This section presents introductory information and general design guidelines. Heat transfer and fluid flow correlations and rigorous analytical methods are referenced. manual contents: Heat Transfer Economic Pressure Drop and Velocity Flow Splitting, Fouling Tube Vibration Enhanced Surfaces
200 Design Background Abstract This section presents introductory information and general design guidelines Heat transfer and fluid flow correlations and rigorous analytical methods are referenced Chevron Corporation Contents Page 210 Heat Transfer 200-2 211 Mean Temperature Difference 212 Overall Heat Transfer Coefficient 213 Single-phase 214 Two-Phase Liquid/Gas Heat Transfer 215 Boiling 216 Condensing 220 Economic Pressure Drop and Velocity 200-12 230 Flow Splitting 200-14 240 Fouling 200-15 250 Overdesign 200-16 260 Tube Vibration 200-17 270 Enhanced Surfaces 200-19 280 Computer Program Abstracts 200-19 200-1 December 1989 200 Design Background Heat Exchanger and Cooling Tower Manual 210 Heat Transfer The rate equation for heat transfer is: Q = U ⋅ A ⋅ MTD (Eq 200-1) where: Q = Heat duty, Btu/hr U = Overall heat transfer coefficient, Btu/(hr ⋅ °F ⋅ ft2) A = Heat transfer area, ft2 MTD = Mean temperature difference, °F The area, A, and the reference area in the overall coefficient, U, must be the same Outside bare tube surface area is the usual reference area for tubular equipment Estimation methods for mean temperature differences and overall heat transfer coefficients are discussed in the following sections Flow in tubes and flow across bare tube bundles for single-phase and some multiphase cases are covered 211 Mean Temperature Difference Mean temperature difference, MTD, is defined as the area weighted average temperature difference between hot and cold streams in the heat exchanger: MTD ≡ (1/A) A∫ (Th-tc) dA (Eq 200-2) where: Th = Local hot fluid temperature, °F tc = Local cold fluid temperature, °F MTD depends on flow arrangement and has been calculated in dimensionless form for commonly used flow arrangements, with the assumptions that U is constant and the heat content of each stream varies linearly with temperature Several different forms of MTD charts are available in the literature The most useful form involves cross plots of four dimensionless parameters (R, N, P and θ) defined as follows R = Relative Heat Capacity Ratio = Cc/Ch = (Ti-To)/(to-ti) N = Number of Transfer Units = U⋅A/Cc P = Thermal Effectiveness = (to-ti)/(Ti-ti) θ = Dimensionless MTD = MTD/(Ti-ti) December 1989 200-2 Chevron Corporation Heat Exchanger and Cooling Tower Manual 200 Design Background where: Ch = (M ⋅ Cp)h = Hot stream heat capacity rate, Btu/hr ⋅ °F Cc = (M ⋅ Cp)c = Cold stream heat capacity rate, Btu/hr ⋅ °F M = Mass flow Rate, lb/hr Cp = Specific heat, Btu/lb ⋅ °F Ti = Hot stream inlet temperature, °F To = Hot stream outlet temperature, °F ti = Cold stream inlet temperature, °F to = Cold stream outlet temperature, °F Graphs relating R, N, P, and θ are presented in Appendix A for commonly used flow arrangements These graphs are useful for design of individual exchangers, evaluation of exchanger performance at nondesign conditions, evaluation of heat exchanger network performance, and evaluation of field performance data, as described below For design, terminal temperatures (and therefore P and R) are known, and required area can be calculated from N For evaluation of networks or alternative design conditions, N, R and inlet temperatures are known, and outlet temperatures can be calculated from P For evaluation of field performance data, terminal temperatures (R and P) are known, and U can be calculated from N A more common presentation of MTD information is F-factor graphs, where F is defined as the actual mean temperature difference (MTD) divided by the “log mean temperature difference” (LMTD) LMTD is the actual MTD for counterflow (F=1), and is calculated as follows: LMTD = [(Ti-to)-(To-ti)] / ln [(Ti-to)/(To-ti)] (Eq 200-3) Using the definition of F, Equation 200-1 becomes: Q = U ⋅ A ⋅ F ⋅ (LMTD) (Eq 200-4) F-graphs are plots of F against P with R as a parameter They provide the same information as θ on the more general graphs Both the general graphs and F-graphs are provided in Appendix A for each flow arrangement Only the general graph is given for counterflow (because F≡1 for counterflow) F=1 for pure component (isothermal) boiling and condensing regardless of flow arrangement MTD graphs described above are based on the assumption that the heat content of each stream varies linearly with temperature This may not be correct where phase change is involved For example, cooling a superheated vapor may involve a variable temperature desuperheating zone, an isothermal condensing zone, and a vari- Chevron Corporation 200-3 December 1989 200 Design Background Heat Exchanger and Cooling Tower Manual able temperature subcooling zone Such cases should be analyzed in segments, as indicated in Figure 200-1, so that the linear assumption is valid for each segment Fig 200-1 Nonlinear Heat Release 212 Overall Heat Transfer Coefficient Heat transfer coefficients can also be considered as reciprocals to heat transfer resistances The overall resistance is the sum of the individual resistances illustrated in Figure 200-2 Individual inside and outside film and fouling resistances are customarily referred to the tube