CHEVRON - Shell and Tube Exchanger Component Design Considerations

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500 Shell and Tube Exchanger Component Design Considerations

AbstractThis section discusses the mechanical design of shell and tube heat exchangertheir components. Emphasis is placed on company practices which differ from industry standards.

Contents Page

510 Design Pressure and Temperature 500-3

511 Design Pressure

512 Design Temperature

513 Relief Valves

514 Rupture Surge Pressure

520 Bundle Design 500-5

521 Tubesheet Design

522 Tube-to-Tubesheet Connection

523 Longitudinal Shell Baffles

524 Impingement Devices

525 Retrofitting Floating Head Bundles with U-tubes

530 Channel and Shell Design 500-14

531 General

532 Body Flanges

533 High Pressure Closures

534 Connections

540 Gaskets 500-35

Chevron Corporation 500-1 March 1994

500 Shell and Tube Exchanger Component Design Considerations Heat Exchanger and Cooling Tower Manual

550 Insulation 500-41

551 Reasons for Insulating

552 Types of Insulation

553 Weatherjacketing

554 Flange Insulation

March 1994 500-2 Chevron Corporation

Heat Exchanger and Cooling Tower Manual 500 Shell and Tube Exchanger Component Design Considerations

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510 Design Pressure and Temperature

511 Design Pressure

Internal PressureThe maximum allowable working pressure (MAWP) should normally exceed themaximum expected operating pressure as follows:

The maximum expected operating pressure is the maximum expected pressureinside the heat exchanger under any operating case, startup, or shutdown cond

• All exchangers with liquid or vapor-liquid mixtures on the low pressure sideshould be designed for tube rupture safety. See Section 514 and Appendixfor this procedure.

• In case of large vertical exchangers, the nameplate design pressure is the maximum pressure permissible at the top of the exchanger. Therefore, despressure must be adjusted for any difference in static head that may exist between the part considered and top of vessel.

External PressureExchangers operating at less than atmospheric pressure should be designed foexternal pressure (vacuum) of 15 psig. All exchangers designed for internal presure should also be adequate for at least 7.5 psi external pressure at 450°F when the ratio of D/T exceeds 150. (D = shell O. D., and T = shell thickness excluding cosion allowance.)

512 Design TemperatureThe design temperature for any part of a heat exchanger is the maximum allowoperating temperature the of fluid inside that part (or minimum for cold service design). The following are recommended:

1. The hot service design temperature (-20°F and above) for each side of a unit should be at least 25°F (14°C) above the maximum operating temperature for

Maximum Expected Operating Pressure, psig

Minimum Amount by Which MAWP Exceeds the Maximum Expected Operating Pressure

0–170 25 psi

170–300 15% of max. op. press.

300–450 45 psi

450–1000 10% of max. op. press.

1000 + Not less than 8% of max. op. press.



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the fluid on that side. (Note that tubes are exposed both to shell side and tuside fluids.)

2. The cold service design temperature (below -20°F) for each side of a unit should be at least 5°F (3°C) below the minimum expected operating tempera-ture of the fluid on that side.

The maximum design temperature that is on the name plate of the heat exchangis the temperature at which the ASME Code allowable stress for the componendetermined, and must be above the maximum expected operating temperatureNormal operating temperature is only occasionally related to design temperaturFor example, tubes exposed to treated cooling tower water, irrespective of metalurgy, plug solid if tube surface temperatures exceed about 160°F. The intent of the exchanger design and control system is to maintain temperatures within the funtional range. However, the name plate design temperature is usually the highestemperature at which the specific material maintains its maximum allowable strFor carbon steels, this temperature is 650°F.

Minimum pressurizing temperature, an important design parameter, is discussein detail in the Pressure Vessel Manual. In brief, the reason for establishing a minimum pressurizing temperature is to avoid a brittle fracture. Ordinary carbonsteels, for example, become brittle at low temperatures. The ductile-to-brittle tration temperature may range from well above ambient to well below ambient, depending on grade of steel used, and its thickness. The aim is to choose a mawhich will not suffer brittle fracture under the design operating conditions of theexchanger. This includes hydrotest, which should be done at a temperature abothe minimum pressurizing temperature.

513 Relief Valves

Pressure ReliefThe ASME Code requires that all pressure vessels be provided with protection against overpressure by use of pressure relief devices. The protective devices nnot be directly on the pressure vessel when the source of pressure is external tpressure vessel and the piping does not include any valves between the relief dand the vessel.

Consequently, many heat exchangers do not have pressure relief valves directlthe vessel, but are rather part of an overall hydraulic system which does have ption from overpressure. In many cases, the source of pressure is a pump or compressor external to the exchanger.

The Instrumentation and Control Manual discusses relief sizing in more detail.



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Thermal ReliefThermal relief valves (TRVs) are required by the ASME Code by the following conditions:

1. Either shell or tube side component can be overpressured by heat input frothe other side, and

2. The component can be isolated from the main pressure safety valve (PSV)valves other than PSV maintenance valves, or

3. The component is not protected by a PSV.

To minimize the possibility of a TRV releasing during a PSV relief, the TRV canset at 110% of design pressure as allowed by ASME Code. One thermal relief vcan serve as the protective device for multiple exchangers in series if there areblock valves between them.

514 Rupture Surge PressureAll exchangers with liquid or vapor-liquid mixtures on the low pressure side shobe designed for tube rupture safety. This is accomplished by setting the designsure on the low pressure side equal to the maximum normal operating pressurethe initial surge pressure due to the complete break of one tube. Long term (e.g+ seconds) pressure transients should be prevented with relief devices in the pTubesheets, shells, shell covers, and channels should be designed to this surgsure. Body flanges should also meet ASME Code requirements but not leak tigness requirements at this design pressure.

Tube rupture is particularly a problem in high pressure gas/low pressure coolingwater applications. Appendix F gives a detailed procedure and examples for demining rupture surge pressure and rupture flow rate.

520 Bundle Design

521 Tubesheet DesignThis section covers the applicable codes and industry practices for establishingtubesheet design and tubesheet thicknesses. Tubesheets separate the shell sidtube side fluids and provide the anchor point for tube ends. TEMA standard rulefor calculating tubesheet thickness are used in the industry extensively.

