure drop 214 215 For the tube side Reynolds Number = 36,073 the corresponding tube-side friction factor 216 217 f = 0.00019 ft 2 /in 2 (From fig. 26) 218 219 Straight tube pressure drop + Return Loss pressure drop 220 221 3.3 psf = 0.02 psi 222 (Eq. 7.45; p. 148) 223 A B C D E F G H I J K L Tube-side pressure drop = Straight tube pressure drop = AP t = f * G t 2 * Ln/(5.22*10 10 *D e * s *u t ) = Shell-side Reynolds Number = D e G s /µ = Propanol vapor density = MW / (V 1 ) (T 2 /T 1 ) (P 1 /P 2 ) = Propanol vapor specific gravity = s = Shell Inside Diameter = D s = Shell-side pressure drop = (1/2) [ f *G s 2 D s (N+1) /(5.22 * 10 10 *D e * s)] = Shell-side friction factor for 25% cut segmental baffles = f = Number of shell-side crosses = (N+1) = Horizontal n-Propanol Total Condenser The propanol vapor temperature = Signature Date 18-Feb-04 Checked Date Proj No. Project File CALCULATION SHEET Art Montemayor D. Q. Kern, "Process Heat Transfer"; Mc-Graw Hill; 1950; p. 274 Subject Sheet 6 of 10 224 225 7.3 psi (Eq. 7.46; p.148) 226 227 7.29 psi 228 where, 229 L = tube length, feet 230 n = Number of tube passes 231 u t = The viscosity ratio (µ/µ w )0.14 in the tubes 232 g' = Acceleration of gravity, 32.2 ft/sec 2 233 234 235 Calculation of clean overall coefficient U C : 236 237 152.4 Btu/hr-ft 2 - o F 238 239 240 Calculation of dirt factor R d : 241 242 101 Btu/hr- o F-ft 2 (From line 81) 243 244 0.0033 hr-ft 2 - o F/Btu 245 (Note: In condensation calculations the omission of the tube metal resistance may introduce a significant error and 246 should be checked.) 247 248 249 178 1,075 250 U C = 152.4 251 U D = 101 252 0.0033 253 0.003 254 1.2 7.29 255 2.00 10.00 256 257 Conclusion: 258 The first trial calculated is satisfactory and yields the following exchanger: 259 260 ID = 31 inches 766; 8' - 0" 261 31 inches (approx.) 3/4"; 16 BWG; 15/16", triangular 262 1 4 263 264 It is interesting at this point to compare a vertical condenser with this horizontal model. The horizontal and vertical 265 266 number of tubes in both models is the same. To this end a vertical condenser will be assumed which uses the same 267 tube count as the above except that the tube length may be 12 or 16 ft (as needed) to account for the lower 268 coefficients obtained in the vertical orientation. A B C D E F G H I J K L Shell side Tube side Quantity and length = Baffle spacing = OD, BWG, & pitch = Passes = Passes = condensing film coefficients are both affected by W and N t , and the best basis fof comparison is otained when the Shell side Summary of Results Tube side h (outside) R d calculated = R d required = Horizontal n-Propanol Total Condenser Calculated AP Allowable AP Total tube-side pressure drop = U C = (h io * h o )/(h io + h o ) = Corrected U D = R d = (U C - U D )/(U C * U D ) = Return Loss pressure drop = AP r = (4*n/s) (V 2 /2 g') = Signature Date 18-Feb-04 Checked Date Proj No. Project File CALCULATION SHEET Art Montemayor D. Q. Kern, "Process Heat Transfer"; Mc-Graw Hill; 1950; p. 274 Subject Sheet 7 of 10 269 270 The vertical condenser to be rated will be oriented as seen in the sketch below. The process conditions will be 271 identical to those of the previous horizontal model rated. In order to prevent water corrosion in the carbon steel shell, 272 the water will also be introduced in the tube side. 273 274 275 276 277 278 279 280 281 282 283 284 285 286 287 288 289 290 291 292 293 Btu/hr 294 295 141 o F 296 297 298 102.5 o F 299 300 Trial Calculation: 301 302 a) Assume that the overall dirty heat transfer coefficient, U D = 70 Btu/hr-ft 2 - o F 303 The equation for the condensing film coefficient gives greater values for horizontal tubes than for vertical tubes. 304 It will, consequently, be necessary to reduce the value of U D . 305 306 1,735 ft 2 307 308 The nearest common, available tube length (using the same 766 tubes) is: 309 ## 11.5 feet ( use 12 foot length tubes ) 311 312 313 A B C D E F G H I J K L Vertical n-Propanol Total Condenser Total heat transferred = 17,100,000 Caloric temperature of the water = t C = Heat transfer area = A = Q/U D * AT = Tube length = Log Mean Temperature Difference = LMTD = Caloric temperature of the Propanol vapor = T C Vapor Condensate Cooling water Cooling water in Signature Date 18-Feb-04 Checked Date Proj No. Project File CALCULATION SHEET Art Montemayor D. Q. Kern, "Process Heat Transfer"; Mc-Graw Hill; 1950; p. 274 Subject Sheet 8 of 10 314 315 b) The same tube layout, using 3/4" OD x 16 BWG tubes on 15/16" triangular pitch and 4 passes will also be 316 used. 317 318 c) The