surface at which they occur, whereas the overall resistance and overall heat transfer coefficient are customarily referred to the outside tube surface Individual resistances are therefore added as follows: 1/U = (1/hi + Rfi)(Ao/Ai) + Rw + 1/ho + Rfo (Eq 200-5) where: hi = Inside film coefficient, Btu/hr ⋅ °F ⋅ ft2 Rfi = Inside fouling resistance, hr ⋅ °F ⋅ ft2/Btu Ao = Outside area, ft2 Ai = Inside area, ft2 Rw = Tube wall resistance, hr ⋅ °F ⋅ ft2/Btu December 1989 200-4 Chevron Corporation Heat Exchanger and Cooling Tower Manual 200 Design Background ho = Outside film coefficient, Btu/hr ⋅ °F ⋅ ft2 Rfo = Outside fouling resistance, hr ⋅ °F ⋅ ft2/Btu The tube wall resistance is given by: ( d o ⁄ 12 ) R w = ln 2k w do - - di (Eq 200-6) where: = Tube O.D., inches di = Tube I.D., inches kw = Tube wall thermal conductivity, Btu/hr ⋅ °F ⋅ ft Estimation of film coefficients and the factors that control them are discussed below Fig 200-2 Heat Transfer Through Circular Tubes 213 Single-phase This section discusses single-phase convective heat transfer inside tubes and outside bare tube bundles Externally finned tubes are discussed in Section 270 Analysis of heat transfer in commercial heat exchangers can be extremely complex For in-tube flow, local tube side heat transfer coefficients vary along the length of the tube as the flow structure develops, may include both natural and force convection components, and may involve significant fluid property variation in both radial and axial directions Chevron Corporation 200-5 December 1989 200 Design Background Heat Exchanger and Cooling Tower Manual Shell side flow involves similar effects, but is much more complicated The various flow paths that exist on the shell side of an exchanger are illustrated in Figure 200-3 Fig 200-3 Shell Side Flow Streams Each flow fraction has a different heat transfer effect and may vary along the length of the exchanger “B-stream” flow dominates overall shell side heat transfer It is most effective in the cross flow region between baffle tips and somewhat less effective in the window region around the end of the baffle “A-stream” is effective, but applies only to the short section of tube in the baffle hole “C,” “E,” and “F” streams are relatively ineffective and are usually minimized by the use of seal bars, dummy tubes, and small clearances “C,” “E,” and “F” should each be less than 0.1 Rigorous computer simulation is required to analyze all of the complexities that exist on shell and tube sides of a heat exchanger Good designs, however, fall within narrow limits where design and evaluation procedures can be greatly simplified For example, a typical shell and tube heat exchanger for pumped liquids is a TEMA “AEU” with 3/4-inch, 14 BWG (minimum wall) carbon steel tubes on 1-inch square pitch, 45 degree tube layout angle, 20% cut segmental baffles, TEMA standard clearances, and about 0.25 psi/ft tube side pressure gradient and 0.5 psi/ft (axial) shell side pressure gradient This is near optimum for most pumped liquid cases as discussed in Section 220 Rigorous computer generated heat transfer coefficients for this case are shown in Figure 200-4 These curves may be used for initial estimates and scoping studies Final designs should be checked using the HTRI ST computer program Figure 200-4 applies to water and to hydrocarbons The difference between the water and hydrocarbon curves reflects the difference in fluid properties Hydrocarbon fluid properties also vary widely; however, they vary with each other and with temperature in such a way that heat transfer can be correlated to viscosity and density Figure 200-4 applies to hydrocarbons extending from naphtha to residuum December 1989 200-6 Chevron Corporation Heat Exchanger and Cooling Tower Manual 200 Design Background and for temperatures from ambient to 650°F Overall accuracy of the curves is about 15% Fig 200-4 Heat Transfer Coefficients for Liquid Phase Shell and Tube Heat Exchangers (See Section 213 for instructions and limitations.) Heat transfer film coefficients in Figure 200-4 are presented in the conventional way Tube side coefficients, hi, are referred to the inside tube surface area Shell side coefficients, ho, are referred to the outside tube surface area The overall heat transfer coefficient, referred to the outside surface, is: U = 1/[ 1/ho + Rw + (1/hi)/(Ai/Ao) ] where the wall resistance, Rw, is 0.0003 hr ⋅ °F ⋅ ft2/Btu and the area ratio, Ai/Ao, is 3/4 To compare shell and tube side film coefficients, both coefficients should be referred to the same surface area, usually the outside surface That is, the shell side coefficient should be compared to 0.75 times the tube side coefficient in Figure 200-4 Chevron Corporation 200-7 December 1989 200 Design Background Heat Exchanger and Cooling Tower Manual From Figure 200-4, 75% of the tube side coefficient is approximately equal to the shell side coefficient for both water and hydrocarbon in the turbulent regime (viscosity < centipoise) This is as it should be Equal expenditure of pumping power per unit of heat transfer surface and equally efficient conversion of pressure drop to heat transfer result in equal heat transfer coefficients in the turbulent regime In the laminar regime (viscosities > 10 centipoise), the shell side coefficient is about seven times the tube side coefficient when referred to the same