2Heat Exchangers Introduction ............................................... 28 Shell Side Pressure Drop ............................... 46 TEMA ...................................................... 28 Heat Transfer Coefficients ............................. 47 Selection Guides .......................................... 33 Fouling Resistances ...................................... 47 Design Recommendations .............................. 35 Installation Recommendations ........................ 48 Process Data .............................................. 37 Thermal Conductivity of Metals ...................... 49 Heat Exchanger Configuration and Area ........... 38 Vacuum Condensers .................................... 50 Determining the LMTD Configuration Air-cooled Heat Exchangers: Forced vs. Correction Factor ........................................ 39 Induced Draft ............................................. 51 Tubeside Pressure Drop ................................ 40 Air-cooled Heat Exchangers: Air Data .............. 52 Tube Side Film Coefficient ............................. 40 Air-cooled Heat Exchangers: Thermal Design ..... 52 Air-cooled Heat Exchangers: Pressure Drop, Shell Diameter ............................................ 41 Air Side .................................................... 55 Ideal Shell Side Film Coefficient ...................... 42 Air-cooled Heat Exchangers: Temperature Shell Side Film Coefficient Correction Factors .... 43 Control ..................................................... 55 Overall Heat Transfer Coefficient .................... 45 Rules of Thumb for Chemical Engineers. 27 DOI: 10.1016/B978-0-12-387785-7.00002-5 Copyright Ó 2012 Elsevier Inc. All rights reserved. 28 Rules of Thumb for Chemical Engineers Introduction Heat exchangers are critical elements in every process parameters and application information, proper plant. While the majority of exchangers are the shell-and- sizing and selection of heat exchangers is impossible, tube type, there are several additional important types. The and all aspects of performance will be compromised. major types of heat transfer equipment are: • Codes and design specifications. Specifying a TEMA designation and an ASME pressure and temperature • Shell-and-tube requirement will enhance all heat transfer selections. • Finned tube • Installation. Following appropriate installation • Bare tube recommendations can eliminate most premature • Plate-and-frame failures and greatly enhance the performance and • Spiral efficiency of the heat transfer unit. • Plate coil • Evaluation. Always evaluate the selections in terms This chapter focuses on shell-and-tube exchangers, of a ten-year operational period, considering all covering topics of interest to typical process engineers. factors. Plate-and-frame and spiral exchangers are also discussed. An Excel workbook accompanies this chapter. The Four factors impact the performance, longevity, and workbook performs calculations for a liquid-liquid shell- maintenance requirements for heat-transfer equipment and and-tube heat exchanger and completes the associated related components [22]: TEMA datasheet. • Initial knowledge and documentation of all the operating parameters. Without correct operating TEMA Describe shell-and-tube heat exchangers using used. Enter the TEMA designation (e.g., BEM) into the nomenclature from the Standards of the Tubular cell labeled “Type” on line 6. Enter the TEMA Class (e.g., Exchanger Manufacturers Association (TEMA). Figure R) on line 54. 2-1 illustrates the front head, shell, and rear head types and The process engineer usually works closely with the lists letter designations corresponding to each. Figure 2-2 exchanger manufacturer to complete the datasheet. Heat shows six typical heat exchanger configurations, with exchanger design is often a trial-and-error process, with their corresponding TEMA designation (e.g., BEM). The different combinations of shell diameter, tube size, length, various parts of the exchangers are called out with the key tube passes, and other attributes being tested. All heat to the parts listed in Table 2-1. exchanger manufacturers use sophisticated software for In addition to the exchanger configurations, TEMA thermal and mechanical design, and they are usually more provides design and construction standards for three major than happy to assist customers by running multiple design classes of exchanger, called R, C, and B. Table 2-2 cases. compares attributes of the three exchanger classes. The Although computers solve the design equations for three classes are listed in order of decreasing cost (and most new exchangers, engineers may want to do some mechanical performance). preliminary work using the manual methods as described Use datasheets to tabulate the primary process and later in this chapter. Sophisticated software such as the mechanical requirements for a heat exchanger. TEMA HTRI Xchanger Suite [11] performs rigorous incremental datasheets are recommended because they are well known calculations that account for the highly dynamic nature of by engineers and fabricators. Versions with SI and US heat exchangers. The manual calculation methods use units are given in Figure 2-3 and Figure 2-4. Similar physical properties averaged across the exchanger, and datasheets from other sources, such as heat exchanger provide heat transfer and pressure drop approximations manufacturers and engineering companies, may also be for various zones within the exchanger. Heat Exchangers 29 Figure 2-1. Nomenclature for shell-and-tube heat exchangers [24]. .30 Rules of Thumb for Chemical Engineers Figure 2-2. Typical TEMA heat exchangers [24]. c) all and up and up hydrocarbons Peripheral gasket contact surface Flatness tolerance specified No tolerance specified No tolerance specified Minimum tubesheet thickness with Outside diameter of the tube 0. OD ¾. Stationary Head e Bonnet 21. Stationary Head Nozzle 24. Slop-on Backing Flange 2.6 mm) 0. Lantern Ring 8. Vent Connection 14. 1¼. Tie Rods and Spacers 9. Floating Head Cover 35. and ⅝ in. ¼ inch carbon steel ⅛ inch alloy. Shell Flange e Stationary Head End 29.0625 in (1.75 x tube OK for 1 inch and 0. Packing Follower Ring 7. Transverse Baffles or Support Plates 10. Shell 27. tabulated 6 inch tabulated Longitudinal baffle thickness ¼ inch minimum ⅛ inch alloy. Tubes 26.2 x R þ lane may be 3/ 16 inch in 12 inch lane ¼ inch lane tube OD and smaller shells for ⅝ and ¾ in tubes Minimum shell diameter 8 inch.25 inch for 2 OD 1. Support Bracket 19. 75 psi for 43 to 60 in. Split Shear Ring 38. Drain Connection 15. internal floating head. C.75 x tube OK for 1 inch and expanded tube joints smaller smaller ⅞ inch for 1¼ OD ⅞ inch for 1¼ OD 1 inch for 1½ OD 1 inch for 1½ OD 1. Heat Exchangers 31 Table 2-1 Heat exchanger parts and connections (for Figure 2-2) 1. Gasket materials Metal jacketed or solid metal for Metal jacketed or solid metal for a) Metal jacketed or solid metal for a) a) internal floating head cover. Floating Tubesheet Skirt 4. Impingement Baffle 11. Instrument Connection 16. Lifting Lug 18. ¼ inch carbon steel Floating head cover cross-over 1. Shell Cover 28. Stationary Head Flange e Channel or Bonnet 22. Support Saddle 17. (3. Shell Flange e Rear Head End 30. Floating Head Cover e External 3. 1½. Shell Nozzle 31. b) 300 psi b) 300 psi and up.25 inch for 2 OD Tube hole grooving Two grooves Above 300 psi design pressure or Two grooves 350 F design temperature: 2 grooves Length of expansion Smaller of 2 inch or tubesheet Small of 2 x tube OD or 2 inch Smaller of 2 inch or tubesheet thickness thickness Tubesheet pass partition grooves 3 / 16 inch deep grooves required Over 300 psi: 3/ 16 inch deep Over 300 psi: 3/ 16 inch deep grooves required or other grooves required or other (Continued) . ⅜. Tube pitch and minimum cleaning 1. and 2 in. Longitudinal Baffle 12. Stationary Head e Channel 20.125 in.2 mm) 0. 150 psi for 25 to 42 in. Floating Head Backing Device 37. tabulated 6 inch. and B heat exchangers.6 mm) steel Tube diameters. Packing Box Flange 5. R þ ⅝ in. Liquid Level Connection Table 2-2 Comparison of TEMA class R. Shell Cover Flange 32. Cost decreases from left to right [23] Attribute Class R Class C Class B Application Generally severe requirements Generally moderate requirements General process service such as petroleum and related such as commercial and general processing applications process applications Corrosion allowance on carbon 0. Weir 39.0625 in (1. Floating Head Flange 36. 1. Expansion Joint 33. Pass Partition 13.25 x tube OD R þ ⅜ tubes may be located 1. Floating Tubesheet 34. R þ ¼. Packing 6. Channel Cover 23. b) 300 psi internal floating head.3 x tube flow area Same as tube flow area Same as tube flow area area Lantern ring construction 375 F maximum 600 psi maximum 375 F maximum 300 psi up to 24 inch diameter 300 psi up to 24 inch diameter shell shell 150 psi for 25 to 42 in. 75 psi for 43 to 60 in. ½. Stationary Tubesheet 25. .32 Rules of Thumb for Chemical Engineers Table 2-2 Comparison of TEMA class R. smaller ⅝ inch bolting may be used Figure 2-3. SI units [24]. Data Sheet for shell-and-tube heat exchanger. and B heat exchangers. C. Cost decreases from left to right [23]dcont’d Attribute Class R Class C Class B suitable means for retaining suitable means for retaining gaskets in place gaskets in place Pipe tap connections 6000 psi coupling with bar stock 3000 psi coupling 3000 psi coupling with bar stock plug plug Pressure gage connections Required in nozzles 2 inch and up Specified by purchaser Required in nozzles 2 inch and up Thermometer connections Required in nozzles 4 inch and up Specified by purchaser Required in nozzles 4 inch and up Nozzle construction No reference to flanges No reference to flanges All nozzles larger than one inch must be flanged Minimum bolt size ¾ inch ½ inch recommended. g. US units [24].. Data Sheet for shell-and-tube heat exchanger. thermal duty. choosing the type of heat exchanger to use for a particular phase change). Heat Exchangers 33 Figure 2-4. temperature approach application: • Cleanliness of the streams . Selection Guides The following factors should be considered when • Operating conditions: service requirements (e. 000 psig). reboiling. but detailed design is normally left designs.