corrected U D coefficient, using the 31" shell, is now calculated: 319 1,804 ft 2 320 Q/A AT = 67 Btu/hr- o F-ft 2 321 322 323 Calculations for shell side hot fluid, n-Propanol 324 325 0.0625 ft 326 327 399 lb/hr-lin. ft (Eq. 12.36; p. 265) 328 329 100 Btu/hr-ft 2 - o F 330 331 h iO = 1,075 Btu/hr-ft 2 - o F (Refer to line # 165) 332 where, 333 h iO = The inside film (water) heat transfer coefficient refered to the tube OD, Btu/hr-ft 2 - o F 334 h i = The inside (water) film heat transfer coefficient = 1,300 Btu/hr-ft 2 - o F (From fig. 25 ) 335 336 114.5 o F (Eq. 5.31; p. 98) 337 where, 338 T v = Average temperature of hot fluid (vapor), o F 339 340 179 o F 341 342 0.095 Btu/hr-ft 2 - o F/ft (From Table 4) 343 344 0.80 (From Table 6) 345 346 0.65 cP (From Fig. 14; also, 4*G'/µ = 1,025) 347 348 h' ( µ f 2 / k f 3 µ f 2 g) 1/3 = 1.47 (4 G'/µ f ) -1/3 (Equation 12.39; p. 266) 349 where, 350 h' = Average condensing film coefficient, Btu/hr-ft 2 - o F 351 µ f = film coefficient absolute viscosity = 1.573 lb/ft-hr 352 k f = film coefficient thermal conductivity, Btu/hr-ft 2 - o F/ft 353 µ f = film coefficient density, lb/ft 3 = 49.92 354 g = Acceleration of gravity, ft/hr 2 = 355 G' = Condensate loading for vertical tubes, lb/hr-ft 356 357 Average shell side condensing film coefficient = 104 Btu/hr-ft 2 - o F 358 A B C D E F G H I J K L Viscosity of shell side film = µ f = 4.18E+08 Shell side film temperature = t f = (T v + t w )/2 = Shell side film thermal conductivity = k f = Specific Gravity of shell side film = s f = Vertical n-Propanol Total Condenser Corrected area, A = Q/U D *At = Corrected U D = Tubes' outside diameter, Do = Condensate loading for vertical tubes = W/N t * t * D o = Assume the value of the average condensing film coefficient = h O = h i (ID/OD) = Tube wall temperature = t W = t a + [h O /(h iO + h O )] (T v - t a ) = Signature Date 18-Feb-04 Checked Date Proj No. Project File CALCULATION SHEET Art Montemayor D. Q. Kern, "Process Heat Transfer"; Mc-Graw Hill; 1950; p. 274 Subject Sheet 9 of 10 359 360 Calculations for tube side cold fluid, Water 361 362 The tube-side water conditions and configuration is the same as the horizontal configuration. 363 364 h iO = 1,075 Btu/hr-ft 2 - o F 365 where, 366 h iO = The inside film (water) heat transfer coefficient refered to the tube OD, Btu/hr-ft 2 - o F 367 368 369 Calculations for shell side pressure drop 370 371 It is necessary to arrange the 12-foot tube bundle into a minimum number of bundle crosses, or (N + 1) = 5. 372 The spacing between baffles will be: 373 B = 29 inches 374 375 1.24 ft 2 (Eq. 7.1; p.138) 376 where, 377 ID = Shell inside diameter, in. 378 C' = Clearance between tubes, in. 379 B = Baffle spacing, in. 380 P T = Tube pitch, in. 381 382 48,387 lb/hr-ft 2 383 384 244 o F 385 Propanol vapor viscosity = 0.010 cP = 0.0242 lb/ft-hr (From fig. 15) 386 387 0.0458 ft (From table in fig. 28) 388 389 91,642 390 391 0.00140 ft 2 /in 2 (From fig. 29) 392 393 5 394 395 0.00378 396 397 2.58 ft 398 399 2.3 psi (Eq. 12.47; p.273) 400 401 This pressure drop prediction is high, and if it cannot be compensated for by elevating the condenser, it will be 402 necessary to use the half-circle (50% cut) support baffles as shown in Example 7-8. 403 A B C D E F G H I J K L Propanol vapor specific gravity = s = (Same as line 206) Shell Inside Diameter = D s = Shell-side pressure drop = (1/2) [ f *G s 2 D s (N+1) /(5.22 * 10 10 *D e * s)] = Equivalent diameter for pressure drop = D e = Shell-side Reynolds Number = Re S = D e * G S / µ = Shell-side friction factor for 25% cut segmental baffles = f = Number of shell-side crosses = (N+1) = Vertical n-Propanol Total Condenser h i (ID/OD) = Shell-side (bundle) crossflow area = a s = (ID * C' * B)/(P T * 144) = The shell-side mass velocity = G s = W / a S = The propanol vapor temperature = Signature Date 18-Feb-04 Checked Date Proj No. Project File CALCULATION SHEET Art Montemayor D. Q. Kern, "Process Heat Transfer"; Mc-Graw Hill; 1950; p. 274 Subject Sheet 10 of 10 404 405 Calculations for tube side pressure drop 406 407 The basic data is the same as in the horizontal model example, except for the tube length. 408 409 5.0 psf = 0.03 psi 410 (Eq. 7.45; p. 148) 411 412 7.3 psi (Eq. 7.46; p.148) 413 414 7.30 psi 415 416 417 Calculation of clean overall coefficient U C : 418 419 95.0 Btu/hr-ft 2 - o F 420 421 422 Calculation of dirt factor R d : 423 424 67 Btu/hr- o F-ft 2 (From line 81) 425 426 0.0043 hr-ft 2 - o F/Btu 427 428 429 430 104 1,075 431 U C = 95.0 432 U D = 67 433 0.0043 434 0.003 435 2.3 7.30 436 2.00 10.00 437 438 Conclusion: 439 Shell side Tube side 440 ID = 31 inches 766; 12' - 0" 441 29 inches (approx.) 