surface area In-tube laminar flow heat exchange is never economical, sometimes leads to “viscosity plugging” (see Section 610), and should be avoided The pumping power expended per unit of heat transfer surface is the controlling factor influencing heat transfer in turbulent flow Where expensive alloys are needed, it is usually economical to spend more on pumping power to save on exchanger costs Film coefficients vary at about the 0.4 power of the pressure gradient General heat transfer correlations for liquids and gases are given in Appendix B Most of this information is taken from the HTRI Design Manual and is proprietary Equation B-1 in Appendix B applies to turbulent flow of liquids or gases in tubes and is accurate within about 15% This is a Dittus-Boelter type correlation and is similar to those found in general heat transfer text books Corresponding text book correlations are based on open literature data with accuracy in the 20% to 30% range Appendix B also gives approximate methods to estimate laminar and transition flow heat transfer in tubes This information may be off by a factor of two When accurate laminar flow heat transfer information is needed, the HTRI ST simulation programs should be used Simple shell side heat transfer correlations for liquids and gases are included in Appendix B They apply to well proportioned shell and tube exchangers with turbulent flow only Extreme geometries or low shell side velocities require computer analysis The best conversion of pressure drop to heat transfer occurs with rotated square tube layout (45 degrees) in liquid service Inline square tube layout (90 degrees) is slightly better in gas service where Reynolds numbers are typically very high 214 Two-Phase Liquid/Gas Heat Transfer This section describes sensible heat transfer in two-phase hydroprocessing feed effluent heat exchangers, where phase change (boiling or condensing) is not controlling Two-phase heat exchangers are usually designed to operate in the annular or churn flow regimes on both shell and tube sides For these flow regimes, liquid coats the heat transfer surfaces and a continuous gas phase flows in the core Liquid droplets are entrained in the core flow Heat transfer resistance between the gas and the liquid is negligible and can be ignored The governing resistance is convection and conduction in the liquid film covering the heat transfer surface This type of flow can be modeled as a pseudo single-phase fluid using “no-slip” mixture density and December 1989 200-8 Chevron Corporation Heat Exchanger and Cooling Tower Manual 200 Design Background velocity, mixture heat capacity (or enthalpy), liquid viscosity and liquid thermal conductivity “No-slip” means that the liquid and gas are assumed to flow at the same velocity These pseudo single-phase fluid properties can be input to the HTRI single-phase simulation program or used with the simpler methods described in the previous section This approach has been validated against Company in-plant test data for well mixed flow only When horizontal two-phase heat exchangers operate in stratified flow regimes, liquid accumulates in the bottom of the exchanger If the outlet nozzle is on top, a stagnant liquid level rises in the exchanger as needed to force the upper part of the exchanger into a mixed phase flow regime where net liquid transport is possible For this situation, heat transfer varies as flow rate to the 1.6 power, and pressure drop is nearly independent of flow rate This is simply the effect of varying level of stagnant liquid If the outlet nozzle is on the bottom, vapor flows in the upper part and liquid flows in the lower part The fraction of the exchanger associated with each phase is more or less in proportion to their relative volume flow rates Stratified two-phase flow in exchangers results in very poor thermal performance and should be avoided Appendix C defines appropriate flow regime boundaries and gives the equations used to calculate pseudo single-phase properties 215 Boiling Pool boiling data are the foundation for correlating the performance of commercial boiling equipment A typical pool boiling curve, from the HTRI Design Manual, is shown in Figure 200-5 It applies to boiling water at atmospheric pressure This type of boiling is obtained when the heated surface is surrounded by a fluid that is not flowing Agitation is produced by natural convection currents and bubble motion The physical condition associated with various parts of the curve is illustrated in Figure 200-5 and briefly described below A-B Natural convection (no boiling) B-C Incipient boiling (surface boiling with subcooled bulk fluid) C-D Nucleate boiling (bulk fluid at saturation temperature) D-E Transition to film boiling (unstable) E-F-G Stable film boiling (heated surface is not wetted) Commercial boiling equipment is intended to operate in the incipient boiling or nucleate boiling regions Operation in the nonboiling A-B region or the D-E-F-G film boiling regions may result in severe fouling and/or mechanical failure of the equipment The pool boiling curves for single component fluids have been correlated in terms of critical pressure (Pc) and reduced pressure (Pr) and are given in Appendix D Chevron Corporation 200-9 December 1989 200 Design Background Heat Exchanger and Cooling Tower Manual Critical pressure is the pressure above which