-psig). mechanically Yes. and gas cooling applications. especially when made of carbon steel. mechanically mechanically. mechanically Yes. The most efficient design is achieved when the hot and cold fluid flow rates are approximately the same. easier maintenance. and can handle certain fouling surfaces and gaskets fluids. outlet temperature) to a shell-and-tube exchanger that is spec- This is the most common type of heat exchanger used in ified for the same service where the engineer specifies the chemical process industries. The plate-and-frame units have • Heating or cooling application higher heat transfer coefficients – often three to four • Maintenance requirements times that of a shell-and-tube exchanger.e.). and product of shell diameter times pressure not exceeding 315. resulting in similar velocities on both sides of the plates.000 in. mechanically or Chemically only Yes. and are usually used specialized and often proprietary. Manufacturers provide for liquid-liquid.34 Rules of Thumb for Chemical Engineers • Maximum design pressure and temperature shell-and-tube designs. They are • Material compatibility with process fluids: wetted compact. These are true counter- In appropriate circumstances. reduced fouling. the most common. Table 2-3 Shell-and-tube exchanger selection guide (cost increases from left to right) [1] Floating Head Floating Head Floating Head Split Pull-Through Type of Design “U” Tube Fixed Tubesheet Outside Packed Backing Ring Bundle Provision for differential Individual tubes Expansion joint Floating head Floating head Floating head expansion free to expand in shell Removable bundle Yes No Yes Yes Yes Replacement bundle Yes Not practical Yes Yes Yes possible Individual tubes Only those in Yes Yes Yes Yes replaceable outside row Tube interiors cleanable Difficult to do Yes. cost effective. some curves and software for use by end users (for TEMA Class exchangers are used for most custom example. Increased turbulent heat transfer. TEMA guidelines are limited to a shell diameter of 1524 mm (60 in. see Ref [10]). or chemically or chemically or chemically or chemically can do chemically Tube exteriors with Chemically only Chemically only Chemically only Chemically only Chemically only triangular pitch cleanable Tube exteriors with square Yes. Shell-and-Tube Heat Exchangers This may require different process parameters (i. and smaller size characterize the Plate-and-Frame Heat Exchangers performance of spiral heat exchangers when compared with shell-and-tube exchangers. Moretta has summarized the design calcu- exchangers offer many advantages compared with lations for heat transfer and pressure drop [17]. shelf models are available in fixed tubesheet and U-tube The design of plate-and-frame exchangers is highly design configurations in smaller sizes.000 mm-bar (60. Off-the- coefficient. with TEMA B (chemical industry service) being to the manufacturers. It is often the lowest cost a high shellside flow rate to maximize the shellside film option. mechanically Yes. mechanically Yes.. working pressure of 207 bar Spiral Heat Exchangers (3. plate-and-frame heat current units. mechanically pitch cleanable chemically or chemically or chemically or chemically Number of tube passes Any practical even Normally no Normally no Normally no Normally no number possible limitations limitations limitations limitations Internal gaskets eliminated Yes Yes Yes No No . mechanically Yes. 650 ft2) 4310 m2 (46.32 to 23.5 to 1. fatigue characteristics of the metal plate may be limiting if plate-and-frame temperature or pressure cycling is a process characteristic Other characteristics are similar to the gasketed plate-and-frame exchangers Spiral Up to 500 C (930 F) and 25 bar (360 psig). engineers conditions (including start-up and turndown scenarios). viscous liquids.7 ft2) area per plate 1. training are recommended.400 ft2) Gasketed plate-and-frame Up to 180 C (350 F) and 20 bar (300 psig). fatigue characteristics of the metal plate may be limiting if temperature or pressure cycling is a process characteristic Up to 2800 m2 (30. so hot and cold fluids should have roughly equivalent flow rates Significant size reduction and weight savings compared with shell-and-tube Gasketed exchangers may be unsuitable for use in highly aggressive media or when leakage is not tolerable Welded.000 ft2) heat transfer area in a single unit Typically designed with 70 kPa to 100 kPa (10 to 15 psi) pressure drop Maximum flow 2500 m3/h (11.02 to 0. After collecting and tabulating • Which of the following parameters can float? To thermodynamic properties for the major fluid components.) thick 0.. tantalum) is required. 1380 bar (20. although “deep groove” or “wide gap” plate designs can tolerate up to 18 mm particles [14]. Where process fluids undergo a change in the process engineer’s responsibility to address potential state (condensers and boilers). at least one parameter is create heat and material balances for normal operating determined from the other five: hot and cold stream .05 in. or fusion-sealed Up to 450 C (850 F) and 40 bar (600 psig). close the heat balance.) Length 9m (30 ft) horizontal 12 m (40 ft) 25 m (75 ft) vertical Area 1270 m2 (13. or when enhanced energy recovery is important High turbulence High heat transfer coefficients High fouling resistance Not available in carbon steel Hot and cold side channels have nearly identical geometry.000 ft2) heat transfer area Typical maximum sizes Floating Head Fixed Head or U-Tube Diameter 1524 mm (60 in.3 ft/s) Plates 0.1 m/s (0.03 to 2. so little or no insulation is required Consider when a high-grade.) 2000 mm (80 in.5 mm). limits vary depending on size and material of construction Up to 350 m3/h (1500 gpm). 310 bar (4.000 gpm) Minimum velocity 0.2 m2 (0. brazed. the design calculations are performance differences among alternative design much more complex.2 mm (0.g. when space is tight. Here are guidance questions for the process engineer: Process engineers should start with a full understanding of the duty requirements.2 in.06 to 0. especially. Heat Exchangers 35 Table 2-4 Compact heat exchanger attributes Exchanger Type Attributes Shell-and-tube Up to 650 C (1200 F). Usually only used for liquid-liquid service. limited due to single channel 0.500 psig) in the shell.000 psig) in the tubes Up to 4650 m2 (50. expensive construction material (e. and specialized software and solutions. can easily model liquid-liquid shell-and-tube heat There may be design trade-off decisions and it is usually exchangers. Operates efficiently with crossing temperatures and close approach temperatures Only the plate edges are exposed to atmosphere. and liquids with solids in suspension Design Recommendations For conceptual and preliminary design work.) spacing between plates Typically used in clean service (no particles larger than 2. sedimentation (compared with shell-and-tube) Particularly effective in handling sludges.5 to 500 m2 (5 to 5400 ft2) heat transfer area in one spiral body Countercurrent design allows for very deep temperature cross and close approach High turbulence reduces fouling and.0 mm (0.5 to 5. heat exchanger. dependent. between shutdowns. what evaluated at the average temperature for each stream are would happen as the temperature of the stagnant fluid fine. and elevation results in an acceptable pressure drop.36 Rules of Thumb for Chemical Engineers inlet temperature. alcohols) and non. However.g. computer programs they have at their disposal. a reduction in duty due to process varia- • What variation in temperature of the fluids is tions. in the heat exchanger cools? Calculate the total duty for the exchanger in Watts. . and flow rate. rithmic average of the components’ mass-weighted At this point the top part of the datasheet can be viscosities (see Equation 27-3 in Chapter 27). is a major contributor to the heat transfer • Are there conditions that could result in freezing. tubes to achieve the surface area based on the assumed • Are there physical limitations? Consider the available overall heat transfer coefficient. The calculations can temperature of the hot fluid is lower than the outlet be solved with spreadsheets to provide a platform for temperature of the cold fluid. requirements (the relationship with associated Pick a shell type based on the process requirements. Then use the tabulated • Are thermodynamic properties for the hot and cold “typical” heat transfer coefficients to compute streams available. is usually more important for The answer is often flexible. • What are the maximum allowable pressure drops Evaluate the design problem using physical properties through the equipment for the two streams? Be sure appropriate to the temperature of the fluids. • Is this a batch or continuous process? Operating Estimate the pressure drop through the shell using the efficiency. Assume a tube diameter instances. properties stopped while continuing the cold stream flow. if a liquid-liquid compact heat • How will the fluid flow rates be controlled? If it’s exchanger is anticipated. Determine its diameter by the tube layout and passes. or years. equipment such as columns and pumps). mechanical design of the exchanger and will be deter- for instance. or • Similarly. mined later. Add a safety factor of 10% which includes stream flow stops without interrupting the hot fouling and uncertainty (or another factor