3/4"; 16 BWG; 15/16", triangular 442 1 4 443 444 This vertical condenser is somewhat secure in performing the specified heat transfer duty but it exceeds the 445 allowable pressure drop, although not seriously. The advantage of horizontal condensation may be observed 446 from the U C of 148.5 in the horizontal condenser as compared with the 93.2 in the vertical unit in identical service. 447 The vertical unit has an inherent advantage, however, when the condensate is to be subcooled. 448 A B C D E F G H I J K L Passes = Passes = Shell side Summary of Results Tube side h (outside) R d calculated = R d required = Calculated AP Allowable AP f * G t 2 * Ln/(5.22*10 10 *D e * s *u t ) = Return Loss pressure drop = AP r = (4*n/s) (V 2 /2 g') = Total tube-side pressure drop = U C = (h io * h o )/(h io + h o ) = Quantity and length = Baffle spacing = OD, BWG, & pitch = Corrected U D = R d = (U C - U D )/(U C * U D ) = Vertical n-Propanol Total Condenser Straight tube pressure drop = AP t = Some of this data was taken from Standards of the Tubular Exchanger Manufacturers Association (TEMA); 7th Edition (1988); page 178. Note: some of the tabular TEMA data contained ERRATA, but this was corrected with this spreadsheet's formulas. Outside Inside 7 8 0.165 9 0.148 10 0.134 0.482 11 0.120 0.510 12 0.109 0.282 0.0625 0.1309 0.0738 0.456 0.532 13 0.095 0.560 14 0.083 0.334 0.0876 0.1309 0.0874 0.370 0.584 15 0.072 0.606 16 0.065 0.370 0.1075 0.1309 0.0969 0.302 168 0.620 17 0.058 0.634 18 0.049 0.402 0.1269 0.1309 0.1052 0.236 198 0.652 20 0.035 0.430 0.1452 0.1309 0.1126 0.174 227 0.680 22 0.028 0.444 0.1548 0.1309 0.1162 0.141 241 NOTES: * The weight of the condenser tubes is based on low carbon steel with a density of 0.2836 lbs/in 3 . For other metal materials multiply by the following factors: Factor 0.35 0.58 0.99 1.02 1.04 1.06 1.07 1.09 1.13 1.12 1.14 ** Liquid Velocity within the tubes = (Lbs Per Tube Hour) / (C * Liquid Specific Gravity) in feet per sec. (Specific gravity of Water @ 60 o F = 1.00) Copper and Cupro-Nickels Material Aluminum Titanium A.I.S.I. 300 Series Stainless Steels A.I.S.I. 400 Series Stainless Steels Aluminum Bronze Aluminum Brass Nickel-Chrome-Iron Admiralty Nickel-Copper Nickel Surface area per linear foot, ft 2 Tube weight per linear foot, lb of steel* Constant C ** Tube I. D. inches BWG Wall thickness inches 1/2" O. D. Condenser tube 3/4" O. D. Condenser tube Tube I. D. inches Tube flow area in 2 Some of this data was taken from Standards of the Tubular Exchanger Manufacturers Association (TEMA); 7th Edition (1988); page 178. Note: some of the tabular TEMA data contained ERRATA, but this was corrected with this spreadsheet's formulas. Outside Inside Outside Inside 0.670 0.3526 0.2618 0.1754 0.704 0.3893 0.2618 0.1843 0.1825 0.1963 0.1262 0.883 285 0.732 0.4208 0.2618 0.1916 0.2043 0.1963 0.1335 0.808 319 0.760 0.4536 0.2618 0.1990 0.2223 0.1963 0.1393 0.747 347 0.782 0.4803 0.2618 0.2047 0.2463 0.1963 0.1466 0.665 384 0.810 0.5153 0.2618 0.2121 0.2679 0.1963 0.1529 0.592 418 0.834 0.5463 0.2618 0.2183 0.2884 0.1963 0.1587 0.522 450 0.856 0.5755 0.2618 0.2241 0.3019 0.1963 0.1623 0.476 471 0.870 0.5945 0.2618 0.2278 0.3157 0.1963 0.1660 0.429 492 0.884 0.6138 0.2618 0.2314 0.3339 0.1963 0.1707 0.367 521 0.902 0.6390 0.2618 0.2361 0.3632 0.1963 0.1780 0.268 567 0.930 0.6793 0.2618 0.2435 * The weight of the condenser tubes is based on low carbon steel with a density of 0.2836 lbs/in 3 . For other metal materials multiply by the following factors: ** Liquid Velocity within the tubes = (Lbs Per Tube Hour) / (C * Liquid Specific Gravity) in feet per sec. (Specific gravity of Water @ 60 o F = 1.00) 1" O. D. Condenser tube Surface area per linear foot, ft 2 Tube I. D. inches Tube flow area in 2 3/4" O. D. Condenser tube Surface area per linear foot, ft 2 Tube weight per linear foot, lb of steel Constant C ** Tube flow area in 2 Some of this data was taken from Standards of the Tubular Exchanger Manufacturers Association (TEMA); 7th Edition (1988); page 178. Note: some of the tabular TEMA data contained ERRATA, but this was corrected with this spreadsheet's formulas. Outside Inside 0.890 0.6221 0.3272 0.2330 2.059 970 1.473 550 0.920 0.6648 0.3272 0.2409 1.914 1,037 1.170 1.348 0.954 0.7148 0.3272 0.2498 1.744 1.200 1.241 656 0.982 0.7574 0.3272 0.2571 1.599 1,182 1.230 1.129 708 1.010 0.8012 0.3272 0.2644 1.450 1,250 1.260 1.038 749 1.030 0.8332 0.3272 0.2697 1.341 1,305 1.280 0.919 804 1.060 