the distinction between liquid and vapor vanishes Reduced pressure is the ratio of operating pressure to critical pressure Boiling is not possible at or above the critical pressure The critical pressure of most hydrocarbons is between 400 and 600 psia Fig 200-5 Pool Boiling Curve for Water at Atmospheric Pressure The nucleate boiling heat flux for multicomponent mixtures is less than that of pure components at the same surface-to-bulk-temperature difference More volatile components boil first, concentrating less volatile components near the heat transfer surface This reduction in heat flux correlates with boiling range (dewpoint— bubble point) and critical temperature The maximum nucleate boiling heat flux for mixtures (point D in Figure 200-5) satisfactorily correlates to mixture critical pressure Boiling curves for horizontal tube bundles are markedly different than for single tubes Figure 200-6 shows an example from the HTRI Design Manual The bundle curve was obtained by boiling normal pentane in the shell at saturation temperature with a large excess of saturated steam in the tubes to minimize tube side resistance The Overall Temperature Difference in Figure 200-6 is the steam saturation temperature minus the n-pentane saturation temperature Nucleate boiling heat flux for bundles is much higher than for single tubes, but maximum bundle heat flux (dryout) is much lower Vapor generated by each tube enhances circulation and heat transfer for neighboring tubes at low heat flux Excessive vapor generation, however, prevents wetting some interior tubes above the incipient dryout heat flux Incipient dryout of central tubes usually occurs at 50% to December 1989 200-10 Chevron Corporation Heat Exchanger and Cooling Tower Manual Fig 200-6 200 Design Background Example: Comparison of Boiling Curves for a Single Tube and a Bundle, N-Pentane at 115 psia 70% of the maximum bundle heat flux Extensive dryout in the interior of the bundle exists at the maximum bundle heat flux Enhanced nucleate boiling heat flux, incipient dryout heat flux, and maximum bundle heat flux correlate with the ratio of heat transfer area to peripheral inflow/outflow area These correlations are given in Appendix D At or above incipient dryout heat flux, a fraction of the liquid entering the lower part of the bundle becomes completely vaporized in the interior Any solids in the liquid will deposit at this point Solids deposition in the bundle restricts flow, reduces dryout heat flux, and extends the deposition region Because most commercial streams contain a few parts per million solids, the usual result of operation above the incipient dryout heat flux is plugging of the bundle Operation of natural circulation boilers in a nonboiling region (A-B in Figure 200-5) results in a similar plugging problem Fluid shear is usually not adequate to keep trace amounts of solids in the liquid from accumulating and adhering to the heated surface and eventually stopping circulation Vigorous agitation associated with nucleate boiling is more than adequate to keep solids suspended The minimum heat flux to ensure nucleation is about 2000 Btu/hr ⋅ ft2 This may impose a turn-down limit, particularly for grossly oversized boilers Chevron Corporation 200-11 December 1989 200 Design Background Heat Exchanger and Cooling Tower Manual Reboilers should be designed reasonably close to but not exceeding incipient dryout heat flux This minimizes the size and cost of the reboiler, maximizes allowable turndown, and minimizes the required temperature of the heat medium Design considerations for reboilers are given in Section 360 216 Condensing This section discusses the principles of process condensation and gives rules-ofthumb for steam condensation Condensing surfaces are below the dewpoint temperature of the condensing fluid and are covered by a film of condensate The resistance to heat transfer between the vapor and the condensate is negligible for single component condensation Essentially all of the resistance to heat transfer is convection and conduction across the condensate film Major factors affecting condensate film thickness and heat transfer are whether the condensing fluid flows vertically downward or horizontally, whether the film is laminar or turbulent, whether the film is vapor shear controlled (high vapor velocity) or gravity controlled (low vapor velocity), and whether the condensing fluid is flowing in tubes or outside tube bundles The resistance to heat transfer between the vapor and the condensate film is significant for multicomponent condensation Least volatile components condense first, concentrating more volatile components near the vapor-liquid interface This tends to inhibit condensation This vapor phase resistance is governed by vapor shear and turbulent mixing in the vapor Total condensers usually operate in the vapor shear controlled turbulent film regime near the inlet, in gravity controlled regimes in the middle, and may have a liquid flooded zone near the outlet The heat transfer coefficients may vary from about a thousand at the inlet to about one Btu/hr ⋅ °F ⋅ ft2 at the outlet This type of condenser is usually analyzed by solving for the limits of the various flow regimes and applying the correlations appropriate to each regime This is an iterative process and involves a very large number of correlations Computer