TEMATEMA covers procedures to establish tubesheet thickness for U-tube bundles, floating head bundles, and fixed tubesheet construction.

ASME Code, Section VIII, Division 1ASME Code, Section VIII, Division 1, Appendix AA, covers tubesheet thicknesscalculations for U-tubes of various configuration.



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ASME Code, Section VIII, Division 1, Appendix A, tells how to calculate allow-able loads for tube-to-tubesheet joints. This calculation may have an effect on tubesheet thickness, the method of joining tube to tubesheet, or the need to proan expansion joint in the shell of an exchanger. Appendices A and AA of SectioVIII are both nonmandatory and therefore do not have to be followed by a vendunless required by the Company.

Waste Heat Boiler–Fixed Tubesheet Exchanger TypeWaste heat boiler tubesheets are designed in accordance with the ASME Boiler and Pressure Vessel Code, Section I, Paragraphs PG-49.1, PW-19.1 and PFT-27, whicaccount for staying capacity of the tubes. Tubesheet thickness is governed by tlargest unstayed area, which is usually the annular space between the bundle athe shell.

Tubesheets designed by TEMA rules would be much thicker and are unacceptafor high temperature steam generators because of high thermal stresses.

Tubesheet Thicknesses and Tolerances: TEMA and Chevron Practice• It is Company practice to use TEMA, Paragraph F-2, tolerances for thickne

and API 660, Paragraph 7.8, for flatness tolerance on new tubesheets, alththis is generally not a problem.

• Some Company locations add a “maintenance” allowance (usually 1/8 inchonto the channel side tubesheet thicknesses beyond TEMA minimum requment to compensate for any surface repairs required due to maintenance aties.

Clad TubesheetsFor clad tubesheets with rolled tube-to-tubesheet joints, the nominal cladding thness should be 1/2 inch minimum, and one of the grooves or serrations in eachhole should be completely within the cladding. The cladding thickness may be lfor welded tube-to-tubesheet joints.

Roll-clad is the preferred method of cladding or overlay. However, explosion clading is sometimes used, especially for small pieces like tubesheets where roll cding is not economical. For other requirements on cladding, refer to EXH-MS-2583, included in this manual.

Bundle Pull Hole DesignRemovable bundles which are 20 inches or more in diameter should have four tapped holes in the channel side of the stationary tubesheet for bundle pulling heads. The holes should be symmetric about the bundle centerline and locatedtube positions between 3 7/8 inches and 5 3/4 inches from both horizontal and vertical centerlines. Pull hole size and thread engagement should take into acctubesheet material and be designed for a maximum pulling load equal to twice bundle weight. The threads should be National Course Series below 1 inch andeight-pitch series for 1 inch and above.



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Bundles smaller than 20 inches in diameter may have two tapped holes for pulleyes. Small pre-engineered exchangers are usually supplied without pulling ho

These guidelines may have to be modified or waived for special construction, sas for thin clad tubesheets. Pull holes should be protected in service by threadeplugs.

522 Tube-to-Tubesheet ConnectionThe main function of tube-to-tubesheet joint is to seal the tubes tightly to the tubesheet, and for some exchangers, an additional function is to support the tubesheet against pressure induced load. Tubes are sealed inside the tubesheethe following methods.

• Expanding tube inside tubehole• Welding tubes to tubesheet

Expanding Tubes Inside TubeholesExpanded tube-to-tubesheet joints are industry standard. In this case tubes areexpanded inside tubeholes by such methods as rolling or applying hydraulic presure directly to the tube end. Properly rolled joints have uniform tightness to mimize tube fractures, stress corrosion, tubesheet ligament enlargement, and disof the tubesheet. Rolling to 95% of tubesheet thickness is recommended. Rollinor beyond the tubesheet thickness is not recommended—for it may damage thetubes.

For moderate general process requirements (less than 300 psi and less than 35°F) tubesheet holes without grooves are standard. For all other services with expantubes at least two grooves are machined (1/8 inch wide by 1/64 inch deep) in etube hole. See Figure 500-1.

Fig. 500-1 Rolled Tube - Tubesheet Connection



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Expanding the tubes into the grooved tube holes provides a stronger joint but results in greater difficulties during tube removal.

The following steps must be taken when tubes are rolled inside tube holes:

1. Tubes should be expanded to provide an initial contact of the tube to the tuhole.

2. Tubes should be seal welded if required. (See the seal welding procedure below.)

3. Tubes should be given final roll. A reduction in wall thickness of 5% is somtimes used as an indicator of adequate rolling.

4. Hydrotest the shell side after the final rolling.

Welding the Tubes to TubesheetAdditional tightness beyond that of the tube rolling is sometimes required in thefollowing areas:

• Steam generators when design pressure is greater that 450 psi• Boiler feedwater heaters• Feed/effluent heat exchangers in hydroprocessing plants• Any exchangers where cross-contamination must be scrupulously avoided

In these cases, tubes can be rolled and then seal or strength welded to the tubSeal welding is defined as a very small bead of weld around the tubes where ncredit can be taken for strength of that weld for calculation of tube-to-tubesheetjoint load. Figure A-2 of Appendix A of ASME, Section VIII, Division 1, shows some acceptable strength weld geometries.

Cleanliness in seal welding is of the utmost importance and care must be exercduring all steps of assembly not to contaminate cleaned parts. A chronic probleespecially in sour services is contamination of the weld with sulfur or iron sulfidcoming from a dirty tubesheet face or tube hole. This contamination makes it impossible to make a leak-free weld. The following procedure summarizes the requirements for seal welding heat exchanger tubes to tubesheets.

1. Use new tubesheets if possible. If old tubesheets are used, make as muchnew as possible. After machining, degrease by steam cleaning.

2. Clean tube ends with tube polisher.

3. Clean tubes (full length), tubesheets, and bundle carcass by immersion in aalkaline detergent solution.

4. Rinse cleaned parts with hot water and inspect.

5. With carcass in horizontal position, place all tubes.

6. Give tubes a light roll.



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7. Adjust tubes for 1/32 inch to 1/8 inch projection, and tack weld each tube tothe front tubesheet.