depending on stream? the specific design problem). or can they be predicted from the the required heat transfer area.. For preliminary design work. It is physically possible evaluating alternatives or rating existing exchangers in true counter-current equipment such as a spiral without involving vendors or consultants. and Pick either the hot or cold fluid to flow inside the tubes a single-pass type BEM shell-and-tube unit. in terms such as pumping cost and method given in this chapter. outlet temperature. Then manipulate the space for installation (including logistics of rigging exchanger length and number of tube passes. ranges. what outcome is expected if the cold Btu/h. outside dew point temperature. In many (for a shell-and-tube exchanger). coefficient. the mixed liquid viscosity is a loga. water. example. This is espe- that unintended vaporization would not occur as cially important for viscosity which is highly temperature pressure is reduced. Newtonian emulsions and slurries. For example. for streams should be within about 20% of each other. to use shell-and-tube equipment. The properties of the pure components? There are many actual required heat transfer area depends on the miscible liquids that behave rationally when mixed. a double-pipe exchanger. or a lower than planned cooling water flow rate expected? This is particularly pertinent for cooling due to oversizing the exchanger may result in tower water that has a temperature that varies with the excessive fouling. the process engineer may also (or instead) use the • Is a temperature cross expected and if so can it be approximate methods and procedure given below to come avoided? A temperature cross occurs when the outlet up with a reasonable design solution. meaning that two or exchangers that are in continuous operation for three of the parameters may be safely varied within months. maintenance. completed and sent to a vendor or heat exchanger engineer However. multiple (usually start with ¾ inch) and calculate the total length of shells are required. the flow rate of the two planned to control the flow rate of cooling water. other mixtures deviate widely such as to design an exchanger using one of the sophisticated polar liquids (e. and plays a central role in pressure drop precipitation. calculating the exchanger into place). or fouling? If the hot stream flow is calculations. maintenance (with an the pressure drop through the tubes until a combination allocation for removing tubes). This is conceptual. and thermal DH ¼ W Cp ðtout tin Þ (2-1) conductivity. It also has point values for molecular weight. The worksheet called “Fluid Data” The change in enthalpy for each stream is evaluated tabulates temperature-correlated coefficients for vapor using the equation: pressure. DH ¼ enthalpy change. heat of vaporization. kg/h or lb/h values even though at least one of these must be adjusted Cp ¼ specific heat. specific heat. many design solutions are the fluids between tube and shell side. What is “reasonable?” There’s no one include: “correct” answer which is why experience and expertise • Heat transfer rate (“U”) are important characteristics for the designer. There are input cells for all six flow and temperature W ¼ mass flow rate. . typically. F to satisfy the heat balance. There are also inputs for tout ¼ temperature at exchanger outlet. through the design steps for a shell-and-tube exchanger in See Figure 2-5. C or F Figure 2-5. the radio buttons identify the unknown. C or Btu/lb. One of the values is calculated based on the other five to close the heat balance. reliability. and fouling resistance. allowable pressure drop. safety. pressure drop. Heat Exchangers 37 Iterate the preceding two steps using different The proper selection of a heat exchanger depends on assumptions (e. Compare with the original assumption • Materials of construction and iterate. Where: The fundamental process parameters – flow and temperature – are entered on the “Process Data” work. of the assumption. using the newly computed coefficient in place • Miscellaneous factors such as leak-tightness. density. viscosity. • Cost (operating and maintenance over the expected Calculate heat transfer film coefficients for the tube and life of the exchanger or 10 years) shell side and combine with the tube resistance and • Pumping power assumed fouling factors to compute an overall heat • Size and weight transfer coefficient. Fundamental process data includes flow and temperature information for the hot and cold streams..) to find compared before a final design is accepted. liquid-liquid service. swapping interrelated factors. through the design steps if necessary. kJ/h or Btu/h sheet. tube diameter. Factors a reasonable design.g. etc. and noise Process Data The Excel spreadsheet accompanying this chapter steps pressure. kJ/kg. C or F tin ¼ temperature at exchanger inlet. and flash point. W or Btu/h to reduce cost. temperatures are changed during the design procedure. C or Btu/ (sometimes just faced). return to this step and update the assumed overall heat transfer coefficient to equal that which was deter. F need to be made of the corrosion-resistant alloy. Other designs use the following formula for LMTD and mined by the procedure. Select a tube size DTmean ðcocurrentÞ ¼ ðTin tin Þ (Table 2-2). The two results (for hot and cold streams) are added in The stream properties are evaluated as follows. Rules of thumb to help decide include: tube diameter.000 W properties at the inlet and outlet of the exchanger. put it in the shell. put it inside the tubes Q ¼ heat transferred. tubesheets U ¼ overall heat transfer coefficient. weighted by the mass fraction of the components (see Equation 27-3 in Chapter 27).” When the heat balance is Density. HeatBalance has at the inlet and outlet temperatures for each stream. DH for the hot side is a negative value and it is evaluated for each component of the hot and cold streams positive for the cold side. length ln ðTout tout Þ (typically 4 ft. In this example the cold stream the stream then summed. and mean temperature difference: the correction factor is less than about 0. capacity variable with temperature. Then only the tubes.38 Rules of Thumb for Chemical Engineers Note that the specific heat is equal to the average of the it is easy to implement and allows for changing of the heat values at inlet and outlet temperatures. GoalSeek is used by the spreadsheet because is taken. C or F other.80 then consider adding shells to achieve a result that is closer to counter- Q A ¼ (2-2) current design. (2-3) After completing all of the calculations in the following sections. number of tube and shell passes. and thermal conductivity are satisfied. Heat Exchanger Configuration and Area Pick either the hot or cold stream to flow through the A ¼ heat transfer area. put it inside the tubes.000 Btu/h) are transferred. and piping h-ft2. or 20 ft). and 79. as the (270. • If one fluid is much more severely fouling than the Determine the mean temperature difference (MTD) by other place it in the tubes. specific heat. especially when mechanical means such then applying a correction factor that is based on the as brushes are used. usually calculated at the outside tubes. tube channels. 15 ft. DTmean ¼ mean temperature difference (MTD) between • If one fluid is at a much higher pressure than the hot and cold streams. overall countercurrent exchanger (shell passes ¼ tube passes). m2 or ft2 • If one fluid is highly corrosive. 12 ft. 8 ft. F ¼ 1. This gives an estimate for the temperature is found to be 10 C (50 F). there is no correction drop. For a strict cocurrent flow • If one fluid has a very limited allowable pressure design (single pass shell and tube). Tubes are easier to clean calculating the log-mean temperature difference (LMTD) than shells. heat capacity is assumed to be constant (which is a good Viscosity is also tabulated and the logarithmic average assumption). factor and this equation applies: Characterize the tube side by assuming an overall heat transfer coefficient (see Table 2-8 on page 47) and a safety ðTin tin Þ ðTout tout Þ factor primarily to account for fouling. They a zero value. a cell named “HeatBalance. wall thickness (start with 14 BWG). W/m2. The heat balance can also be solved algebraically if the the properties are immediately updated. and number of passes (either 1-pass or an even number up to about 14). different shell and tube configurations. U DTmean .0 for a true The heat transfer area is related to the heat duty. Iterate until the calculated overall a correction factor read from graphs corresponding to coefficient equals the assumed one. and Excel’s GoalSeek function is used to find are multiplied by the mass fraction of the component in the unknown variable. Therefore. If heat transfer coefficient. heat transfer area. 12. Determine the minimum number of tubes by dividing sionless (see next section) the total length. Bowman compiled formulae that F ¼ W’ (2-7) W ’ 1 accurately represent the graphs for every configuration of ’ þ pffiffiffi shell-and-tube exchanger system [4]. U. etc. Heat Exchangers 39 t ¼ inlet and outlet temperatures of the cold stream. is constant R 1 throughout the heat exchanger 1 P R 1=N • The rate of flow of each fluid is constant W ¼ • The specific heat of each fluid is constant 1 P • There is no condensation of vapor or boiling of liquid Tin Tout R ¼ in a part of the exchanger tout tin • Heat losses are negligible tout tin • There is equal heat transfer surface area in each pass P ¼ Tin tin • The temperature of the shell-side fluid in any shell- side pass is uniform over any cross section For the special case when R ¼ 1 (and the logarithms • There is no leakage of fluid or heat across the cannot be evaluated): transverse baffle separating two shell passes . and ¼ F ðTin tout Þ safety factor. 