simulation with the HTRI CST program is recommended for these cases Some partial condensers operate with vapor shear controlled turbulent film condensation throughout, which is very similar to two-phase heat transfer discussed earlier The two-phase methods discussed earlier may be applied to condensers in this case Condensing steam is a common and efficient heat medium Steam condensing coefficients are usually between 2000 and 3000 Btu/hr ⋅ °F ⋅ ft2 in the vapor shear controlled regime and a few hundred in the gravity controlled regimes Most steam heated exchangers involve total steam condensation and have an outlet condensate pot to avoid a condensate flooded zone in the exchanger For these cases, a design steam side heat transfer coefficient of 1000 Btu/hr ⋅ °F ⋅ ft2 is recommended 220 Economic Pressure Drop and Velocity Pressure drop is required to change fluid momentum (increase velocity and/or change direction), and to overcome friction Friction pressure loss is the dominant December 1989 200-12 Chevron Corporation Heat Exchanger and Cooling Tower Manual 200 Design Background cause of pressure drop in most single- and two-phase heat exchangers Acceleration (momentum) loss is most important in boilers and is usually provided by differential static head Momentum and friction losses dominate in the inlet regions of condensers (and overall), with some vapor deceleration pressure recovery occurring in the bundle Both heat transfer and pressure drop increase with increasing fluid velocity Optimum pressure drop for single-phase heat exchangers and some two-phase heat exchangers (excluding boilers and condensers) is determined by balancing the cost of pressure drop (e.g., pumping power) against the cost of heat transfer surface area Inevitable pressure losses associated with nozzles, including entrance and exit zones of exchangers, not substantially affect heat transfer Exchanger nozzles usually match the piping size and pipe size is usually based on the same principle of balancing pipe costs against pumping costs These pressure losses are therefore a more-or-less constant fraction of the exchanger friction losses For pumped liquids, compressed vapors, or both (two-phase) in carbon steel exchangers, the economic expenditure of pumping power on each side of the exchanger (including nozzle, entrance, and exit losses) is about hp per 1000 ft2 of heat exchanger surface area The optimum is relatively flat Power expenditures between and hp/1000 ft2 have a significant effect on exchanger and pump or compressor costs, but result in approximately the same overall cost (5%) Designs anywhere in this range are reasonable Optimum power expenditure per unit area is approximately the same for any fluid density and tube size The above economic range of power expenditure can be expressed in terms of velocity and pressure gradient Economic velocity, ft/sec (independent of tube size): (28 to 36)/(Density, lb/ft3)1/3 Tube side ( to 10)/(Density, lb/ft3)1/3 Shell side The shell side velocity is for the “B” stream (crossflow) used in the HTRI programs and is shown in Figure 200-3 “B” stream flow is usually about 70% of the total shell side flow Typical values for low viscosity hydrocarbon liquids and water are: to ft/sec Tube side to ft/sec Shell side Economic pressure gradient, psi/ft (for 3/4-inch, 13 BWG tubes): (0.05 to 0.08) ⋅(Density, lb/ft3)1/3 Tube side (0.1 to 0.16) ⋅(Density, lb/ft3)1/3 Shell side The tube side flow path is the straight tube length times the number of tube passes The shell side flow path is the axial shell length times the number of shell passes Typical values for low viscosity hydrocarbon liquids and water are: Chevron Corporation 200-13 December 1989 200 Design Background Heat Exchanger and Cooling Tower Manual 0.2 to 0.3 psi/ft Tube side 0.4 to 0.6 psi/axial ft Shell side The factor of two difference between shell and tube side pressure gradients simply reflects the fact that actual shell side flow path is about twice the axial length Economic pressure gradient for 1-inch tubes is about 70% of that for 3/4-inch tubes These simple rules-of-thumb are insensitive to wide variations in energy costs, because exchanger costs are energy intensive and track well with energy costs Exchangers with similar tube side and shell side heat transfer coefficients (the norm) should be designed to the economic parameters particular to each side If one side of the exchanger limits (much lower coefficient than the other side), the above economic parameters should be used for the limiting side, and less power expended on the side that does not limit Power expenditures on the high side, or more, are justified for expensive alloy exchangers In some cases (e.g., desalter effluent water coolers), available pressure drop is dictated by other process requirements (e.g., suppressing vaporization in the desalter) and should be used to the maximum practical extent to reduce exchanger size These guidelines not apply to reboilers and condensers Reboilers are usually driven by natural circulation rather than pumps Pressure drop in condensers impacts column and reboiler design as well pumping/compression costs At atmospheric pressure or above, condenser pressure drop is typically the smaller of about psi or 10% of the absolute pressure At very low pressures, the type and performance of vacuum equipment governs condenser