8. Trim other end of tubes to same extension and tack about one-quarter of thtubes.

9. Turn tube bundle in horizontal position so tube ends are in vertical rows anweld with MIG short-arc.

10. Clean and dye-penetrant-inspect all welds. Repair as required.

11. Reposition bundle and complete opposite end as required above. If desiredboth ends may be welded at once.

12. Give tubes full roll.

13. Place bundle in shell and test. Repair as required and repeat dye-checkingations.

Note It is important that a specific weld procedure be developed for the work athe actual materials used. The shop doing the work must demonstrate qualificato use this procedure. Consult with a local welding specialist or the Material Divsion Welding Specialist for help in developing the weld procedure.

523 Longitudinal Shell BafflesIn the design of heat exchangers, it is sometimes advantageous to use a TEMAType “F” (two-pass shell), “G” (split-flow shell), or “H” (double-split-flow shell). All of these require a longitudinal baffle to control the shell side flow. To preventbypassing, the seal between the longitudinal baffle and the shell is most comma “Lamiflex” type. More recently, Richmond Refinery has been using with good results Thermo-Ceram fabric for seal between long baffle and shell joint on services such as water or lube oil.

Longitudinal Baffle ThicknessThe Company recommends that the longitudinal baffle thickness be the largestthese three: (1) 1/4 inch, or (2) the thickness required by TEMA, for transverse baffles, or (3) the thickness for differential pressure loading.

Attachment to TubesheetThe longitudinal baffle should be fillet welded to the tubesheet.

Lamiflex BaffleThe stack of flexible strips is most commonly attached to the edge of the longitudinal baffle by sandwiching them between the longitudinal baffle and a bolting sas shown in Figure 500-2.

A typical seal consists of a long stack of eight strips, each 0.004 inch thick. Thisthickness represents a compromise: A thin strip is fragile and vulnerable to mecical damage when the tube bundle is handled and inserted into the shell, and is



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susceptible to damage by corrosion. However, only thin strips can flex adequatto seal.

The most common material for the strips is Type 304 stainless steel. But other rials could also be used depending on process requirements (such as hydroprocessing systems).

The angle of contact between shell and flexible strips should be small so that frtion during installation is minimized and the differential pressure has the greateeffect in causing the strips to seal. To this end, it is recommended that dimensio“A” and “B” in Figure 500-3 should be about equal, with both in the range of 1/23/4 inch.

Fig. 500-2 Lamiflex Baffle

Fig. 500-3 Lamiflex Baffle Dimensions



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Protection during installation. The lamiflex baffle must be protected with crib-bing to avoid damage during rigging operations.

Fiber Fabric—“Thermo Ceram” BaffleSince 1985 Richmond Refinery has been using, on existing units that have edgseal problems, a refractory textile product woven from white ceramic (alumina silica) fibers. The results have been successful in water or lube oil systems servThe Company has not yet used it in other services. Some of the advantages offabric “Thermo Ceram” over flexible strips are:

• The fiber fabric is less prone to being damaged when the bundle is removefrom the shell. Lamiflex baffles normally have to be replaced when the bundis removed.

• Thermo Ceram has no sharp edges to cut personnel or crane slings.

• The fiber fabric conforms closely to shell irregularities.

• It is very economical.

• It is made of nonasbestos fabric, good to 2200°F.

A method of attaching the fiber fabric is shown on Figure 500-4.

(Thermo Ceram can be obtained from Allied Packing in Oakland, California, telphone 654-3274).

Fig. 500-4 Ceramic Fiber-Type Seal



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Transverse Baffles. The transverse baffles must be notched to provide clearancefor the fiber fabric seal (See Figure 500-4). The clearance area at this notch shobe minimized, for it adds to the leakage through the transverse baffle.

524 Impingement DevicesThis section compares Chevron criteria against TEMA Recommended Guidelinfor impingement devices. It also discusses the types of impingement devices usin the Company.

The tubes directly underneath the shell inlet nozzle may need to be protected against impinging fluid. Lack of proper impingement devices can cause tube faiby corrosion, erosion, or vibration. However, use of an impingement device wheis not needed increases exchanger diameter and cost.

Chevron PracticesTEMA recommends impingement plates for most services. Impingement plateshave been a chronic cause of both erosion and vibration problems. Removing impingement plates has been a common solution.

Chevron’s normal practice is to put two staggered rows of impinging rods in theprojection of the inlet nozzle to serve as an impingement device and also to distribute flow in the bundle. The impingement rods are recommended for all exchangers (regardless of service) where shell diameter is 20 inches or larger. Impingement devices are not practical in small exchangers (shell diameter less20 inches) and are usually not provided.

Impingement RodsImpingement rods are preferred to an impingement plate for several reasons. Fthe plate creates a dead space directly beneath it, lowering the heat transfer in tubes. Also, if the plate blocks too much of the inlet area, then the fluid may accerate into the remaining gap causing serious erosion of the tubes in that area.

Designing the rods is recommended as follows:

• The rods should consist of 1/2 inch solid rod inside 3/4 inch tube spacers which are the same diameter as the active tubes.

• The two rows of rods replace the first two tube rows which extend past the nozzle projection.

• The distance between the center-lines of the outermost rods in the first rowat least equal to the inside diameter of the shell inlet nozzle.

• The effective length of the rods is at least 20 percent greater than the diamof the shell inlet nozzle. The actual length of the rods may extend beyond teffective length as required for construction.

• For staggered tube layouts (30° and 45°), the impingement rods should be of the same layout as the active tubes. For inline tube layout (90°), the impinge-ment rods should have a 45° staggered layout.



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TEMA GuidelineThe TEMA Standard provides a minimum guideline for determining when an impingement device should be used. This guideline is appropriate for Companyalso. Impingement protection underneath the shell inlet nozzle is recommendedthe following:

• All noncorrosive, nonabrasive, single phase fluids with ρV2 >1500.

• All other liquids, including liquids at their boiling point ρV2 >500.

• All gases and vapors, including all nominally saturated vapors, and for liquid/vapor mixtures

The TEMA Standard also recommends that in no case should the shell or bundentrance or exit area produce a value of ρV2 in excess of 4000.

“V” is the linear velocity of the fluid in feet per second and “ρ” is its density in pounds per cubic foot.