8. This Excel formula gives the answer: ntubes ¼ ROUNDðLengthOfAllTubes=ðTubeLength TubePassesÞÞ þ 0:5. C to the next integer that is evenly divisible by the number of or F tube passes. dimen. by tube length and rounding up T¼ inlet and outlet temperatures of the hot stream. C DTmean ðcountercurrentÞ or F ðTin tout Þ ðTout tin Þ (2-4) From the tube outside diameter. tube passes) [6]. 0Þ TubePasses Determining the LMTD Configuration Correction Factor Many references present F factors in graphical form pffiffiffi 1 W ’ 2 (for example: Perry’s). with 2 shell passes there may be any multiple of 2N tube Where: passes or 4. LAlltubes. N NP S ln W W’ ¼ F ¼ (2-6) N NP þ P 1 þ W S þ SW ln 1 þ W þ S SW And: DTmean ¼ F ðTout tin Þ Where: pffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi Assumptions for the F factor equations and charts are: R2 þ 1 S ¼ • The overall heat transfer coefficient. calculate the total tube length: ln ðTout tin Þ A Fsafety LAll tubes ¼ (2-5) p do F¼ LMTD configuration correction factor. Fakheri then 1 W 2 ln ’ collapsed the correlations into a single algebraic equation W 1 that is applicable to shell-and-tube heat exchangers with N pffiffiffi 1 W’ 2 shell passes and 2NM tube passes per shell (for example. Tw ¼ t þ ðT tÞ (2-9) hi cp m NPr ¼ Prandtl Number ¼ k Where: m ¼ viscosity. with [15]: Adjust physical parameters (tube size. shell side fluid.000 the tube pathway (e. transitional. Use the Hausen correlation for laminar flow terms of the surface area inside the tubes. Calculate the mean wall temperature. C or Btu/ft2. pass then L is the total length of tubing divided Ui by 10). Pa or psf (divide by 1. expressed in 2. A good average fluid temperature and L is the length for initial guess for the film coefficient is about 2. 144 for psi) 2-10. so use an assumed value k 0:0668 NRe NPr ðdi =LÞ m 0:14 for both then iterate through all of the calculations 3:66 þ di 1 þ 0:40 ½NRe NPr ðdi =LÞ2=3 mw until the assumed values match the calculated ones.000) [2] . C or Btu/ft2. if there are 10 tubes per W/m2- C or 400 Btu/ft2- F. exchanger length. estimate the pressure drop for gc ¼ conversion factor. gc iterate until a “reasonable” configuration is attained.40 Rules of Thumb for Chemical Engineers Tubeside Pressure Drop Calculate the pressure drop in two parts then add Where: together: DPt ¼ pressure drop through turns. m/s or ft/s tube passes. Use the Sieder Tate equation for turbulent flow t ¼ average temperature.17 ft/s2 turning the flow through the heads or channels Compare the calculated and allowable pressure drops. W/m2. then evaluate the viscosity at that temperature. 1 m/s2 or 32. C or F 3.. Neither of these values is hi ¼ known until the calculations for both the tube side " # and shell side are complete. and 2-11 to compute pressure drop through the np ¼ number of passes tubes. Use Ui ¼ overall heat transfer coefficient based on inside the correlation that corresponds to the flow regime area. use equations 2-9. The overall coefficient was already assumed to (2-10) estimate the heat transfer area. C or F properties evaluated at the average fluid temperature. From the velocity in the tubes and number of u ¼ velocity in tubes. The “Tube Pressure Drop” and “F Factor” worksheets do the calculations just described. ¼ Uo di 1. W/m2. mPa-s or lbm/ft-h T w ¼ average inside wall temperature. or turbulent) for the tube side film do coefficient. The formula uses hi ¼ inside film coefficient.g. r ¼ density. F the overall heat transfer coefficient. and pressure drop. it was based on the Where the properties are evaluated at the outside area of the tubes (see page 38). F (laminar. and number of tube passes) and repeat the calculations for 2 ðnp 1Þ r u2 DPt ¼ (2-8) heat exchanger area. C or F (Reynolds number >¼ 10. total tube length. Tube Side Film Coefficient Compute the tube side film coefficient from physical T ¼ average temperature. kg/m3 or lb/ft3 2. Using the mass flow rate per tube. and the (Reynolds number <¼ 2000) [2]: inside film coefficient. tube-side fluid. Note the calculated and turbulent coefficients. (2-16) The area of the triangle is one-half of the area np ¼ number of tube passes in the shell required to accommodate one tube. tube pitch. minimum ¼ 20. Shell Diameter The shell diameter is related to the number of tubes.25 inches (inside do ¼ outside diameter of tubes. TEMA and many others publish tables that list Dtight ¼ 2 (2-15) p the number of tubes that will fit into shells of standard diameters. uses this to calculate the wall temperature and evaluate the transition coefficient is bounded by the laminar the viscosity at that temperature. the spreadsheet there is a possibility of flow oscillations. pffiffiffi 2 3 0:5 Area1 tube. The area of the square is equal Acorrected. if Ds.33.25. nt ¼ number of tubes in the shell For a quick estimation which should suffice for 3. For each tube pass greater than one. to the area required to accommodate one tube. area for all tubes in the shell. For Where: example.5) standard size which is 21. Calculate the diameter of a circle that equates to the tube passes. Avoid the transition region if possible because the cient using the formulae in this section. Similarly. mm or in. diameter) . add cross preliminary design work. For triangular pitch. triangular ¼ 2 ðPR do Þ (2-13) Acorrected 4 Ds. coefficient in Cell D44 and make one or two iterations based on the laminar and turbulent equations. or 1.min ¼ 2 þ 2 do (2-17) 2 p Area1 tube. tube pitch layout. Input an assumed heat transfer coefficient is very unpredictable and value for the film coefficient in Cell D7. for 4. 1. 1. Heat Exchangers 41 0:14 k m NRe 2000 hi ¼ 0:023 NRe0:8 NPr1=3 (2-11) ðhi ÞT ¼ hi þ ðhi hi Þ (2-12) di mw 8000 The “Tubes htc” worksheet calculates the film coeffi- 4. Calculate the cross-sectional area occupied by each Acorrected ¼ Dtight do ðnp 1Þ þ ðNt Areatube Þ tube. Finally. 1. tube diameter.5 inches. and tube omissions to allow space for impingement baffles or to decrease the number of tubes in the baffle 0:5 Nt Areatube windows. round up to the next standard shell size. use this procedure (easily sectional area to account for the pass partition by implemented in Excel): multiplying the tube diameter by Dtight. Calculate the minimum shell diameter by adding square pitch draw the square with corners at the two tube diameters to the circle equating to center of four tubes. However.285. square ¼ ðPR do Þ (2-14) 5. 2. draw the equilateral triangle with vertices at the center of three tubes. and a plausible equation. use the next PR ¼ tube pitch ratio (usually 1. is by changing the assumed value to equal the calculated [2]: result. 5 26.73 0. and 3. bundle bypass flow. and 60 layout with pt/do >¼ 3.1 0.5 0.519 4.408 0. then a total correction of 0. many geometrical properties of the shell such as baffle cut.60 may be used (ho ¼ 0.732: bundle must be known or estimated.5 60 10000 þ 0.476 6.022 6. baffle leakage effects.5 90 0e10 0.476 7 0.52 45 10e100 0.52 45 100e1000 0.602 6.519 48 1 7 0. 45 layout with spacing.593 0.45 0.187 0.5 30 10000 þ 0.657 1.57 0.37 0.37 32.631 1.148 6.123 7 0.45 0.519 45.5 60 1000e10000 0.2 0.4 0.37 0.42 Rules of Thumb for Chemical Engineers Ideal Shell Side Film Coefficient Use the Bell-Delaware method to compute the shell Where: side film coefficient.519 0.37 0.707. Pitch. the equation is: Calculate the ideal heat transfer coefficient for pure pt do crossflow in an ideal tube bank from [13]: As ¼ Lbc Ds Dotl þ ðDotl do Þ !0:14 pn 2=3 ws ks ms hideal ¼ Jideal cps pt ¼ PR do .388 1.93 0.45 0.486 0. bank ho ¼ hideal Jc Jl Jb Js Jr (2-18) As ¼ bundle crossflow area at the centerline of the shell between two baffles Implied by the nature of the correction factors.5 0.395 1. baffle For 30 and 90 tube layout bundles.37 6.9 0.187 0.1 0.707 has “long been used as a rule of thumb” [16].4 0.486 0.97 0.396 1. at the wall temperature.136 6.6 hideal) since this For a 45 and 60 layouts with ratios less than 1.973 7 0.372 0.378 .321 0.477 1.55 0. which is the Pitch Ratio x tube OD As cps ms ms.w pn ¼ pitch normal to the flow direction (see Table 2-6) (2-19) Lbc ¼ baffle spacing Table 2-5 Correlation coefficients for Jideal and fideal [13] Pitch Layout Reynolds Number a1 a2 a3 a4 b1 b2 b3 b4 30 0e10 1.913 6.45 0.498 0.93 0.5 1.519 0.667 1.46 1.388 1.321 0. The Bell-Delaware method computes the heat transfer film coefficient for an ideal bank of tubes.59 0.3 0.3 0.656 1.477 1.378 90 10e100 0. The procedure uses p n do the geometrical properties to calculate each factor.123 7 0.37 0.36 0.5 30 1000e10000 0.59 0.519 0. and Ws ¼ mass flow rate of shell side fluid across the tube adverse temperature gradient build-up if laminar flow.36 0.732 respectively. able baffle spacing in the inlet and outlet sections.52 60 0e10 1.45 0.45 0.388 1.126 6.59 0. vari.657 1.391 0.59 0.52 45 10000 þ 0.378 90 10000 þ 0.5 32 1 6.303 0. As ¼ Lbc Ds Dotl þ ðDotl do Þ pn If the geometrical properties are unknown.57 0.5 3.396 1.45 0. shell diameter.187 0.5 30 10e100 1.52 45 1000e10000 0.107 0.519 45.1 0.266 1.321 0.667 1. subscript w is spacing.3 0.5 60 100e1000 0.37 0.59 0.519 0.519 4.152 7 0.5 60 10e100 1.45 0.0963 6.973 7 0.388 1.593 0.93 0.519 48 1 7 0.667 1.372 0.09 0.45 0.3 0.152 7 0.378 90 100e1000 0.5 30 100e1000 0. as described by Bejan and Kraus [1] Jideal ¼ the Colburn factor for an ideal tube bank and many others.187 0. and outside diameter of the tube pt/do >¼ 1.0815 0.93 0.3 0. The subscript s stands for physical properties at the then applies correction factors to account for baffle cut and average temperature of the shell side fluid.45 0.321 0.5 45 0e10 1.378 90 1000e10000 0.667 1.187 0.37 35 1 6.476 7 0.5 0.333 0.93 0. JL fluid flows more longitudinally. where the shell fluid bypasses the normal flow ranges from about 0. Heat Exchangers 43 Table 2-6 Tube geometry as a function of tube pitch. and baffle cut.8 [13].15 path. 1:33 a Jideal ¼ a1 NRe. listed in Table 2-5. It is flow. It is related to the shell This factor includes tube-to-shell and tube-to-baffle diameter. This factor takes into account the heat transfer rate that occurs in the baffle window where the shell side Baffle Leakage Effects. pn Pitch Parallel to Flow. 1 Fc ¼ ½p þ 2 f sinðarccos fÞ 2 arccos f tional information about the construction of the heat p exchanger. the equation assumes single segmental baffles: Jl ¼ 0:44 ð1 ra Þ (2-23) Jc ¼ 0:55 þ 0:72 Fc (2-22) þ ½1 0:044 ð1 ra Þ expð2:2 rb Þ . depend on the tube Calculate Jideal from the following relationship: pitch layout and Reynolds number.s ¼ (2-20) ms As The coefficients. Dotl ¼ outside diameter of the tube bundle. If baffles are too closely spaced.sa4 NRe. The value leakage. pt Tube Layout Pitch Normal to Flow. If flow in the leakage stream increases compared with cross there are no tubes in the window Jc ¼ 1. Ds 2 lc f ¼ Dotl lc ¼ baffle cut ¼ distance from the baffle to the inside of Baffle Cut and Spacing. Jc the shell. Fc [1].53 for a large baffle cut up to 1.7 and 0. Use this expressed as a fraction of the number of tubes in cross formula [1]: flow. pp pffiffiffi! 