0.8825 0.3272 0.2775 1.173 1,377 1.310 0.814 852 1.080 0.9161 0.3272 0.2827 1.059 1,440 1.330 0.714 898 1.110 0.9677 0.3272 0.2906 0.883 1.360 0.650 927 1.120 0.9852 0.3272 0.2932 0.824 1,537 1.370 0.584 1.130 1.0029 0.3272 0.2958 0.763 1.380 0.498 997 1.150 1.0387 0.3272 0.3011 0.641 1,626 1.400 0.361 1,060 1.180 1.0936 0.3272 0.3089 0.455 1,706 Surface area per linear foot, ft 2 Tube weight per linear foot, lb of steel Constant C ** Tube I. D. inches 1-1/2" O. D. Condenser tube 1" O. D. Condenser tube 1-1/4" O. D. Condenser tube Tube weight per linear foot, lb of steel Constant C ** Tube I. D. inches Tube flow area in 2 Outside Inside Outside Inside 1.0751 0.3927 0.3063 2.355 1.1310 0.3927 0.3142 2.165 1.1882 0.3927 0.3220 1.970 1,860 1.2469 0.3927 0.3299 1.771 1.760 2.4328 0.5236 0.4608 1.2868 0.3927 0.3351 1.635 2,014 1.782 2.4941 0.5236 0.4665 1.3478 0.3927 0.3430 1.427 1.810 2.5730 0.5236 0.4739 1.3893 0.3927 0.3482 1.286 2,180 1.834 2.6417 0.5236 0.4801 1.4527 0.3927 0.3560 1.070 1.4741 0.3927 0.3587 0.997 2,300 1.4957 0.3927 0.3613 0.924 1.5394 0.3927 0.3665 0.775 Tube flow area in 2 Surface area per linear foot, ft 2 Tube weight per linear foot, lb of steel Tube flow area in 3 Surface area per linear foot, ft 3 1-1/2" O. D. Condenser tube 2" O. D. Condenser tube Constant C ** Tube I. D. inches 2.412 3,795 2.204 3,891 1.935 4,014 1.701 4,121 Constant C ** Tube weight per linear foot, 2" O. D. Condenser tube Art Montemayor Heat Exchanger Tubesheets Tubesheet Thickness October 09, 1991 Rev: 0 F = 1.25 G = 12 inches P = 350 psig S = 17,500 psi T = 1.06 inches T = F = = = G = P = design pressure, psig S = tubesheets' material allowable stress, psi 100 200 300 400 500 17,500 17,500 17,500 17,500 17,500 -- 17,700 16,100 15,900 -- 15,000 15,000 15,000 15,000 15,000 17,500 16,500 15,500 14,800 14,700 -- 12,500 10,500 2,000 -- 12,500 10,500 10,400 10,400 10,400 6,600 5,700 5,000 -- -- Tubesheet thickness, inches a factor 1.0 for stationary and floating-head tubesheets SB-402 Copper Nickel 1.25 for U-tube tubesheets shell internal diameter, as calculated from transfer surface and tube dimensions, inches Values of S for some common materials are shown in the following table. With this table and the other terms, SB-11 Copper tubesheet thickness can be calculated in this spreadsheet. Material Temperature, o F SA-516 Grade 70 Stainless Steel Monel 1.25Cr - 0.5Mo - Si Steel SB-171 Naval Brass From: Chemical Engineering Magazine; Plant Notebook; May 12, 1975 The thickness of heat exchanger tubesheets is an important consideration in cost-estimating and selecting design alternatives for process heat systems. According to the Tubular Exchanger Manufactureres Assn. (TEMA) standards, the tubesheet thickness for shell-and-tube exchangers is given by the formula: TEMA gives precise rules for determining the variables F, G, P, and S for exchanger design. For estimating purposes, however, these terms can be taken as: S P G F T 2 = Page 311 of 327 FileName: 136851736.xls.ms_office WorkSheet: TubeSheet 1 TubePass 2 TubePass 4 TubePass 6 TubePass 8 TubePass 1 TubePass 2 TubePass 4 TubePass 6 TubePass 8 TubePass 1 TubePass 2 TubePass 8 32 26 20 20 21 16 14 10 52 52 40 36 32 32 26 24 16 12 12 81 76 68 68 60 48 45 40 38 36 30 24 13-1/4 97 90 82 76 70 61 56 52 48 44 32 30 15-1/4 137 124 116 108 108 81 76 68 68 64 44 40 17-1/4 177 166 158 150 142 112 112 96 90 82 56 53 19-1/4 224 220 204 192 188 138 132 128 122 116 78 73 21-1/4 277 270 246 240 234 177 166 158 152 148 96 90 23-1/4 341 324 308 302 292 213 208 192 184 184 127 112 25 413 394 370 356 346 260 252 238 226 222 140 135 27 481 460 432 420 408 300 288 278 268 260 166 160 29 553 526 480 468 456 341 326 300 294 286 193 188 31 657 640 600 580 560 406 398 380 368 358 226 220 33 749 718 688 676 648 465 460 432 420 414 258 252 35 845 824 780 766 748 522 518 488 484 472 293 287 37 934 914 886 866 838 596 574 562 544 532 334 322 39 1049 1024 982 968 948 665 644 624 612 600 370 362 Note: where, C = P = L = N = C * (L/P) 2 0.75 (a constant for Square pitch) the tube spacing, in inches the Outer Tube Limit, in inches These tube counts can be taken only as an estimate. For accurate tube counts, an actual scaled layout should be done. Kern does not reveal where he obtained this information and he is not specific in giving details to what TEMA type, orientation, and Outer Tube Limits (OTL) this data applies. Consequently, the user is advised to scrutinize this information before using it. The number of heat exchanger tubes can be estimated from the equation SHELL AND TUBE HEAT EXCHANGER TUBESHEET LAYOUTS (TUBE COUNTS) Source: "Process Heat Transfer"; Donald Q. Kern, McGraw-Hill Book Co. (1950); page 841 Shell I. D. Inches 3/4" O. D. tubes on 1-inch square pitch 1" O. D. tubes on 1-1/4 inch square pitch 1-1/4" O. D. tubes on 1-9/16 inch square pitch Another estimating method for tube counts is found in "Petroleum Refinery Engineering"; Nelson; McGraw-Hill; Page 544: 1.5 inches 13.5 inches 61 Tube Spacing = Outer Tube Limit = Number of Tubes = The OTL is about 1-1/2" less than the inside diameter of the shell in floating head exchangers. It is about 5/8" less than the shell inside diameter of fixed-head or U-tube construction. 