pressure drop The economics of pumping/compression costs versus exchanger cost is one aspect of optimum exchanger utilization Equally important is balancing the value of heat exchange against the cost of achieving it (ie., determining the appropriate duty of the exchanger) Determination of exchanger duty is beyond the scope of this manual 230 Flow Splitting Flow splitting is sometimes a problem in networks and in multiple bundle services with parallel branches Flow splitting can always be controlled with valves, but that wastes pumping power and is undesirable Flow splitting in treated cooling tower water systems has been a problem due to oversizing some condensers and throttling water flow for control, or designing different services with different numbers of bundles in series and with large differences in design pressure drop Throttling treated cooling tower water is always undesirable, as discussed in Section 310 All water cooler services in a particular circulating system should be designed with the same overall pressure drop and unrestricted water flow whenever possible December 1989 200-14 Chevron Corporation Heat Exchanger and Cooling Tower Manual 200 Design Background Splitting two-phase streams has been a problem when the piping is not symmetrical, even though care was taken to balance pressure drops in parallel paths Reliable two-phase flow splitting is accomplished by splitting both branches along the same axis and at right angles to the original combined stream in a standard pipe tee, and designing for equal pressure drop in each branch 240 Fouling This section discusses principal fouling mechanisms, and services and conditions where they may occur In most cases, fouling conditions can be avoided by appropriate process and exchanger design The principal fouling mechanisms are: • • • • • • Particulate fouling Salt precipitation fouling Chemical reaction fouling Filming amine fouling Biological fouling Corrosion fouling Particulate fouling is possible for streams that contain a few parts per billion of solids Particle sizes between 0.001 and micron contribute most to the deposit Particles over about 100 microns usually erode smaller particles and inhibit fouling Simple particulate deposits are weak and yield readily to fluid shear Exchangers that operate with fluid shear stresses greater than about 0.001 psi are usually not subject to simple particulate fouling Economic liquid velocities discussed in Section 220 result in nearly four times the shear stress (or twice the velocity) needed to prevent simple particulate fouling Economic velocities for most two-phase exchangers are also above the threshold for particulate fouling, except at high vapor fraction Economic gas velocities are not adequate to keep small particles moving For example, in FCC flue gas coolers the process and/or exchanger design must be adjusted to control gas side fouling This example is discussed in Section 390 Salt precipitation fouling usually involves liquid-to-solid phase transition at the heat transfer surface It occurs where an aqueous phase contacts the heat transfer surface and the aqueous phase is supersaturated with respect to one or more of the dissolved salts Most salt deposits can not be eroded at economic velocities A few salt deposits are erodible at economic liquid velocities, and these are candidates for aqueous phase corrosion inhibitors (See Section 310) Salt precipitation fouling may occur alone or in combination with particulate fouling Gas-to-solid phase transition (sublimation) is a less common type of salt precipitation that can occur in overhead condensers and effluent streams in hydroprocessing plants NH4Cl is usually the salt involved Addition of an aqueous phase and/or control of its composition is the normal method of eliminating salt precipitation conditions in heat exchangers This method Chevron Corporation 200-15 December 1989 200 Design Background Heat Exchanger and Cooling Tower Manual is practiced in cooling tower water exchangers, sea water coolers, crude oil preheat exchangers, crude unit atmospheric overhead condensers and some other services Chemical reaction fouling involves chemical bonding between thermally unstable organic compounds to the extent that liquid-to-solid phase transition occurs at the heat transfer surface This is a slow process in liquids and usually does not cause fouling at economic velocities in the absence of particulate matter However, chemical reaction fouling can occur in combination with particulate fouling and affect fouling by increasing the strength of predominantly particulate deposits Chemical reaction fouling increases exponentially with temperature above a certain threshold temperature It is controlled by maintaining heat transfer surface temperatures below the threshold The threshold fouling temperature for naturally occurring hydrocarbons is about 600°F for the heaviest components and higher for lighter components Residuum fouling is the most common type of chemical reaction fouling in refineries Filming amine fouling usually occurs with liquids in combination with particulate fouling and affects fouling by increasing the strength of predominantly particulate deposits Economic liquid velocities will not stop filming amine fouling Filming amines can cause fouling in two other ways Filming amines dissolved in light hydrocarbon will “deposit” as a liquid (like tar) where the light hydrocarbon is evaporated to extinction Most filming amines also decompose to solids above a certain temperature, usually between 300°F and 500°F Filming amines are marketed as corrosion inhibitors, dispersants, antioxidants, metal deactivators and antifoulants They are commonly injected into crude unit atmospheric overhead systems and dispersed throughout the refinery Biological fouling occurs spontaneously in oxygenated waters between 32°F and 120°F, unless significant toxic material is present Cooling tower water systems and sea water cooling systems are subject to biological fouling, usually in combination with particulate fouling Chlorination is the most common method of biological fouling control Redundant equipment and/or frequent cleaning is an alternative Corrosion fouling involves irregular loss of metal and accumulation of corrosion products Increased surface roughness improves convective heat transfer and compensates for added thermal resistance of corrosion products that remain on the surface Pressure drop increases initially due to the roughness but may decrease if metal loss is significant These effects are usually small in heat exchangers and are usually ignored in design Quantitative fouling information is given in Section 300 250 Overdesign Typical Company practice is to rate plants and most component equipment at 80% of their calculated capacity This rating allows for normal uncertainty in design, repairable deterioration during one turn-around cycle, and a margin for control That is, most plants could ideally run at 125% of rated capacity initially and should be able to run at 100% of rated capacity at planned end of run December 1989 200-16 Chevron Corporation Heat Exchanger and Cooling Tower Manual 200 Design Background Heat exchanger thermal and hydraulic overcapacity should be comparable to that of other equipment in the plant More overcapacity may be justified for unconventional processes with little commercial experience Additional heat exchanger surface area is needed in services where fouling is intended (corrosion inhibitor films) or is allowed to occur Corrosion inhibitor fouling is asymptotic (self-limiting) at temperatures below the thermal stability limits of the protective films Affected exchangers should be designed for 80% of the asymptotically fouled performance (design U = 0.8 × fouled U) Asymptotic fouling resistance and film stability limits are given in Section 310 for treated cooling tower water and in Section 380 for filming amine corrosion inhibitors Most fouling situations are not asymptotic (fouling resistance increases with time indefinitely) Redundant (spare) heat exchangers with isolation block valves and frequent cleaning are practical methods of accommodating nonasymptotic fouling Although 50% to 100% excess area for both the operating exchanger and its spare with cleaning every month or two is usually effective, it may not be economical This approach has been used for some crude preheat systems, and to heat large quantities of recovered oil with thermally unstable components Fouling prevention is usually more economical than fouling accommodation Sections 330, 350, 360 and 380 discuss fouling prevention for closed loop water cooling, high temperature steam generators, reboilers, and crude unit heat exchangers, respectively Biological fouling from sea water is nonasymptotic but is an order of magnitude slower than other types of fouling Both fouling prevention and fouling accommodation are common in sea water exchangers Section 320 discusses biological fouling rates of untreated sea water, the conditions when it occurs, and other limitations Although most commercial streams will foul under some conditions, most exchangers are sized for clean operation only when the expected range of operation conditions is outside the range where fouling occurs Section 250 discusses principal fouling mechanisms, and the services and conditions where they may occur The exchanger designer normally only specifies the calculated clean pressure drop Hydraulic overdesign is usually specified by the pump or compressor designer and is usually 125% of design flow (150% of clean pressure drop), plus a 30% control allowance This factor is usually more than adequate to accommodate pressure drop increase due to fouling The exchanger designer should check all shell and tube exchangers for tube vibration at 125% of design flow, as discussed in Section 260 below 260 Tube Vibration This section discusses tube vibration mechanisms, tube failure locations, and design criteria to prevent failures Damaging heat exchanger tube vibration mechanisms are vortex shedding and fluid elastic whirling Acoustic resonance can occur in exchangers but does not damage tubes and is not covered here Chevron Corporation 200-17 December 1989 200 Design Background Heat Exchanger and Cooling Tower Manual Vortex shedding downstream of a single tube in cross flow is illustrated in Figure 200-7, and is a “snap shot” of a vortex forming near the tube and three other vortices that were shed earlier in sequence Fig 200-7 Illustration of Vortex Shedding from a Single Tube Vortices are alternately shed on each side of the tube and exert an alternating pressure (force) in the direction perpendicular to the crossflow direction This alternating pressure makes