30 Triangular Staggered Array pt 3 pffiffiffi pt 2 60 Rotated Triangular Staggered Array 3 pt pt 2 90 Square Inline Array pt pffiffiffi pt 45 Rotated Square Staggered Array 2 pt pt pffiffiffi 2 The Colburn factor is a function of the shell side Where: Reynolds number: a3 a ¼ d o Ws 1 þ 0:14 NRe. deviating from the ideal cross-flow arrangement. as noted. mm or in. Some of the equations require addi. It is typically between 0. the fraction of for small windows with a high window velocity.0 [13]. tube diameter.sa2 (2-21) PR=do Shell Side Film Coefficient Correction Factors Where: This section describes each of the five Bell-Delaware correction factors. mm or in. Use these formulae Ds 2 l c to calculate Jb [1]: q3 ¼ 2 arccos Ds C1 1 C1 ¼ Ds Dotl . Jb p do ð1 Fw Þ Nt dtb Atb ¼ . A rule of thumb is to use one pair of sealing strips for 2p tubes in one window approximately every six tube rows [2]. Sealing strips can increase the value Fw ¼ . Atb. but may be reduced to 0.300 This parameter strongly influences the calculation of Jl. Db ¼ baffle diameter Bundle and Partition Bypass Effects. For exchangers with very small clearances the factor is about 0. ra ¼ Asb þ Atb Aw ¼ Awg Awt .9. and Aw as follows: Where: 1 Ds Asb ¼ ðp q1 Þ Ds dsb . Asb þ Atb (2-26) rb ¼ Aw free area for fluid flow in one window section Calculate Asb.620 0. but larger clear- Where: ances are required for a pull-through floating head where q3 sin q3 the factor is about 0.445 0.004 times the shell diameter limit the baffle-to- shell leak stream. The clearance may be reduced to 0.0156 in to reduce the leak stream between tube and Asb baffle hole [19]). shell-to-outer tube limit distance Jb ¼ exp ½C rc ð1 2 z1=3 Þ for z < (2-27) 2 dtb ¼ baffle-hole diameter tube OD (usually 0.0035 to 0. shell-to-baffle spacing.7.715 0. Awg ¼ ðq2 sin q2 Þ. This factor corrects for flow that bypasses the tube 4 (2-25) bundle due to clearance between the outermost tubes and tube-to-baffle leakage area the shell and pass dividers.100 350 to 425 14 to 17 Pipe 3. .810 0. See Table 2-7. number of tubes in the window Ds dsb ¼ Ds Db .. gross window area 2 8 (2-24) 1 2 lc shell-to-baffle leakage area q2 ¼ arccos Ds p Where: Awt ¼ ntw do .225 1375 to 1500 55 to 60 Rolled 7.8 1 mm or 0.175 0.44 Rules of Thumb for Chemical Engineers Where: 0.125 450 to 575 18 to 23 Pipe 3.03125 in.150 600 to 975 24 to 39 Rolled 4. based on TEMA class R [24] Nominal Shell Diameter Shell Type Difference in Shell-to-Baffle Diameter DN Inches Millimeters Inches 200 to 325 8 to 13 Pipe 2. but only for rolled shells and only if necessary since it is hard to guarantee compliance [19].540 0.4 mm or Or Jb ¼ 1 for z 2 Table 2-7 Diametric shell-to-baffle clearance. fraction of the total number of [13].175 1000 to 1350 40 to 54 Rolled 5. area occupied by tubes in one window 2 lc 4 q1 ¼ arccos 1 ntw ¼ Fw nt . from the baffle tips Lbi and for every 5 to 7 tube pitches thereafter [19].0 [13]. The clean coefficient is: do and di ¼ outside and inside tube diameter. and Lbc are baffle spacing at inlet. Heat Exchangers 45 Where: nb 1 þ ðLi Þð1 nÞ þ ðLo Þð1 nÞ C ¼ 1. 1 in to 3 in.cc nb ¼ number of baffles in the exchanger from 25 mm to 75 mm.s > 100 Js ¼ (2-28) nb 1 þ ðLi Þ þ ðLo Þ Abp rc ¼ As Where: nss z ¼ (API Standard 660 requires a seal device nr. and typically ranges from perform a linear interpolation between the two extreme 0. If NRE.35 for NRE.25 for NRE. and tube wall thermal conductivity.17 for this parameter) Lbo nss ¼ number of sealing strip pairs Lo ¼ Lbc Ds 2 lc nr. The final correction factor is used when the Reynolds assume 2 x Tube OD) number on the shell side is less than 100.85 to 1. through one crossflow section. pp ¼ longitudinal tube pitch and central respectively Abp ¼ Lbc ðDs Dotl þ 0:5 ndp wp Þ Lbc ¼ central baffle spacing. Uo ¼ overall heat transfer coefficient based on the calculate the overall heat transfer coefficient for both the outside area of the tubes clean and fouled conditions. This factor accounts for the conse. Lbo .s < 100.cc ¼ n ¼ 3/5 for turbulent flow or 1/3 for laminar flow pp And Lbi .i do lnðdo =di Þ 1 2-9) þ þ þ Rf . Jr the crossflow stream wp ¼ width of the bypass divider lane (if unknown. Overall Heat Transfer Coefficient Where: Given the tube (inside) and shell (outside) film coeffi- cients.clean ¼ (2-30) ho and hi ¼ outside and inside film coefficients.cc When baffle spacing is increased at the ends of the exchanger to accommodate the nozzles. mm or in.. do do lnðdo =di Þ 1 þ þ respectively di hi 2k ho Rf .s >¼ 100. 1 respectively Uo.s <¼ 100 or 1. Js 10 0:18 Jr ¼ (2-29) nr.cc is the number of effective tube rows crossed flow velocity occur. outlet. leading Li ¼ Lbc to the rule of thumb of 0. ndp ¼ number of bypass divider lanes that are parallel to Temperature Gradient for Laminar Flow Regime. local decreases in Where nr.s <¼ 20: Variations in Baffle Spacing. respectively 1 k ¼ thermal conductivity of the tube material (see Table Uo. fouled ¼ do do Rf . Calculate Js with [25]: values [1].o þ di hi di 2k ho It is good practice to limit the reduction in heat transfer (2-31) due to fouling to about 80% of the clean heat transfer .i ¼ fouling factors on the shell and tube And the coefficient in the fouled condition is: side. fouling factors.0 for NRE.o and Rf . For 20 < NRE. quent decrease in heat transfer. It is equal to 1. Use the b coefficients in Table 2-5 to compute the friction Ws ð2 þ 0:6 ntw Þ DPw. The entrance and exit sections. þ 2 DPb. bypass and baffle leakage effects. from the nozzle to the from 0.0 for z>¼ 0. The baffle windows. first baffle window. Cbp ¼ 4.tw ¼ pp The bundle bypass correction factor uses parameters determined for Jb. it typically ranges from 0.tw and rb (see page 28). Reynolds number > 100. limit for Rb is 1.ideal Rb 1 þ nr. calculations until the values are in reasonable agreement. For a Reynolds number <¼ 100. This is done by instituting a cleaning schedule Use this calculated overall heat transfer coefficient to that removes accumulations before they become too update the assumed coefficient (page 18) and iterate the severe. Shell Side Pressure Drop The Bell-Delaware method accounts for tube bundle 2.ideal Þ Rb þ nb DPw.ideal ¼ (2-33) 0:8 ½lc 0:5 ðDs Dotl þ do Þ 2 rs gc As m s nr.cc mw 0:14 DPb.ideal ¼ 26 pffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi þ As Aw r pt do Dw2 Where: Ws þ b3 As Aw r b ¼ (2-37) 1 þ 0:14 NRe.4 to 0. Cbp ¼ 3. the total pressure drop through the exchanger (excluding pffiffiffiffiffiffi the nozzles) is: Rb ¼ exp½Cbp rc ð1 3 2 zÞ (2-34) DPs ¼ ½ðnb 1Þ ðDPb.5. which depends on the tube 2 gc As Aw rs layout and Reynolds number: If NRe < 100: 1:33 b fideal ¼ b1 NRe. The Combined with the crossflow and baffle window findings.tw Lbc PR=do DPw.5.46 Rules of Thumb for Chemical Engineers coefficient. calculate the pressure drop using the drop that is 20% to 30% of that calculated without the equation corresponding to the flow regime.7.sb2 (2-32) m s Ws nr.cc Rl ¼ exp½1:33 ð1 þ ra Þ rbc (2-35) (2-38) c ¼ 0:15 ð1 þ ra Þ þ 0:8 . It computes a pressure For an ideal window.ideal ¼ (2-36) factor for an ideal tube bank. bypass and leakage effects.ideal Rl The baffle leakage correction factor is a function of ra nr. the film coefficient correction factor for bundle and partition bypass effects. For NRe >¼ 100: 1. The crossflow section between the interior baffles.8 [13].5 to 0. it typically ranges 3.5.sb4 4 Aw The pressure drop for one ideal crossflow section is: Dw ¼ p do ntw þ Ds q2 =2 4 fideal Ws2 nr. rust. Common solution: velocity control. mechanisms [5]: Common solution: reducing the temperature of the heat transfer surface often softens the deposits. • Crystallization. these dissolved salts will . atmospheric Water 450e1140 80e200 High boiling hydrocarbon. particularly at the tube wall. brine 280e680 50e120 Organic solvents. vacuum and high non-condensables Water. Depositing of dirt. water is used. C Btu/h-ft2. atmospheric and high non-condensable Water. Water 570e1140 100e200 Saturated organic solvents with some non-cond Water. and natural waters have a lower solubility in warm water other small particles is also common when fresh than in cold. brine 60e280 10e50 Low boiling hydrocarbon. atmospheric with non-condensables Water 30e170 5e30 Organic solvents. Heat Exchangers 47 Heat Transfer Coefficients Table 2-8 Approximate overall heat transfer coefficients [21] U U Hot Fluid Cold Fluid W/m2. vacuum Water 60e170 10e30 Fouling Resistances The following are the more common fouling crystallize on the surface in the form of scale. Certain salts commonly present in • Sedimentation. Therefore. brine 280e680 50e120 Aromatic