4 TubePass 6 TubePass 8 TubePass 1 TubePass 2 TubePass 4 TubePass 6 TubePass 8 TubePass 10 22 16 16 16 16 12 12 30 22 22 22 22 16 16 37 35 31 29 29 25 24 22 51 48 44 39 39 34 32 29 71 64 56 50 48 45 43 39 86 82 78 62 60 57 54 50 106 102 96 78 74 70 66 62 127 123 115 94 90 86 84 78 151 146 140 112 108 102 98 94 178 174 166 131 127 120 116 112 209 202 193 151 146 141 138 131 244 238 226 176 170 164 160 151 275 268 258 202 196 188 182 176 311 304 293 224 220 217 210 202 348 342 336 252 246 267 230 224 1-1/2" O. D. tubes on 1-7/8 inch square pitch These tube counts can be taken only as an estimate. For accurate tube counts, an actual scaled layout should be done. Kern does not reveal where he obtained this information and he is not specific in giving details to what TEMA type, orientation, and Outer Tube Limits (OTL) this data applies. SHELL AND TUBE HEAT EXCHANGER TUBESHEET LAYOUTS (TUBE COUNTS) Source: "Process Heat Transfer"; Donald Q. Kern, McGraw-Hill Book Co. (1950); page 841 1-1/4" O. D. tubes on 1-9/16 inch square pitch 1 TubePass 2 TubePass 4 TubePass 6 TubePass 8 TubePass 1 TubePass 2 TubePass 4 TubePass 8 36 32 26 24 18 37 30 24 10 62 56 47 42 36 61 52 40 12 109 98 86 82 78 92 82 76 13-1/4 127 114 96 90 86 109 106 86 15-1/4 170 160 140 136 128 151 138 122 17-1/4 239 224 194 188 178 203 196 178 19-1/4 301 282 252 244 234 262 250 226 21-1/4 361 342 314 306 290 316 302 278 23-1/4 442 420 386 378 364 384 376 352 25 532 506 468 446 434 470 452 422 27 637 602 550 536 524 559 534 488 29 721 692 640 620 594 630 604 556 31 847 822 766 722 720 745 728 678 33 974 938 878 852 826 856 830 774 35 1102 1068 1004 988 958 970 938 882 37 1240 1200 1144 1104 1072 1074 1044 1012 39 1377 1330 1258 1248 1212 1206 1176 1128 Note: As an example of a discrepancy, refer to the 8" shell with 3/4" tubes on 15/16" triangular pitch and 2-passes. An actual layout yields 48 tubes with 3/16" OTL, as compared with the listed 32 tubes. where, C = P = L = 1.5 inches 13.5 inches 70 Outer Tube Limit = Number of Tubes = 0.86 (a constant for Triangular pitch) the tube spacing, in inches The OTL is about 1-1/2" less than the inside diameter of the shell in floating head exchangers. It is about 5/8" less than the shell inside diameter of fixed-head or U-tube construction. the Outer Tube Limit, in inches Shell I. D. Inches 3/4" O. D. tubes on 15/16-inch triangular pitch 3/4" O. D. tubes on 1-inch triangular pitch Tube Spacing = Another estimating method for tube counts is found in "Petroleum Refinery Engineering"; Nelson; McGraw-Hill; Page 544: The number of heat exchanger tubes can be estimated from the equation N = C * (L/P) 2 Triangular pitch should never be used with a dirty or fouling fluid on the shellside of an exchanger. This configuration is impossible to clean mechanically. SHELL AND TUBE HEAT EXCHANGER TUBESHEET LAYOUTS (TUBE COUNTS) Source: "Process Heat Transfer"; Donald Q. Kern, McGraw-Hill Book Co. (1950); page 842 These tube counts can be taken only as an estimate. For accurate tube counts, an actual scaled layout should be done. Kern does not reveal where he obtained this information and he is not specific in giving details to what TEMA type, orientation, and Outer Tube Limits (OTL) this data applies. Consequently, the user is advised to scrutinize this information before using it. 6 TubePass 8 TubePass 1 TubePass 2 TubePass 4 TubePass 6 TubePass 8 TubePass 1 TubePass 2 TubePass 24 21 16 16 14 36 32 32 26 24 20 18 74 70 55 52 48 46 4 32 30 82 74 68 66 58 54 50 38 36 118 110 91 86 80 74 72 54 51 172 166 131 118 106 104 94 69 66 216 210 163 152 140 136 128 95 91 272 260 199 188 170 164 160 117 112 342 328 241 232 212 212 202 140 136 394 382 294 282 256 252 242 170 164 474 464 349 334 302 296 286 202 196 538 508 397 376 338 334 316 235 228 666 640 472 454 430 424 400 275 270 760 732 538 522 486 470 454 315 305 864 848 608 592 562 546 532 357 348 986 870 674 664 632 614 598 407 390 1100 1078 766 736 700 688 672 449 436 As an example of a discrepancy, refer to the 8" shell with 3/4" tubes on 15/16" triangular pitch and 2-passes. An actual layout yields 48 tubes with 3/16" OTL, as compared with the listed 32 tubes. The OTL is about 1-1/2" less than the inside diameter of the shell in floating head exchangers. It is about 5/8" less than the shell inside diameter of fixed-head or U-tube construction. 