Other Types of Impingement Devices

Impingement Plate. A circular or rectangular plate is placed directly underneaththe inlet nozzle perpendicular to nozzle flow. This plate could be welded to the shell, bolted to clips which are welded to the shell, or bolted to baffles on eitherside of shell inlet nozzle. The preferred construction is to attach the plate to thebundle.

If an impingement plate is used, it must be at least 1/4 inch thick and extendedminimum of 1 inch (or 10% nozzle diameter, whichever is greater) on each sidethe projected nozzle bore. Also the flow area off the impingement plate should more than the inlet nozzle flow area. Impingement plates, however, are not recomended because of the problems stated above.

Distribution Belt. A distribution belt consists of a collar that fits around the shellthe inlet and/or the outlet. The shell nozzle attaches to this collar. The fluid entethrough the nozzle and flows through the annulus between the belt and the sheThe fluid enters the tube bundle through windows cut in the shell, with a reducevelocity. Distribution belts are not widely used in the Company. They are expenand have maintenance problems.

525 Retrofitting Floating Head Bundles with U-tubesIt is often advantageous to change a floating head bundle to a U-tube bundle. Tchange may be warranted because of excessive leaks between floating head fland the tubesheet. Recent progress in U-tube bundle cleaning methods allows of U-tubes in many more services.



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Retrofitting a floating head to a U-tube bundle requires thermal, hydraulic, and vibration redesign. Once the need for retrofitting has been established, the following steps can be taken.

1. Obtain existing exchanger data sheet and fabrication drawings.

2. Put the following data on a new exchanger data sheet:

– Performance requirement of the new exchanger. This could be either thexisting exchanger performance requirement or new data as specified the process engineer based on information from the field about the opetion and fouling of the old exchanger. In case of split ring type floating head where possibly less heat transfer area will be available, the re-evation of performance data may be required.

– Existing exchanger’s shell and channel inside diameters– Locations and sizes of shell inlet and outlet nozzles– Maximum allowable length of bundle. Allow minimum of 2 inch clear-

ance between end of U-bends and inside of rear shell cover.– Location and thickness of existing channel pass partition plates– Material of construction for the bundle– Tubesheet thicknesses– Tube sizes, pitch, and layout preference– Baffles type, cut, and spacing preference– Impingement device requirement

3. Note that all the above data are subject to re-evaluation for the new bundleonly criterion is that the new bundle must fit in the existing shell, rear shell, and channel.

4. Design a U-tube bundle based on the new data sheet. This can be done byusing the Company/HTRI Programs or by using an exchanger design contractor.

– Compare cost of retrofit to cost of new exchanger: extensive modificatito channel or shell may justify purchase of a complete new exchanger.

– Consider the possibility that it may be necessary to remove the channepass partition plates on the existing unit and install new ones. This is nconsidered extensive modification.

– Consider the effects of excessive vibration and its prevention (see Standard Drawing GC-E1048).

530 Channel and Shell DesignThis section covers mechanical design of the channel and shell on a shell and exchanger. Refer to EXH-MS-2583 for more details on channel and shell consttion.



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531 General

Channel and Shell Thickness• The channel and shell contain the two separated fluids in the exchanger. T

are almost always cylindrical in shape and follow rules and regulations of ASME Code for structural integrity. ASME Code, Section VIII, establishes minimum metal thickness of cylindrical channels or shells.

Stacking Restrictions• Avoid stacking more than three exchangers. More than three can cause ma

nance, handling, and shipping difficulties.

• Piping and shell stresses in stacked exchangers should be within acceptablimits.

• The lower shells of stacked removable-bundle heat exchangers should be designed to withstand the superimposed loads due to exchanger operatingweight or bundle pull-out, without suffering distortion that could cause bindiof the tube bundles.

532 Body FlangesBody flanges are used to permit disassembly and removal or cleaning of internparts of a heat exchanger. Integral flanges (hub or weld neck) are flanges that aintegral with the exchanger wall or neck. This type of flange is recommended oservices and pressures except in water service for pressure up to 150 pounds. flanges (slip on) should be reviewed by a specialist. See Figure 500-5.

For pressures over 1,000 psig, special closures should be considered, such asgral construction (no flanges), welded diaphragm seals, or breech lock closuresWelded diaphragm and breech lock closures are discussed in Section 533.

The ASME Boiler and Pressure Vessel Code establishes the minimum requiremfor a flange design and provides a method of calculation (Section VIII, Division Mandatory Appendix 2 and Non-mandatory Appendix S). Deficiencies in ASMECode designed flanges, from a leakage standpoint, have been recognized for stime. Although records are not routinely kept, a recent Company survey found tabout half of the heat exchanger body flanges were chronic leakers. For servicebelow 250°F, ASME Code flanges are normally adequate.

Chevron has developed a flange design method which corrects the deficienciesthe ASME CODE. Appendix G presents the Chevron and ASME Code design methods for heat exchanger body flanges. The Chevron design method is recomended for all heat exchanger body flanges with design temperatures above 2°F.

ASME Code vs. Chevron Design MethodsThe current Code formulas are deficient in two ways: (1) They are based on depressure and ignore bolt loads and flange stresses required to pass hydrotest,



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(2) they ignore the hydrostatic end force due to operating (design) pressure in tbolt load for gasket seating.

The basic differences between the ASME Code and Company flange design methods are in the design bolt loads, W1 and W2. The design bolt load is defined asthe larger of W1 and W2.

The ASME Code defines W1 as the bolt load required to balance the sum of gaskreaction and the hydrostatic end force due to design pressure.

W1 = 0.785 G2 Pd + (2b 3.14 GmPd)

Company practice defines W1 as the bolt load required to balance the sum of gasreaction and the hydrostatic end force due to hydrotest pressure.

W1 = 0.785 G2 Ph + 2 b (3.14 G + Lp) m Ph

The ASME Code defines W2 as the bolt load required to seat the gasket at zero psure.

W2 = 3.14 b G y

Company practice defines W2 as bolt the load required to seat the gasket at design pressure. This is the hydrostatic pressure end force at design pressure plus thedefined gasket seating force.

W2 = 0.785 G2 Pd + b ( 3.14 G + Lp ) y

Fig. 500-5 Body Flange Configuration



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Using the Chevron design method will increase flange thickness by approximat50% (or more) and increase the number of bolts, depending on size, geometry,gasket material and design pressure. The benefits are no leakage and lower mnance costs (Section 1000).