vapors. when cooling water is heated. sand. F Sensible Heat Transfer (No Change of Phase) Water Water 850e1700 150e300 Organic solvents Water 280e850 50e150 Gases Water 20e280 3e50 Light oils Water 340e900 60e160 Heavy oils Water 60e280 10e50 Organic solvents Light oil 110e400 20e70 Water Brine 570e1140 100e200 Organic solvents Brine 170e510 30e90 Gases Brine 20e280 3e50 Organic solvents Organic solvents 110e340 20e60 Heavy oils Heavy oils 50e280 8e50 Heaters Steam Water 1400e4300 250e750 Steam Light oils 280e850 50e150 Steam Heavy oils 60e450 10e80 Steam Organic solvents 570e1140 100e200 Steam Gases 30e280 5e50 Dowtherm Gases 20e230 4e40 Dowtherm Heavy oils 50e340 8e60 Flue gas Aromatic HC and Steam 30e85 5e15 Evaporators Steam Water 2000e4300 350e750 Steam Light oils 450e1000 80e180 Steam Heavy oils (vacuum) 140e430 25e75 Steam Organic solvents 570e1140 100e200 Water Refrigerants 430e850 75e150 Organic solvents Refrigerants 170e570 30e100 Condensers Steam (pressure) Water 2000e4300 350e750 Steam (vacuum) Water 1700e3400 300e600 Saturated organic solvents near atmos. For handles are recommended if available for the pipe exchangers with fixed tube sheets. not plates.al. transfer surface. Also. If a stream is known to foul. the method provides much longer 5. geothermal providing little or no excess surface area (that normally is brines.4 mm (1 in. (570 F). do not use a fouling factor for that that will accommodate additional plates in the event that stream. select a frame size be non-fouling. Here is a summary of the no-foul design method from Nesta: Installation Recommendations Here are some installation tips for typical shell-and.1 mm (1.2 ft/s) for 31. This appears where and 38. Smooth surfaces (e. Design for high velocities within erosion and to medium through high boiling point liquid hydrocarbon vibration limits (per the no-foul design method).) tubes. use impingement rods. (consider the possibility of using brushes that would tube heat exchangers [12] and [22]): be at least as long as the tubes). This is a common engineering error provided this “most basic” design algorithm [3]: that leads to oversizing the plate-and-frame exchanger. to fouling than shell-and-tube units. Where impingement protection is temperature gradient between the fluid and the heat required. Check company experience with the heat exchanger The general practice is to specify plate-and-frame to be designed exchangers with no fouling factor.. room to remove the heads and clean the tubes . chrome plated) and mm (0.5 in. because they 6. Keep overdesign between 0% and 20%. By increasing Exceptions to this general high-velocity rule for the velocity of the hydrocarbon above threshold values and fouling mitigation include corrosion. Maximum temperature at the tube wall: 300 C • Corrosion. Common solution: material 1. practices. but to specify a percent 2. Recent research by HTRI [11] shows that fouling in 3.) and 25. allow enough sizes. Decide on fouling factors. using minimum velocities.75 mm (1. and slurries that present an erosion limit.2 m/s (7. mixtures with API gravity less than 45 [19]. side to facilitate cleaning.48 Rules of Thumb for Chemical Engineers • Biological growth. 3.6 tion. minimum copper or copper alloys reduce biological growth. If a stream is determined to of excess surface area instead. Plate-and-frame heat exchangers are usually less prone but do not apply a fouling factor. allocated for fouling). Bennett.75 in. and the composition of the stream. Common solution: reducing the temperature m/s (2 ft/s). Provide up to 20% excess surface area when both streams are within the scope of this design practice.g. between the fluid and the heat transfer surface.) tubes. et.) • Chemical reaction coking. calculated overall U. Also. 5. Common solution: reducing the E and J shells. Shell design should use single segmental baffles surface can cause solidification of some of the fluid with 20% cut. Place the most heavily fouling stream on the tube- crude oil preheat service depends primarily on velocity. run time than traditional designs. • Freezing fouling. • Provide valves and bypasses in the piping system for • Provide sufficient clearance for removing the tube both the shell and tube sides. Shell side: minimum cross-flow stream velocity 0. Provide pressure drop as required to achieve the have much higher overall heat transfer coefficients. 1. and to avoid surface temperature. Ball valves with locking bundle at the head end of the exchanger. the same fouling resistance values as for a shell-and-tube exchanger has a proportionally greater effect on the Building on the no-foul design method.6 ft/s) for 19 selection.25 in. if necessary. Tube side: minimum velocity 2 m/s (6. 2. oriented horizontally for TEMA Type stream components. use a fouling more surface is needed because of a loss of performance factor in accordance with the company’s best due to fouling. the areas of low velocity that occur on the shellside Nesta outlined a “no foul design method” that is applicable 4. velocity 2. Overcooling at the heat transfer 4. Common solution: material selection. hydrocarbons deposit in a high temperature applica. 4 Monel 26 15 Nickel 90 52 Tantalum 54 31 Titanium 21 12 Type 316 stainless steel 16. • Condensate drainage pipes should have a vertical drop-leg of at least 18 inches from the exchanger to the trap. use a steam trap. Install a condensate drip steam service.500 kg/h (8.9 14. Condensate • Re-torque all external bolted joints after installation accumulating in the exchanger results in water and again after the exchanger has been heated to hammer and poor temperature control. possible flooding due to faulty trap operation. no vertical lift before or after steam traps.4 . Table 2-9 Thermal conductivity of metals used in heat exchangers Heat Exchanger Tube Material k.5 8.5% C) 36 @ 20 C 21 @ 68 F 33 @ 400 C 19 @ 750 F Copper 386 223 Hastelloy C 8.000 lb/h) in the piping at each inlet and outlet. Install a strainer in front of the control valve. be omitted from the piping. • Exchangers in condensing steam duty should be • Quick-opening and closing valves controlling fluids installed at a 3 to 4 slope.5% C) 54 @ 20 C 31 @ 68 F Carbon steel (1. Some exchangers are that use a control valve with level controller. Btu/h-ft- F Aluminum 147 85 Brass. and vacuum breakers for exchangers in condensate to the boiler. Admiralty 111 64 Brass. Thermal conductivity is the quantity of heat transferred through a unit thickness. Red 159 92 Carbon steel (0. • Insulate all heat-transfer-exposed surface areas. designed with these features. • Ensure that foundations are adequately sized. valve. corrosion prevent leaks and blowing out of gaskets. to to or from an exchanger may cause water-hammer. problems may also occur. for capacities higher than to the unit as practicable.7 5 Inconel 14.4 Type 410 stainless steel 24. • Do not pipe drain connections to a common closed • Loosen foundation bolts at one end of unit to allow manifold. The normal locations are close to the pocket with a steam trap in front of the steam control steam inlet or on the top portion of the shell. facilitate drainage of condensate. Thermal Conductivity of Metals Use the values in Table 2-9 when computing overall heat transfer coefficients (page 45). free expansion and contraction of the heat exchanger • Install a gage glass in a vapor or gas space to indicate shell. foundation bolts set in pipe sleeves the exchanger. located as close or less.3 9. toward the shell outlet. Heat Exchangers 49 • Provide thermowells and pressure gage connections • For condensate capacities of 3. in which case they can • If the steam supply is modulated with a control valve. all condensate drains must flow by gravity to • Provide valves to allow venting of gas vapor from the a collection tank or pumping system to return the exchanger. and use a pipe size equal to or larger of larger size than the bolt size will allow for than the inlet connection to the unit. In Locate the valve at least 10 pipe diameters away from concrete footings. W/m-K k. adjustment after the foundation has set. Heat exchangers and care should be taken for proper selection of such should be installed to promote gravity drainage with equipment. heat transfer as required by the controller. This will Control. The inerts then pile up at the outlet end lowering instrumentation. and lower the percentage of surface blanketed. variable leakage of inerts. It is necessary to have some over-surface and increase cooling and bring the pressure back down to the to have a proper baffling to allow for pressure control desired level. A relatively Figure 2-6. The condenser must be considered part of the Figure 2-6 shows typical baffling. The column Outlet Temperature and Pressure. This is simply a matter of adjusting the designer adds 50% to the calculated length for the over. The inerts move control system (similar to extra trays in a fractionator) to through the first part of the condenser as directed by the allow for process swings not controlled by conventional baffles. vacuum system along with the inerts. etc. The condensers for vacuum distillation [20]. It is important to pressure is controlled by varying the percentage of the have proper subcooling in the vent end of the unit to tube surface blanketed. One below that desired. . When the desired pressure is prevent large amounts of process vapors from going to the exceeded.50 Rules of Thumb for Chemical Engineers Vacuum Condensers This section provides tips for designing overhead The inerts will “blanket” a portion of the tubes. the vacuum system will suck out more inerts. blanketed portion has very poor heat transfer. The vapor inlet nozzle is expanded to five times its area. surface. heat transfer coefficient to heat balance the system. The reverse happens if the pressure falls during process