the upstream flow first favor one side of the tube and then the other The vortex shedding phenomenon for a single tube occurs in heat exchanger bundles at peripheral tubes and along unblocked pass partition lanes in the vicinity of inlet and outlet nozzles, but does not penetrate far into the bundle The location of the next row of tubes affects the frequency of formation of vortices on the first row Interaction between vortices from adjacent tubes usually degenerates the vortices to harmless random noise after about the third row into the bundle That is why only peripheral tubes are of concern Vortex shedding frequency depends on tube layout angle The distinction between 30 degrees and 60 degrees, and between 45 degrees and 90 degrees, however, is academic because of the extreme divergence or convergence of nozzle flows The 60-degree layout angle governs for all triangular layouts; the 45-degree layout angle governs for all square layouts Damaging tube vibration occurs when the vortex shedding frequency matches one of the tube natural frequencies The maximum allowable unsupported tube span is set so that the highest anticipated crossflow velocity will not excite the first mode natural frequency of the tube Higher mode resonant vibration occurs at higher velocities Figure in Standard Drawing GC-E1048 defines maximum unsupported tube spans for inlet and outlet regions of shell and tube exchangers Vibration control involves adding partial support near nozzles as needed These support plates not affect thermal or hydraulic performance December 1989 200-18 Chevron Corporation Heat Exchanger and Cooling Tower Manual 200 Design Background Fluid elastic whirling may occur in the first two rows beyond the baffle cut in the interior of the bundle only These are the spans labeled “L4" in Figure in Standard Drawing GC-E1048 Clusters of at least three tubes in at least two different rows vibrate in harmony, as illustrated in Figure 200-8 Fig 200-8 Fluid Elastic Whirling Tube Motion Figure in Standard Drawing GC-E1048 defines maximum unsupported tube spans for interior tubes The velocities that initiate fluid elastic whirling are greater than economic velocities and therefore rarely affect exchanger design Problems occur when abnormally high velocities and/or abnormally long unsupported spans are used 270 Enhanced Surfaces Enhanced surfaces are justified when one side of a heat exchanger has an inherently lower heat transfer coefficient than the other side (e.g., low gas side coefficient and high liquid coefficient) Air cooled exchangers are the most common example See Section 600 Water cooled compressor inter- and after-coolers are also appropriately designed with enhanced surfaces Low fin tubes with area ratios between three and five to one are commonly used Optimum fin area depends on the gas pressure The HTRI ST-5 program can be used to select the most appropriate type of tube There are few other appropriate exchanger applications of enhanced surfaces in the petroleum and petrochemical business 280 Computer Program Abstracts The following HTRI heat exchanger simulation computer programs are available on the Company’s mainframe computer system ST-5—Single-phase shell and tube heat exchangers This program is primarily used to model single-phase vapor and liquid It is also used for two-phase mixtures modeled as pseudo single-phase fluids, described in Chevron Corporation 200-19 December 1989 200 Design Background Heat Exchanger and Cooling Tower Manual Section 214 This program should also be used when the limiting side is singlephase even though boiling or condensing occurs on the other side Steam condensers and steam generators are usually modeled on ST-5 CST-2—Condensing shell and tube heat exchangers This program is used to model multicomponent condensation in horizontal and vertical down-flow shell and tube exchangers The coolant may be single-phase or boiling RTF-2—In-tube boiling shell and tube exchangers This program is used to model vertical thermosiphon reboilers, and horizontal or vertical forced flow reboilers RKH-2—Boiling outside horizontal bundles This program is used to model boiling in horizontal bundles in enlarged shells (kettle reboilers) or without shells (internal column reboilers) The program does a simplified analysis of flow in inlet and outlet piping It uses average fluid properties and does not calculate circulation rigorously The program should be used with caution ACE-2—Air cooler heat exchangers This program is used to model horizontal air coolers with single-phase or condensing fluids in the tubes Instructions for access and use of these computer programs are given in the Heat Exchanger Design Program Users Guide maintained separately December 1989 200-20 Chevron Corporation ... The exchanger designer normally only specifies the calculated clean pressure drop Hydraulic overdesign is usually specified by the pump or compressor designer and is usually 125% of design flow... should be designed with the same overall pressure drop and unrestricted water flow whenever possible December 1989 200-14 Chevron Corporation Heat Exchanger and Cooling Tower Manual 200 Design Background. .. oversized boilers Chevron Corporation 200-11 December 1989 200 Design Background Heat Exchanger and Cooling Tower Manual Reboilers should be designed reasonably close to but not exceeding incipient