Applying the Chevron modifications to cover plate design gives the following criteria.

The minimum cover plate thickness, t, is the larger of th or tso below:

Hydrotest:

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Operation:

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Hydrotest conditions almost always govern the cover plate thickness. Seating during operation may govern at low pressures and high temperatures when Sd is much less than Sc.

In the above equations:

b = Effective gasket seating width (in.), from Figure G-3, Appendix

G = Diameter at location of gasket load reaction (in.), from Figure G-3, Appendix G

hG = Gasket Moment Arm (in.), from Figure 500-6

Lp = Total length of gasket pass partition rib(s) (in.)

m = Gasket factor, from Figure G-2, Appendix G

Pd = Design pressure (psig)

Ph = Hydrotest pressure, normally equal to (1.5)(Pd)(psig)

Sc = Allowable flange (or cover plate) stress at ambient temperature(psi)

Sd = Allowable flange (or cover plate) stress at design temperature (psi)

t = Flange (or cover plate) thickness (in.)

th = Cover plate thickness for hydrotest conditions in.)

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tso = Cover plate thickness for operating conditions (in.)

W = Design bolt load (lbf)

W1 = Bolt load required to pass hydrotest (lbf)

W2 = Bolt load required to reseat a gasket in service (lbf)

Y = Factor, from Figure G-7, Appendix G

y = Gasket seating stress (psi), from Figure G-2, Appendix G

Designing And Evaluating Body FlangesChevron personnel seldom design heat exchanger body flanges from scratch, boften evaluate vendor designs or existing flanges. The PCFLANGE program, provided on a floppy in the back of this manual, automates the calculations necsary for the evaluation of flanges. Appendix H describes the operation of the PCFLANGE program.

Flange design requires decisions regarding geometry, materials, gaskets, and bThe design of a flange may be iterative, as the required bolting may dictate an increase in flange OD, which may, in turn, increase the bolt size or number.

The flange ID is set by the shell ID, which is set by the process and thermal desof the heat exchanger. The materials are dictated by the operating temperature

Fig. 500-6 Channel Cover Dimensions



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the corrosive nature of the fluids. The gasket type is dictated by the anticipatedmovement at the gasket surface due to thermal stresses and piping stresses. Tsize and number and the flange thickness are dictated by the pressures and strThe flange OD is dictated by the bolt size and number.

Materials. The choice of flange and bolt materials is based on design temperatuand the corrosive nature of the process fluid. The stress should be below creepstress limits at design fluid temperature. Flange creep is not a problem in low asteels below 750°F. At temperatures above 750°F creep may be a problem. When designing flanges in this range, consult the Materials Unit of CRTC.

Refer to the ASME Code to define the following allowable stresses:

Sa, allowable bolt stress at ambient temperature

Sb, allowable bolt stress at design temperature;

Sc, allowable flange stress at ambient temperature;

Sd, allowable flange stress at design temperature;

Se, allowable shell stress at ambient temperature;

Sf, allowable shell stress at design temperature.

Gaskets. Selection of the proper gasket is essential in flange design. See Section 540 for recommended gasket materials. The Code specifies minimum recommended gasket stress for the different gasket types. Gasket manufactureoften supply maximum stress values. One manufacturer recommends maximumspiral wound gasket stresses of 25,000 psi for asbestos filled, 13,000 psi for TFfilled, and 20,000 for GRAFOIL filled gaskets. Another manufacturer suggests 15,000 psi for a general upper limit.

Spiral wound gaskets and double jacketed asbestos gaskets are commonly useSpiral wound gaskets that are not in a recessed groove should have an I.D. comsion stop ring, or an O.D. centering ring and an I.D. compression stop ring. Bolstop rings should be on the gasket ID. A bolt stop ring on a gasket OD can actuunload a gasket as bolts are tightened. Specify 125 micro-inch finish on flange surfaces which will contact the gasket.

Gasket resilience, the ability of a gasket to maintain a seal when the two matingflanges move relative to each other, is an important gasket parameter. Solid megaskets have almost no resilience. Double jacketed gaskets can tolerate 1 to 2 of axial movement at the gasket surface. Spiral would gaskets can tolerate 4 tomils of axial movement at the gasket surface. This makes spiral wound gasketsgood replacement for double jacketed or solid gaskets for leaking flanges. However, spiral wound gaskets are usually wider than double jacketed or solid gaskets. Not all flanges have wide enough gasket seating surfaces to accommoa spiral wound retrofit. Ideally, a spiral wound gasket should incorporate a bolt sring on the gasket ID, however, this makes it even wider and harder to retrofit inplace of a double jacketed gasket. Manufacturers can supply gaskets with centtabs which aid installation.



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Gasket parameters required for a flange analysis include the gasket ID, OD, anCode values of minimum seating stress and gasket factor.

Flanges and Bolts. The flanges and bolts should be of material with similar coeffcients of thermal expansion, i.e., B-7 studs for low alloy flanges. If the materialsthe flanges and bolts are not similar, an analysis should be done to confirm thadifferential thermal expansion at design fluid temperature will not unseat the gaor yield the bolts or flanges.

Bolt relaxation (creep) is a function of both temperature and actual bolt stress. Tfollowing equations are for avoiding creep in new designs or evaluating for creeexisting designs. To avoid relaxation(creep), bolts should be used at temperatubelow the following criteria:

T < 920 - (S/180) for B-7 bolts

T < 1030 - (S/180) for B-16 bolts where:

T = Operating (design) temperature (F)

S = Target or actual bolt-up bolt stress (psi)

Code rules, as indicated in Appendix S of the ASME Pressure Vessel Code, recnize that normal bolt-up practices are not precise. Actual loadings often signifi-cantly exceed design loads. For example, in order to hydrotest a code designedflange, bolt stress must exceed Code allowable by about 50 %. Section VIII, Dision 1, rules are intended to permit this practice. However, bolt and flange streswill be below Code allowable at hydrotest for a Chevron designed flange.

Use the smallest bolts that will satisfy the spacing requirements and flange dimsions shown on Figure G-4. The number of bolts should be divisible by 4 to conform to symmetrically oriented bolting equipment. Bolt area should be calculated based on the thread root area shown on Figure G-4 (Appendix G). Bolt hodiameter should be 1/8 inch larger than bolt diameter.