3/4" O. D. tubes on 1-inch triangular pitch 1" O. D. tubes on 1-1/4 inch triangular pitch 1-1/4" O. D. tubes on 1-9/16 inch triangular pitch Another estimating method for tube counts is found in "Petroleum Refinery Engineering"; Nelson; McGraw-Hill; Page 544: Triangular pitch should never be used with a dirty or fouling fluid on the shellside of an exchanger. This configuration is impossible to clean mechanically. SHELL AND TUBE HEAT EXCHANGER TUBESHEET LAYOUTS (TUBE COUNTS) Source: "Process Heat Transfer"; Donald Q. Kern, McGraw-Hill Book Co. (1950); page 842 These tube counts can be taken only as an estimate. For accurate tube counts, an actual scaled layout should be done. Kern does not reveal where he obtained this information and he is not specific in giving details to what TEMA type, orientation, and Outer Tube Limits (OTL) this data applies. 4 TubePass 6 TubePass 8 TubePass 1 TubePass 2 TubePass 4 TubePass 6 TubePass 8 TubePass 14 26 22 20 18 14 14 12 12 32 28 26 27 22 18 16 14 45 42 38 36 34 32 30 27 62 58 54 48 44 42 38 36 86 78 69 61 58 55 51 48 105 101 95 76 72 70 66 61 130 123 117 95 91 86 80 76 155 150 140 115 110 105 98 95 185 179 170 136 131 125 118 115 217 212 202 160 154 147 141 136 255 245 235 184 177 172 165 160 297 288 275 215 206 200 190 184 335 327 315 246 238 230 220 215 380 374 357 275 268 260 252 246 425 419 407 307 299 290 284 275 1-1/4" O. D. tubes on 1-9/16 inch triangular pitch 1-1/2" O. D. tubes on 1-7/8 inch triangular pitch Art Montemayor November 03, 1997 TOTAL NUMBER OF TUBES IN AN EXCHANGER, N t: If not known by direct count, find the tube quantity in the tube count table as a function of D otl , the tube pitch, p, and the layout. The shell diameter D i and outer tube limit D otl given in the table are those for a conventional split-ring floating head design, fully tubed out. For a given shell diameter, the value of D otl will be greater than that shown for a fixed tube sheet design and smaller for a pull-through floating head. In any case, the tube count can be reasonably interpolated from the Table using the known or specified D otl , asuming that the tube count is proportional to (D otl ) 2 . All tube count tables are only approximate since the actual number of tubes that can be fitted into a given tubesheet depends upon the pass partition pattern, the thickness of the pass dividers and exactly where the drilling pattern is started relative to the dividers and the outer tube limit. Additional tubes will be lost from the bundle for a U-tube design because the minimum bending radius prevents tubes from being inserted in some, or all, of the possible drilling positions near the centerline of the U-tube pattern. Tubes will also be lost if an impingement plate is inserted underneath the nozzle. For a no-tubes-in-the-window design, the actual number of tubes in the bundle is F c N t . F c is the fraction of total tubes in crossflow. 0.75 0.9375 Triang. 38 32 26 24 0.75 1.0000 Square 32 26 20 20 0.75 1.0000 Triang. 37 30 24 24 1.00 1.2500 Square 21 16 16 14 1.00 1.2500 Triang. 22 18 16 14 0.75 0.9375 Triang. 62 56 47 42 0.75 1.0000 Square 52 52 40 36 0.75 1.0000 Triang. 61 52 48 48 1.00 1.2500 Square 32 32 26 24 1.00 1.2500 Triang. 37 32 28 28 0.75 0.9375 Triang. 109 98 86 82 0.75 1.0000 Square 80 72 68 68 0.75 1.0000 Triang. 90 84 72 70 1.00 1.2500 Square 48 44 40 38 1.00 1.2500 Triang. 57 52 44 42 0.75 0.9375 Triang. 127 114 96 90 0.75 1.0000 Square 95 90 81 77 0.75 1.0000 Triang. 110 101 90 88 1.00 1.2500 Square 60 56 51 46 1.00 1.2500 Triang. 67 63 56 54 0.75 0.9375 Triang. 170 160 140 136 0.75 1.0000 Square 138 132 116 112 0.75 1.0000 Triang. 163 152 136 133 1.00 1.2500 Square 88 82 75 70 1.00 1.2500 Triang. 96 92 86 84 0.75 0.9375 Triang. 239 224 194 188 0.75 1.0000 Square 188 178 168 164 0.75 1.0000 Triang. 211 201 181 176 1.00 1.2500 Square 112 110 102 98 1.00 1.2500 Triang. 130 124 116 110 0.75 0.9375 Triang. 301 282 252 244 0.75 1.0000 Square 236 224 216 208 0.75 1.0000 Triang. 273 256 242 236 1.00 1.2500 Square 148 142 136 129 1.00 1.2500 Triang. 172 162 152 148 0.75 0.9375 Triang. 