swings. Baffling and inlet “bathtub” are shown in this typical vacuum condenser design. Air-cooled Heat Exchangers: Forced vs. Baffles should prevent bypass of inlet condensing section and decrease heat transfer until the vapor into the vent. can be cooled to about 10 C to located on the suction side of the fan. Note very poor. bearings. of the fan. at about 2½ times velocity and absence of stack the intake velocity. Then the vacuum valve opens Typical Condenser. Enough baffles pressure drop as low as possible. the inerts build up in the Bypassing. rature of the air.4-mm (1-in. rain. 12. Pressure Drop. Baffling must be designed to keep the and the side-to-side (40% cut) baffles. Pressure drop inerts which are subcooled before being pulled out as is lower at the outlet end because of smaller mass flow. Air flows at a velocity of 3 to 6 m/s (10 to Typically. belts. Without proper baffles. or about 450 m/min (25 ft/s) Influence of weather conditions Total exposure of tubes to sun. or induced draft when the tube section is flowing inside the tubes. The higher the pressure must be used in the inlet end for minimum tube support. Heat Exchangers 51 large section must be covered by more or less stagnant job of attaining proper vent end subcooling.) OD carbon steel tubes are 20 ft/s). and other components in the air stream Temperature limit e tubeside Limited by tube components Limited to 175 C (350 F) because fan failure could process fluid subject fan blades and bearings to excessive temperatures Maintenance Better access to mechanical Mechanical components are more difficult to access components because they are above the tubes . In drop the higher the energy consumption and the harder the the last 25% of the outlet end a spacing of 1/10 of a diam- eter is recommended. Under these conditions pressure control will be entrance to exchanger and across first rows of tubes. Induced Draft Air-cooled heat exchangers are classified as forced inch). and hail because and hail 60% of face is covered Freezing conditions Easily adaptable for warm air recirculation Warm discharge air not recirculated during freezing conditions Result of fan failure Low natural draft capability on Natural draft stack effect is greater than fan failure due to small stack forced draft type effect Power requirement Slightly lower fan power because Slightly higher fan power because the fan is the fan is located in the cold air located in the hot air stream (air has lower density) stream (air has higher density) Temperature limit e discharge No limit Limited to about 95 C (200 F) to prevent potential air stream damage to fan blades. pressure gets too high.9 mm high (½ to ⅝ Table 2-10 Comparison of forced draft and induced draft air-cooled heat exchangers [8] Attribute Forced Draft Induced Draft Distribution of air across section Poor distribution of air over the Better section Effluent air recirculation Greatly increased possibility of hot Lower possibility because fan discharges air to intake air recirculation due to low discharge upward.7 to 15. fitted with aluminum fins. rain. needed. pulling process vapor and inerts into the vacuum “bathtub” used for low vacuums to limit pressure drop at system. Figure 2-6 illustrates an inlet wider. providing outside surface area about 14 to 21 times draft when the tube section is located on the discharge side greater than the area of the bare tubes. Forced draft units 15 C (20 F to 30 F) above the dry-bulb tempe- are more common. This is very important. away from the tubes. Less effect from sun. The process stream. 25. the staggered baffle spacing with large spacing at inlet. 07 0. On the other hand.92 0.67 1.11 0. order to obtain a feel for the performance range to expect. Also.93 0.75 900 (3.73 0. Wind and rain patterns should also be considered [7].000) 0.80 0.000) 1.82 0. the corrected log.88 0.000) 0.15 1.76 0. optimizing the design of an air-cooled heat exchanger. and because they provide an external surface area about 20 power). For four or more tube passes those for shell-and-tube exchangers. calculation of the air side film coefficient is a good design.86 0. As a result.000) 0.69 1. Q ¼ U A MTD. the finned tubes partially offset the poor thermal performance The air density affects fan design (flow.70 0. m (ft) L 20 C (0 F) 20 C (70 F) 40 C (100 F) 90 C (200 F) 0 1. Table 2-11 Approximate correction factor for air density as a function of altitude and temperature Air Temperature Altitude. The air flows on the outside of the the design temperatures or number of passes to obtain tubes.00 0. head. tubes.92 0.800 (6. The performance of air-coolers is tied to the dry-bulb Air data should include environmental characteristics.99 0. in design accommodations such as increased tube pitch. air temperature.000) 1. fins.000) 0.96 0. Air-cooled exchangers trans- • Annual temperature-probability curve fer less than 10% of that of water-cooled shell-and-tube • Typical daily temperature curves units.89 0.100 (7.400 (8.79 0. Use the charts for one.64 2. The process fluid the correction factor is 1.90 0.and two-pass coolers.83 0. it is slightly less than 1 for three- flows inside the tubes. which is generally in the order of 60 the design air temperature [7]: W/m2- C (10 Btu/h-ft2- F).500 (5. and the inside heat transfer film pass units. which varies considerably throughout the Marine air or sulfur dioxide content can be corrosive to year. Table 2-11 gives values for correction factors for times that of plain tubes. mean temperature difference is determined from charts Mukherjee discussed each of these variables in terms of . Engineers can juggle at least nine variables when For the heat balance. complicated.03 0. and structures. indicating incorporation of performance of the cooler at the higher end of the fouling factors in the design and possibly suggesting temperatures that are known to occur at the plant site.62 2.77 0.000) 0.85 0. Assume a design temperature that is exceeded during fans. altitude and temperature.80 300 (1.8 then strongly consider changing shell-and-tube units.77 600 (2.91 0. the specific heat of air is only 25% that of • Duration-frequency curves for the occurrence of the water (on a mass basis). some guidance is given later in this section.200 (4.60 Air-cooled Heat Exchangers: Thermal Design Thermal performance calculations are analogous with (Figure 2-7 and Figure 2-8).96 0.86 0. air coolers are very maximum dry-bulb temperature large relative to water coolers.000) 1.52 Rules of Thumb for Chemical Engineers Air-cooled Heat Exchangers: Air Data The overall heat transfer coefficient is governed by the Obtain the following data to get a realistic estimate of air film heat transfer. but calculate the lead to increased fouling.74 0. If coefficient is calculated exactly the same way as with the factor is less than 0.72 1. Dusty atmospheres may 2% to 5% of the annual time period. cross-flow. Heat Exchangers 53 Figure 2-7. MTD correction factors for air-cooled heat exchangers (1-pass. both fluids unmixed) [8]. . cross-flow. both fluids unmixed) [8]. Figure 2-8. MTD correction factors for air-cooled heat exchangers (2-pass. but can range from three to ten rows Air-side film coefficient varies inversely with number of tube rows More rows advantage: more heat transfer area in the same bundle width.) As tube pitch is decreased.5 power of air mass velocity Air-side pressure drop varies to the 1.9 mm (3/8 in.54 Rules of Thumb for Chemical Engineers economic impact. effective Table 2-12 Variables that must be optimized for air-cooled heat exchanger design [18] Variable Considerations Air flow rate Rule of thumb for face velocity approaching the tube bundle (total flow divided by total area of bundle): e 3 row coil: 240 to 275 m/min (800 to 900 ft/min) e 4 row coil: 150 to 210 m/min (500 to 700 ft/min) e 5 row coil: 140 to 180 m/min (450 to 600 ft/min) e 6 row coil: 100 to 150 m/min (350 to 500 ft/min) Air-side film coefficient varies to the 0. air-cooler [7]: calculate the tube-side pressure drop and check this against the allowable pressure drop. 5. There are usually two bundles in a section.5 m (10 ft to 11. Ganapathy has described a procedure for designing an 4. fans are commonly 3.5 ft).375 in. H2O and 0. Therefore. calculate film coefficients and overall heat transfer coefficient.6 m to 4. 12.) 25 mm / 57 mm / 67 mm (1 in / 2.3 m (12 ft to 14 ft) in diameter. and 5/8 in. six or seven tube rows for gas coolers and viscous liquid hydrocarbon coolers Number of tube Distribution of tubes in the various passes need not be uniform. Identify all process and site data.) Selection depends on relative values of air-side and tube-side film coefficients With higher fins.5 mm.3 in. When the required surface fits the assumed layout. Tube outside Cost of exchanger is lower with smaller diameter tubes diameter Cleaning is more difficult with smaller diameter Minimum recommended (and most common) tube size is 25 mm (1 in) OD Optimize with pressure drop by adjusting the number of passes and tube size Fin height Usual fin heights are 9.25 in / 2.75 power of the air mass velocity Optimum air mass velocity is higher when air-side heat transfer coefficient is highly controlling (e. 3. tubes are typically in the range of 8 m to 10 m long (26 ft to 33 ft).625 in. and two fans per section. Bundle width normally limited to 3. use lower density for gas coolers and viscous liquid hydrocarbon coolers Tube pitch Staggered pattern almost invariably employed Designers tend to use the following combinations of bare-tube OD. use higher fins for steam condensers and water coolers Typically.2 m to 3.g. surface against the assumed layout. API 661 specifies minimum fan coverage of 40%.. finned-tube OD. H2O .75 power of air mass velocity Tube length Length is established in conjunction with the bundle width. use higher density for steam condensers and water coolers Typically. and 15. four or five tube rows for steam condensers and water coolers Typically. air tempera- fied. 1. steam condensers and water coolers) Exchangers are usually designed with a pressure drop between 0.7 mm. Assume the layout of the tube bundle. air-side pressure drop and power consumption increase more rapidly than the air-side heat transfer coefficient Number of tube Most exchangers have four to six tube rows.7 in.. For the assumed values. highlights are given in Table 2-12 temperature difference.) Typically. and surface area. horsepower. and fin geometry. When surface and tube-side pressure drop are veri- 2.. use lower fins for gas coolers and viscous liquid hydrocarbon coolers Fin spacing Spacing usually varies between 276 to 433 fins/m (7 to 11 fins/in. and tube pitch: 25 mm / 50 mm /60 mm (1 in / 2 in / 2. 