A flange analysis requires specification of the number of bolts and the root meaarea of the bolts. The root mean bolt area is shown in a table below.

Flange Geometry And Stresses. Flange thickness and hub dimensions are the main variables that control the magnitude of the stresses in the flange. The PCFLANGE program prints out the stresses in the various parts of the flange athe corresponding code allowable limits. The program can be run with various flange thicknesses and hub dimensions until all the stresses are at or below coallowables. The program runs both the Code and Chevron methods so the diffeence in flange thickness for the two methods can be compared. Arbitrary bolt stresses can be specified in the program to investigate the resulting flange strehigh bolt stresses.

Flange Rotation. As flanges are stressed by forces at the bolt circle, gasket, andshell, they pivot, or bend, about the bolt circle and gasket. This bending is callerotation. All flanges rotate to some degree, even at low stresses. The rotation is



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usually not significant for small flanges, however, it can become significant for large diameter flanges and for high stresses. In cases of extreme high stress, thflanges can rotate until metal to metal contact exists between mating flanges atflange OD. Depending on the flange geometry and location of bolt stop rings, rotion can sometimes unload gaskets and cause leaks.

To approximate flange rotation, the flange is considered a free body, disregardimetal in the hub and the restraint of the nozzle neck or shell. These assumptionresult in the following equation, which slightly over-estimates the actual rotation

θ = 1.91 M R / (E b t3)where:

θ = Angle of rotation, radians

M = Total moment (in lbf)

R = Mean radius of flange (in)

b = Radial width of flange (in)

t = Thickness of flange (in)

E = Modulus of Elasticity of flange at temperature (psi)

To calculate the total moment, M, acting on the flange, consider the bolt load toacting at the bolt circle, the hydrostatic load at the inner edge of the flange (if prsured conditions are being considered), and the gasket reaction at the mean gadiameter or the bolt stop ring. Then calculate the total moment on the flange retive to the mean flange radius.

With the rotation and the flange dimensions, the deflections at any point of intercan be calculated. For example, deflection at the flange OD for rotation about thgasket is shown below:

d = θlwhere:

d = Deflection at OD of flange (in)

θ = Angle of rotation (radians)

l = Radial distance from center of gasket to flange OD (in.). Flangrotation can cause problems if deflection at the flange or OD approaches 1/2 the gasket thickness.

Thermal Gradients. Thermal stresses leading to leakage can result from transietemperature differences during start up, steady state temperature differences between tube passes at tubesheet and channel cover flanges, process variationduring operation and, for uninsulated flanges, variations in the weather, particulrain storms. It is often necessary to re-torque uninsulated bolts after each rain storms to stop leaks.



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As a general guideline, thermal effects may be significant when the maximum process temperature exceeds 350°F and/or when process temperature changes across an exchanger exceed approximately 250°F. Significant reduction in these effects can be made by insulating flanges and bolts to achieve more uniform cirferential temperatures. Belleville spring washers have been used with mixed re(see Section on Belleville Spring Washers).

In addition to the effect of circumferential thermal gradients in heat exchangerspiping, pressure vessel and exchanger closures may be subjected to appreciabthermal shocks. When process temperatures decrease rapidly, flange and gaskmaterials respond faster than the bolting. Consequently, under these conditionsleakage may occur when bolt loadings decrease and gaskets have insufficient rience to compensate for the contraction of the closure flange material.

Operation And Maintenance Of Body Flanges

Bolt-up. Following is the recommended general flange assembly procedure:

1. Inspect the gasket seating surfaces for tool marks, cracks, scratches or pittRadial tool marks on a gasket seating surface are virtually impossible to se

2. Inspect the gasket for defects or damage.

3. Inspect bolts, nuts, washer, and flange facings for galling, pitting, dirt, etc.

4. Lubricate all threads and nut facings with temperature appropriate lubricanConsider using an anti-seize compound to facilitate disassembly.

5. If necessary, use a few dabs of gasket cement to keep the gasket in positiountil the flanges are tightened. A gasket designed with centering tabs can aflange assembly.

6. Torque the bolts to no more than 30% of the final torque value following thesequence recommended in Figure 500-7 (found on pages following). Visuacheck the gap between flanges for evenness of fit-up.

7. Torque the bolts to 60% of the final torque value following the same boltingsequence.

8. Torque the bolts to 100% of the final torque value following the same boltinsequence. This may require several retorquings because as one stud is torit will relieve the stress on the adjacent stud until equilibrium is achieved.

9. Retorque the bolts after 4 hours at ambient conditions to compensate for agasket or metal relaxation.

10. Retorque the bolts after 24 hours at operating conditions to compensate fogasket or metal relaxation.

Torque. A flange analysis by PCFLANGE will define a target bolt-up bolt stress.This stress value can be converted to bolt torque using the following equation.



Fig. 500-7 Recommended Sequence for Torquing Bolts During Body Flange Assembly (1 of 2)



Fig. 500-7 Recommended Sequence for Torquing Bolts During Body Flange Assembly (2 of 2)



T = .013 S Dr3where:

T = Torque (ft lbf)

S = Target bolt stress (psi)

Dr = Bolt diameter at the thread root (in)

Bolt thread root diameters are shown below.

NominalBolt Diameter

(in.)

RootDiameter

(in.)

RootArea

(sq. in.)

1/2 0.4005 0.126

5/8 0.5071 0.202

3/4 0.6201 0.302

7/8 0.7404 0.419

1 0.8376 0.551

1 1/8 0.9628 0.728

1 1/4 1.088 0.929

1 3/8 1.213 1.155

1 1/2 1.337 1.405

1 5/8 1.463 1.680

1 3/4 1.588 1.980

1 7/8 1.713 2.304

2 1.838 2.652

2 1/4 2.088 3.423

2 1/2 2.338 4.292

2 3/4 2.588 5.259

3 2.838 6.324

3 1/4 3.088 7.487

3 1/2 3.338 8.749

3 3/4 3.587 10.108

4 3.837 11.566



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The equation for torque assumes a friction coefficient of 0.2. This has been shoto be appropriate for the size of bolts used on heat exchanger body flanges. Thresults are accurate within 10 percent for lubricated bolts and within about 15 percent for unlubricated bolts. If either the flange or nut surface has galled, a haened washer should be used to maintain the correct coefficient of friction.