361 342 314 306 0.75 1.0000 Square 276 264 246 240 19.25 18.00 21.00 19.25 17.25 16.00 12.00 10.75 13.25 12.00 15.25 14.00 8.071 (Sch. 30) 6.82 Tube Layout Number of Tube Passes 4 6 1 10.02 (Sch. 40) 8.77 Shell ID in. Outer Tube Limit Diameter, in. 2 Tube OD in Tube Pitch, in. Page 318 of 327 FileName: 136851736.xls.ms_office WorkSheet: Total Tubes Art Montemayor November 03, 1997 0.75 1.0000 Triang. 318 308 279 269 1.00 1.2500 Square 170 168 157 150 1.00 1.2500 Triang. 199 188 170 164 0.75 0.9375 Triang. 0.75 1.0000 Square 0.75 1.0000 Triang. 1.00 1.2500 Square 1.00 1.2500 Triang. 0.75 0.9375 Triang. 0.75 1.0000 Square 0.75 1.0000 Triang. 1.00 1.2500 Square 1.00 1.2500 Triang. 0.75 0.9375 Triang. 0.75 1.0000 Square 0.75 1.0000 Triang. 1.00 1.2500 Square 1.00 1.2500 Triang. 0.75 0.9375 Triang. 0.75 1.0000 Square 0.75 1.0000 Triang. 1.00 1.2500 Square 1.00 1.2500 Triang. 0.75 0.9375 Triang. 0.75 1.0000 Square 0.75 1.0000 Triang. 1.00 1.2500 Square 1.00 1.2500 Triang. 0.75 0.9375 Triang. 0.75 1.0000 Square 0.75 1.0000 Triang. 1.00 1.2500 Square 1.00 1.2500 Triang. 0.75 0.9375 Triang. 0.75 1.0000 Square 0.75 1.0000 Triang. 1.00 1.2500 Square 1.00 1.2500 Triang. 0.75 0.9375 Triang. 0.75 1.0000 Square 0.75 1.0000 Triang. 1.00 1.2500 Square 1.00 1.2500 Triang. 0.75 0.9375 Triang. 0.75 1.0000 Square 0.75 1.0000 Triang. 1.00 1.2500 Square 1.00 1.2500 Triang. 0.75 0.9375 Triang. 0.75 1.0000 Square 0.75 1.0000 Triang. 1.00 1.2500 Square 1.00 1.2500 Triang. 0.75 0.9375 Triang. 0.75 1.0000 Square 0.75 1.0000 Triang. 1.00 1.2500 Square 1.00 1.2500 Triang. 44.00 42.25 39.00 37.25 42.00 40.25 35.00 33.25 37.00 35.25 31.00 29.25 33.00 31.25 27.00 25.25 29.00 27.25 23.25 21.50 25.00 23.25 21.00 19.25 Page 319 of 327 FileName: 136851736.xls.ms_office WorkSheet: Total Tubes Art Montemayor November 03, 1997 0.75 0.9375 Triang. 0.75 1.0000 Square 0.75 1.0000 Triang. 1.00 1.2500 Square 1.00 1.2500 Triang. 0.75 0.9375 Triang. 0.75 1.0000 Square 0.75 1.0000 Triang. 1.00 1.2500 Square 1.00 1.2500 Triang. 0.75 0.9375 Triang. 0.75 1.0000 Square 0.75 1.0000 Triang. 1.00 1.2500 Square 1.00 1.2500 Triang. 0.75 0.9375 Triang. 0.75 1.0000 Square 0.75 1.0000 Triang. 1.00 1.2500 Square 1.00 1.2500 Triang. 60.00 58.00 48.00 46.00 52.00 50.00 56.00 54.00 Page 320 of 327 FileName: 136851736.xls.ms_office WorkSheet: Total Tubes Art Montemayor November 03, 1997 conventional split-ring floating head design, fully tubed out. For a given shell diameter, the value of D otl will be the tube count is proportional to (D otl ) 2 . All tube count tables are only approximate since the actual number of tubes that can be fitted into a given tubesheet depends upon the pass partition pattern, the thickness of the pass dividers and exactly where the drilling pattern is started relative to the dividers and the outer tube limit. Additional tubes will be lost from the bundle for a U-tube design because the minimum bending radius prevents tubes from will also be lost if an impingement plate is inserted underneath the nozzle. For a no-tubes-in-the-window design, 18 36 60 68 36 40 86 70 74 44 50 128 108 110 64 72 178 142 166 82 94 234 188 210 116 128 290 234 8 Number of Tube Passes Page 321 of 327 FileName: 136851736.xls.ms_office WorkSheet: Total Tubes Art Montemayor November 03, 1997 260 148 160 Page 322 of 327 FileName: 136851736.xls.ms_office WorkSheet: Total Tubes Art Montemayor November 03, 1997 Page 323 of 327 FileName: 136851736.xls.ms_office WorkSheet: Total Tubes Art Montemayor November 03, 1997 Tube OD, in. Tube Pitch, in. Layout P p , in. P n , in. 0.625 0.8125 0.704 0.406 0.750 0.9375 0.814 0.469 0.750 1.0000 1.000 1.000 0.750 1.0000 0.707 0.707 0.750 1.0000 0.866 0.500 1.000 1.2500 1.250 1.250 1.000 1.2500 0.884 0.884 1.000 1.2500 1.082 0.625 Tube Pitch Types: Note: Flow arrows are perpendicular to the baffle cut edge TUBE PITCH PARALLEL TO FLOW, P P , AND NORMAL TO FLOW, P N These quantities are needed only for the purpose of estimating other parameters. If a detailed drawing of the exchanger is available, or if the exchanger itself can be conveniently examined, it is better to obtain these other parameters by direct count or calculation. The quantities are described by Figure 5.2-1 and read from Table IV for the most common tube layouts. 