1/2 in. check this [18]. reducing number of bundles and sections More rows disadvantage: increases fan horsepower for the same air velocity and lowers the Mean Temperature Difference Typically. especially useful in condensers where the flow area in each passes pass can be gradually reduced as the liquid fraction increases progressively Optimize to obtain uniform pressure drop in each pass Fan power Power varies directly with volumetric air flow rate and pressure drop consumption Fan horsepower varies to the 2. fewer tubes can be accommodated per row Typically. calculate the air-side pressure drop and fan ture rise or mass flowrate. m or ft 41 mm (0. W/m2-C or H ¼ height of fin. Btu/h-ft2-F s ¼ spacing between fin centers. to 0. 0.) and fin heights from 1. Exchangers are pt ¼ tube pitch. kair do rair umax 0:68 H Y ho ¼ C ðNPr Þ1=3 mair ¼ viscosity of air. 1. Heat Exchangers 55 Air-Side Heat Transfer Coefficient C ¼ coefficient (includes units conversion). Y ¼ thickness of fin. to 0. 1005 J/kg-C or 0.). mm or in. m/h or ft/h equilateral triangular pitch tube banks with pitches up to umax is related to the face velocity of the air 4.0438 lbm/ft-h do mair s s cp mair (2-39) NPr ¼ Prandtl number. approaching the tube bundle by the ratio of total face 0:2 0:12 area to open area between tubes. First.61 in.134 (US) The Briggs and Young correlation (as reported in [2]) kair ¼ thermal conductivity of air.192 ¼ in. It was developed 0. the tubes were in umax ¼ maximum velocity of air.6 mm (0. The manufacturer of the heat fluid outlet temperature: switching fans on and off. mm or in. Air Side Calculate the air side pressure drop with the Robinson Where: and Briggs correlation (as reported in [2]).4 mm to rair ¼ density of air. Air-cooled Heat Exchangers: Pressure Drop.9 mm to 3 mm (0. 1 m/s2 or 32.0000181 Pa-s or 0.056 in.035 in. 0. mm or in.44 in.5 in.17 ft/s2 friction factor in consistent units: The other variables are the same as for Equation 2-39.3 in H2O and 0. Fin spacings ranged from Table 2-11) 0. dimensionless ¼ kair Where: cp ¼ heat capacity of air. ho.23 kg/m3 or 0.026 W/m-C or solves for the air-side film coefficient.000231 (SI) or 0. Then: 2 f n rair ðumax Þ2 DPair ¼ (2-41) gc Air-cooled Heat Exchangers: Temperature Control Various methods are used to control the process the tube sections.0765 lb/ft3 (see 16.).24 Btu/lb-F ho ¼ air-side heat transfer film coefficient.015 Btu/h-ft-F empirically using data from tube diameters from 11 mm to do ¼ outside diameter of tube (without fins). . and adjustable shutters mounted above the unit. m or ft usually designed with a pressure drop between 75 Pa and n ¼ number of tube rows in the bundle 175 Pa (0. especially for umax. use exchanger will normally recommend the best of two-speed or variable-speed motors.117 in.7 in H2O). 0.652 in.81 ¼ Pa) or lbf/ft2 (x 0. do rair umax 0:32 pt 0:93 f ¼ 9:47 (2-40) but be sure the units are consistent. calculate the gc ¼ conversion factor. H2O). to 1. use of variable solution after consulting with the buyer and designing pitch fan blades. mair do Results will be kg/m2 (x 9. F pt ¼ tube pitch Db ¼ baffle diameter pn ¼ tube pitch normal to the flow direction Dotl ¼ outside diameter of the tube bundle. mm or in. C or Btu/ft. F Btu/h-ft2. C or Btu/h-ft2. W/m2. C or F subscripts) U ¼ overall heat transfer coefficient. 1 m/s2 or 32. C or Btu/lb. W/m2. consistent units do ¼ outside tube diameter.56 Rules of Thumb for Chemical Engineers Nomenclature nb ¼ number of baffles in the exchanger nr. F L ¼ tube length u ¼ velocity in tubes. m2 or ft2 nr. F ¼ mean temperature difference between hot and DT mean J ¼ Bell Delaware correction factor (various cold streams. inside of the shell. mm or in. cP NPr ¼ Prandtl number ¼ z ¼ ratio of sealing strip pairs to tube rows in k dru crossflow section NRe ¼ Reynolds number ¼ m . mm or in. H ¼ height of fin. pp ¼ tube pitch parallel to the flow direction Ds ¼ inside diameter of the shell Q ¼ heat transferred. mm or in.tw ¼ effective tube rows crossed in the window Abp ¼ tube bundle bypass area section As ¼ free flow area through one crossflow section nt ¼ number of tubes evaluated at centerline nss ¼ number of sealing strip pairs Asb ¼ shell to baffle leakage area for a single baffle ntw ¼ number of tubes in a baffle window Atb ¼ tube to baffle leakage area for a single baffle np ¼ number of passes Aw ¼ area available for flow through a single baffle ndp ¼ number of bypass dividers parallel to crossflow window stream Awg ¼ flow area through a single baffle window with DPt ¼ pressure drop through turns. f ¼ friction factor C or F gc ¼ conversion factor. r ¼ density.17 ft/s2 t ¼ inlet and outlet temperatures of the cold stream. Pa or psf (divide no tubes by 144 for psi) Awt ¼ window area that is occupied by tubes PR ¼ pitch ratio cp ¼ heat capacity. m/s or ft/s Lbc ¼ central baffle spacing W ¼ mass flow rate Lbi ¼ baffle spacing at inlet wp ¼ width of bypass divider lanes that are parallel to Lbo ¼ baffle spacing at outlet the crossflow stream lc ¼ baffle cut ¼ distance from the baffle to the Y ¼ thickness of fin. W/m. C or k ¼ thermal conductivity. W or Btu/h Dw ¼ effective diameter of a baffle window Rf ¼ fouling factor di ¼ inside tube diameter. mm or in.cc ¼ effective tube rows crossed through one A ¼ heat transfer area. rb ¼ ðAsb þ Atb Þ=Aw dimensionless Fc ¼ fraction of cross sectional area in the crossflow rc ¼ Abp =As section Fw ¼ fraction of cross sectional area in the baffle s ¼ spacing between fin centers. window T ¼ inlet and outlet temperatures of the hot stream. C or F h ¼ film coefficient. usually calculated at the crossflow section outside tube diameter. consistent units ra ¼ Asb =ðAsb þ Atb Þ F ¼ LMTD configuration correction factor. kg/m3 or lb/ft3 cp m m ¼ viscosity. kJ/kg. U. [16] Leong K.14(3):217–24. Hydrocarbon Processing. Mon MS. [22] Plant Support & Evaluations. downloaded from www. March 27. in January. and Guy Z. Armstrong Pumps. King D. 2001. Heat Exchanger May. [17] Moretta A. May. [11] Heat Transfer Research. Mueller A. Lin KA. Process-design Criteria. Boca Raton: CRC (46):604. System Heat-Transfer Solutions”. vol. Press. [25] Than STM. Chemical Engi- Difference in Design. 2007:40. July. www. Alanis F. 1980. A Best Practices neering Data Book. Inc. John 2009:40–3. Exchangers. Lestina T. Leong Y. Liu H.tema. Kraus A. Design. Journal of [20] Personal communications between Carl Branan. Chemical [21] Pfaudler Corporation.htri. Toh K. Chemical Engineering. 2002. Hydrocarbon Processing. (HTRI). 2011. [15] Kern D. 1998. Wiley & Sons. American Society for Mechanical Jack Hailer. Improving cations. 1. February [1] Bejan A. 2nd ed. 2002:32. Chemical Engineering Progress. Effectively Design Air-Cooled Heat [5] Delta T Heat Exchangers. Inc. employed at El Paso Products Company. Nagle W. 2010:44. Chemical Engi. Chemical Engineering Education. NY.org. (DOE).65. Compact Heat Technologies Program. www. Steam Technical Brief published by the Industrial [9] Gunnarsson J. Department of Energy neering. Kistler RS.9:92. 2008. Engineering. Reduce fouling in shell-and-tube heat [6] Fakheri A. 2004: tion of the Log Mean Temperature Correction Factor 77–82. Fouling in Heat Exchangers. Mueller A. [2] Bell K. 2003.com February. 2007. Society of Mechanical Engineers. exchangers.deltathx.. What’s the Difference Between TEMA Heat Exchangers. International Journal of Engineering Heat Exchanger Designs. Rochester. and Thermal Design. Alabama: Wolverine Tube.com.125:527.net [24] TEMA. Heat Transfer Handbook. A General Expression for the Determina. 2003. www. [19] Nesta J. 9th ed.. Sinclair I. 2007. File 138.wlv. Shell-and-Tube Heat Exchangers”. Moore while all were Engineers (ASME).62:283.. Designing Plate-and-Frame [23] Rubin F. Exchanger Design Software for Educational Appli- [3] Bennett C. Kraus A. Exchanger Classes. World Academy of Science. April.waset. Plate Heat Exchangers: Avoiding Common Misconceptions. June September. Rating. Transactions of the American neering. and Technology. Wolverine Engineering Data New York: McGraw-Hill. Heat Transfer. October.S. Spiral Heat Exchangers: Sizing Units for [4] Bowman R. Heat Exchangers: Selection. June 2003. [18] Mukherjee R. 1978:112–9. SI Version.armstrongpumps. Renewable Energy. 12th ed.org. Engineering [13] Kakaç S.com/products. Extended Surface Heat Transfer. Progress. Energy Efficiency and Exchangers: Improving Recovery. Shell and Tube Heat published online at www. Engi. Mean Temperature Cooling Non-Newtonian Slurries. [7] Ganapathy V. February 2009:44–7. 1997:26–47. June. www. DOE/GO-102003–1738. . Chemical Engineering Progress. May 1940. Inc. for Shell and Tube Heat Exchangers.2004. 1972. Book. “Industrial Steam [8] Gas Processors Suppliers Association (GPSA). Standards of the Tubular Exchanger [12] “Installation and Operating Instructions for Armstrong Manufacturers Association. Polley G. Huntsville. Inc. Heat Exchangers 57 References [14] Kerner J. [10] Haslego C.
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