Regularly calibrated torque wrenches are adequate for bolting well designed flanges. Many other devices and methods have been used for precision boltingsuch as bolt elongation measurements and control of nut methods. Mechanicaltensioning devices may be required for large bolts that are beyond hand torquewrench capabilities. Although these methods produce accurate bolt loads, theycannot compensate for inadequate Code designed flanges. No bolting device cprevent leaks in a flange that yields during hydrotest or yields when uninsulatedbolts shrink during a rain storm.

Insulation. Insulation of flanges and bolts to prevent leakage is appropriate in services operating above 250°F. Leakage problems can be reduced by applicationinsulation or rain shielding over flange surfaces and bolts normally exposed to atmospheric conditions. The sudden cooling effect of heavy rainstorms may creleakage problems that seldom disappear when normal operating conditions areagain reached. Uninsulated flanges operating above 700°F should be analyzed by an expert before insulation is applied.

Use insulation covers designed for safe leakage. See Section 100 of the Insulation and Refractory Manual and Model Specification IRM-MS-4197 for the design of leak-safe, removable insulation covers. Improperly designed insulation will soakleakage and may cause auto-ignition. A 1 or 2 inch air gap between shielding aflange is typical.

Apply insulation when the flange is cold (after hydrotest and before startup) to mmize startup stresses. Insulation may be temporarily removed after startup to inspect for leaks.

Dealing With Leaking Body FlangesSection 1000 of this manual lists common problems with gasketed joints, and gdetails for maintaining flanges. A flange analysis using PCFLANGE should be pof the diagnosis of a leaking flange to see if the design is contributing to the problem.

Bolt Tightening. Many flange leaks are caused by flanges that are too thin, in spof being designed according to the Code. The leaks can often be stopped temprarily by tightening the bolts, even if the flanges are too thin and are yielding. Thtightening should be done with a torque wrench so the bolts stress can be montored. The bolt stress can be entered into the PCFLANGE program to assess thresulting flange stress. A flange designed to Code should not break even if the have to be tightened beyond Code allowable bolt stress to stop a leak.

Gasket Change. A gasket change should be considered for a chronically leakingflange. A solid metal gasket can be upgraded to a clad (double jacketed) gaskeclad gasket to a spiral wound gasket, and a spiral wound gasket to a spiral wou



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gasket with an ID bolt stop ring. Gasket widths or styles can be changed to accdate flanges that have been machined to below design thickness. A change of gasket can also change the required bolting force. An analysis by PCFLANGE asses the flange’s ability to seat the upgraded gasket.

Rain Shields. Most leaks in uninsulated flanges would improve if the flanges weinsulated. This is because the insulated flanges would experience fewer thermatransients. However, reluctance to insulate a leaking flange is understandable. positive intermediate step would be the application of a stainless steel rain shiecompletely covering the flange, bolts, and nuts, with a gap at the bottom for leadetection. If the rain shield reduces or stops the leaks, a flexible insulation covecould be applied over the rain shield later.

Belleville Spring Washers. Belleville spring, or dished, washers (Figure 500-8) have been used in some Company plants since 1965 to compensate for thermacycling. The forced deflection of the spring on tightening keeps a steady force othe stud nut when thermal expansion of the stud occurs. Refinery experience wBelleville washers has been varied. Cracking and failure can occur in corrosiveservices, especially if they are used on internal floating heads. Washer materiamust be selected based on the expected maximum operating temperature to avcreep relaxation.

Fig. 500-8 Belleville Washer Configurations



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for shell-oval

Deflection is directly proportional to load for most Belleville washer designs. Whproperly installed deflection due to changes in loads in operating temperature should not cause washers to completely flatten nor return to an undeflected conration. The load required to deflect the springs is increased by adding them in parallel. A larger deflection for the same load can be obtained by adding springseries. To avoid damaging the spring washers, mount them on the side of the flanges opposite the stud nut which is to be turned. The manufacturer’s recommdations should be consulted and followed. Experience shows that written procedures are necessary to insure proper re-assembly of washers during future maintenance.

Backing Rings. Backing rings have been installed to reinforce inadequate heat exchanger body flanges. This options avoids the need to heat treat shell to flanwelds when flanges are replaced. The rings fit snugly behind the flanges on theshell OD, have the same OD as the flanges they support, and are notched to acmodate the hub of the existing flange. The backing rings are put on the exchanshell as two semi-circles, and then are welded together to form a solid ring. Bolholes in the rings match the bolt holes in the flanges. This option is only viable ithere are no nozzles or other attachments on the heat exchanger shell that wouinterfere with the rings and the longer bolts.

Backing rings are usually sized so that the thickness of the existing flange plus backing ring is 20 percent greater than the thickness of a Chevron designed flafor the same service.

Flange Replacement. The most reliable way to solve a chronic leak caused by aninadequate Code designed flange is to replace it with a Chevron designed flangThis options allows the flange to be designed for an upgraded gasket, with a bostop ring, and for bolts that remain below code allowable stress even during hydrotest. The welds from the flange to shell and channel will usually have to bheat treated, depending on the material. Thicker retrofit flanges can move shellchannel nozzles relative to each other, requiring piping modifications.

533 High Pressure ClosuresUsing alternative sealing techniques - high pressure closures - becomes econofor exchangers over 20 inches in diameter that operate above 1000 psig. For pructs of pressure (psig) times diameter (inches) less than 70,000, welded diaphrclosures are economic. For products of pressure times diameter greater than 8screwed or keyed type closures are economic. Break-even cost is in the 70,00080,000 range.

Figure 500-9 shows a typical welded-diaphragm closure for a high-pressure channel and low-pressure shell.

Figure 500-10 shows a welded-diaphragm closure and welded-tube sheet usedfeed-effluent exchangers with high- pressure on both shell and tube sides. Theto-tube side strength weld is located near the channel end to facilitate easy remand re-welding. The inconel overlay permits rewelding without heat treatment.



Fig. 500-9 Integral Tubesheet and Channel with Seal-Welded Diaphragm Closure



Fig. 500-10 Removable Tube Bundle with Welded Tubesheet and Diaphragm Closure



side alled

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lems ss--nnels

Figures 500-11 and 500-12 show screwed type closures for high pressure, tubeonly, and high pressure both sides, respectively. These screwed closures are c“breech lock closures.”