30 o Triangular 60 o Rotated Triangular Flow Flow Square Rotated Square Art Montemayor Heat Exchanger Temperatures August 21, 2004 Rev: 0 Source: Chemical Engineering Magazine; Plant Notebook Section; Unknown date J. T. Petrosky; Vulcan Materical Co. Wichita, Kansas In specifying heat exchanger sevices for process design, it is frequently necessary to arive at optimum condtions through trial and error. However, the determination of each set of condtions within this trial-and-error also involves calculation of interrelated variables, such as inlet and outlet temperatures and area; and this can result in trial-and-error calculations within the trial-and -error for the optimum. It is, thus, convenient to be able to calculate exchanger outlet conditions directly, based on known or assumed values of inlet temperatures, specific heats, flowing quantities, overall transfer rate, and surface. Such a direct calculation is developed as follows and shown in the sketch. Nomenclature: q = Heat duty, Btu/hr or kcal/hr = 1,000,000 C p = Constant or average specific heat on the shell side, Btu/lb or kcal/kg = 0.5000 c p = Constant or average specific heat on the tube side, Btu/lb or kcal/kg = 1.0000 W = Fluid mass flow rate in shell side, lb/hr or kg/hr = 100,000 w = Fluid mass flow rate in tube side, lb/hr or kg/hr = 45,000 U = Overall heat transfer coefficient, Btu/hr-ft 2 - o F or kcal/hr-m 2 - o C = 125 A = Total exchanger heat transfer area, ft 2 or m 2 = 300.0 T 1 = Shell-side fluid temperature, o F or o C = 250 t 1 = Tube-side fluid temperature, o F or o C = 85 AT m = Log mean temperature difference, o F or o C= 1 = Subscript denoting inlet conditions 2 = Subscript denoting outlet conditions From the derived equations, let: Z = 50,000 B = 45,000 C = 0.920044 Therefore, T 2 = 152 o F or o C Direct Calculation of Exchanger Exit Temperatures B t 1 (1 - C) - T 1 (B - Z)/(Z - BC) = q, U, A, AT m w, c p , t 1 W, C p , T 2 W, C p , T 1 w, c p , t 2 | . | \ | ÷ = B Z UA e C 1 1 Page 325 of 327 FileName: 136851736.xls.ms_office Worksheet: Ht Exchanger Temperatures Art Montemayor Heat Exchanger Temperatures August 21, 2004 Rev: 0 Equations and their derivations: The heat transferred to the tube-side fluid = q = (w) (c p ) (t 2 - t 1 ) The heat transferred to the shell-side fluid = q = (W) (C p ) (T 1 -T 2 ) In specifying heat exchanger sevices for process design, it is frequently necessary to arive at optimum condtions through trial and error. However, the determination of each set of condtions within this trial-and-error also involves Let: B = (w) (c p ) trial-and-error calculations within the trial-and -error for the optimum. It is, thus, convenient to be able to calculate Z = (W) (C p ) exchanger outlet conditions directly, based on known or assumed values of inlet temperatures, specific heats, flowing quantities, overall transfer rate, and surface. Such a direct calculation is developed as follows and shown Combining both above equations, ( ) ( ) ( ) ( ) ( ( ( ( ¸ ( ¸ ÷ ÷ ÷ ÷ ÷ = A = 1 2 2 1 1 2 2 1 ln t T t T t T t T UA T UA also is d transferre heat The m ( ) 1 2 1 2 t T T B Z t + ÷ | . | \ | = ( ) ( ) ( ) ( ) ( ) ( ( ( ( ¸ ( ¸ ÷ ÷ ÷ ÷ ÷ = ÷ 1 2 2 1 1 2 2 1 2 1 ln t T t T t T t T UA T T Z ( ) ( ) ( ) ( ) ( ) ( ( ( ( ¸ ( ¸ ÷ ÷ ÷ + ÷ | . | \ | ÷ = ( ( ( ( ¸ ( ¸ ÷ ÷ ÷ | . | \ | ÷ 2 1 1 2 1 2 1 1 1 2 1 2 1 1 ln T T Z t T t T T B Z T UA t T t T T B Z T ( ) ( ) ( ( ( ( ¸ ( ¸ ÷ + ÷ + | . | \ | + | . | \ | ÷ = ( ( ( ( ¸ ( ¸ ÷ ÷ | . | \ | + | . | \ | ÷ 2 1 1 2 1 2 1 1 1 2 1 2 1 1 ln T T Z t T t B Z T B Z T T UA t T t B Z T B Z T T ( ) ( ) ( ) ( ) ( ) ( ) 2 1 2 1 2 1 2 1 2 2 1 1 1 2 1 2 1 1 ln T T Z T T B Z T T UA T T Z T T T B Z T UA t T t B Z T B Z T ÷ ÷ | . | \ | ÷ ÷ = ( ( ( ( ¸ ( ¸ ÷ ÷ ÷ | . | \ | ÷ = ( ( ( ( ¸ ( ¸ ÷ ÷ | . | \ | + | . | \ | ÷ ( ) | . | \ | ÷ = ( ( ( ( ¸ ( ¸ | . | \ | ÷ = ( ( ( ( ¸ ( ¸ ÷ ÷ | . | \ | + | . | \ | ÷ B Z UA Z B Z UA t T t B Z T B Z T 1 1 1 1 ln 1 2 1 2 1 Page 326 of 327 FileName: 136851736.xls.ms_office Worksheet: Ht Exchanger Temperatures Art Montemayor Heat Exchanger Temperatures August 21, 2004 Rev: 0 ( ) ( ) ( ) ( ) ( ( ( ( ¸ ( ¸ ÷ ÷ ÷ ÷ ÷ = A = 1 2 2 1 1 2 2 1 ln t T t T t T t T UA T UA also is d transferre heat The m ( ) ( ) ( ) ( ) ( ) ( ( ( ( ¸ ( ¸ ÷ ÷ ÷ + ÷ | . | \ | ÷ = ( ( ( ( ¸ ( ¸ ÷ ÷ ÷ | . | \ | ÷ 2 1 1 2 1 2 1 1 1 2 1 2 1 1 ln T T Z t T t T T B Z T UA t T t T T B Z T ( ) ( ) ( ( ( ( ¸ ( ¸ ÷ + ÷ + | . | \ | + | . | \ | ÷ = ( ( ( ( ¸ ( ¸ ÷ ÷ | . | \ | + | . | \ | ÷ 2 1 1 2 1 2 1 1 1 2 1 2 1 1 ln T T Z t T t B Z T B Z T T UA t T t B Z T B Z T T ( ) ( ) ( ) ( ) ( ) ( ) 2 1 2 1 2 1 2 1 2 2 1 1 1 2 1 2 1 1 ln T T Z T T B Z T T UA T T Z T T T B Z T UA t T t B Z T B Z T ÷ ÷ | . | \ | ÷ ÷ = ( ( ( ( ¸ ( ¸ ÷ ÷ ÷ | . | \ | ÷ = ( ( ( ( ¸ ( ¸ ÷ ÷ | . | \ | + | . | \ | ÷ ( ) | . | \ | ÷ = ( ( ( ( ¸ ( ¸ | . | \ | ÷ = ( ( ( ( ¸ ( ¸ ÷ ÷ | . | \ | + | . | \ | ÷ B Z UA Z B Z UA t T t B Z T B Z T 1 1 1 1 ln 1 2 1 2 1 Page 327 of 327 FileName: 136851736.xls.ms_office Worksheet: Ht Exchanger Temperatures