Designs shown in Figures 500-9 through 500-12 are appropriate when all compnents are made of the same material.

Figure 500-13 shows a segmented keyed-type design with a welded tubesheetclosure that can accommodate different metals.

High pressure closures are not commodity items. Industry standards and appropriate codes for their design do not exist. Few manufacturers can design safe, fully serviceable high pressure exchangers. Some of the common problemand solutions are discussed below.

Standard bolting practices that are appropriate for hardened studs and nuts shonot be used in high pressure closures. Bolt stresses should be less than yield sand relaxation stress of the soft base metal of the female threads.

Stainless steel internals with clad low chrome channels have caused many probfor designs shown in Figures 500-9 through 500-12. Welded stainless steel, papartition plates in low chrome channels have caused fatigue, cracking and catastrophic channel failure. Stainless steel pass partition plates in low chrome chashould be made bolted with adequate clearances to accommodate differential thermal expansion.

Fig. 500-11 Integral Tubesheet and Channel with Gasketed Closure



Fig. 500-12 Removable Tube Bundle with Gasketed Tubesheet and Closure

Fig. 500-13 YUBA Patented “Hemilok“ Design with Welded Tubesheet and Closure



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Stainless steel sleeves in Figure 500-12 have jammed set screws. This requiresdrilling them out and retapping larger holes after each shut down. Fully restrainsleeves should be made of the same material as the channel barrel.

Stainless steel diaphragms have cracked, due to thermal fatigue leading to leakand fires. Diaphragms should be made of the same material as the channel, wiappropriate corrosion allowance, or should be replaced at appropriate intervals

The nozzle packing joint in Figure 500-10 will leak between shell and tube sidethe sleeve is stainless steel and the channel is low chrome. When using these mrials, the tubesheet skirt-to-channel weld should be placed between the tubeshand nozzle and the pass partition box bolted.

Screwed covers in Figures 500-11 and 500-12 have jammed due to inappropriamaintenance practice, thread corrosion and severe thermal transients. Large thand clearances mitigate this problem.

One non-Company exchanger (similar to Figure 500-12) failed catastrophically to diaphragm leakage that pressured and expanded the threaded portion of thechannel and allowed the cover to disengage. Some similar Company exchangehave been modified with larger threads, new covers and externally stiffened chaends to prevent this possibility.

The Yuba “Hemilok” channel, shown in Figure 500-13, was developed to accomdate different metals, thermal cycling and extreme thermal transients, and to refirst cost and maintenance costs. The radial key grove clearance and segment are sized to accommodate specified metallurgy and thermal transients. The floapass partition box and flexible tubesheet-to-channel connection are evident in tfigure.

Yuba has invested in 3-D transient elastic-plastic finite element models of their channel to quickly evaluate any specified conditions. Yuba also offers a gasketeclosure that replaces the welded torus in Figure 500-13. The welded torus is mreliable and is recommended.

High pressures closures should be carefully specified including:

• material of all components• design features consistent with specified metallurgy• allowable bolt stresses (if applicable)• steady state design conditions• any transient conditions that may be encountered and their frequency• the design life of the exchanger.

Normal plant startup and shutdown transients have no impact on design. Feed failure transient with continuing recycle gas and full reactor effluent flow shouldprobably be considered for all feed/effluent exchangers in hydroprocessing planReactor temperature excursions followed by rapid depressuring should be consered in plants where excursions are possible. Transient thermal stress analysisaffects clearances and minor design details and has minimal effect on equipmecost.



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534 ConnectionsThis section discusses recommendations for piping and instrument connectionsbody of shell and tube exchangers. In general the number of connections shouminimized to minimize sources of leaks.

Inlet and Outlet NozzlesInlet and outlet nozzles for shell side and tube side fluids are required and theyshould conform to Chevron Piping Standards.

Vent and Drain ConnectionsAll high and low points on the shell and tube sides of an exchanger not otherwivented or drained by nozzles or piping should be provided with 3/4 inch connections for vent and drain. Consider where blinds will be installed in determining tneed for, or location of, vent and drains. Condensers can require dedicated venwhich operate continuously.

Pressure Gage ConnectionsAll inlet and outlet nozzles 2 inches or larger should be provided with a 3/4 inchhorizontal connection for a pressure gage unless special considerations requirebe omitted.

Thermometer ConnectionsAll inlet and outlet nozzles 4 inches or larger should be provided with a 1-inch hzontal connection for a thermowell unless special considerations require it to beomitted.

Company Practices: Construction1. Nozzle projections should be sufficient for removal of studs between flange

and insulation, jacket, shell, or head without removing insulation.

2. Pressure temperature ratings for flanges should be in accordance with ANS16.5. The rating and facing of the flange should match that of the adjoiningpiping.

3. On horizontal units, channels with nozzles not in a vertical plane should beprovided with two 3/4 inch (26.7 mm) nominal-pipe size screwed connectioone at the top and one at the bottom, to be used for venting and draining. Tshould be plugged with solid barstock steel plugs.

4. Chemical cleaning connections, when required, should consist of a pair of flanged and blinded nozzles, one at the inlet and the other at the outlet of eheat exchanger (or each series-connected group of heat exchangers). Theical cleaning connection should be made as branches on the heat exchangnozzles. The size of the chemical cleaning connections should depend on size of the heat exchanger nozzle to which they are attached, as shown in Figure 500-14.



ize sate

s ue to

e

s at

nex-

fe.

s e ets. er

have esis-

5. To double as inspection openings, the following nozzles should be 4-inch sor larger: (a) blowdown nozzle on steam generators, and (b) steam condenoutlet nozzle on vertical reboilers.

6. Steam heated vertical reboilers should be provided with a shell side vent aclose to the top tubesheet as possible to relieve air binding and corrosion dnoncondensibles.

7. Kettle-type steam generators should have one manhole located either in thshell above the tube bundle or in the shell cover.

8. All threaded connections should be welded up to first root valve except in water services.

540 GasketsThere are four types of gaskets commonly used in heat exchanger body flangeChevron facilities:

• Composition Asbestos• Double Jacketed• Spiral Wound• Solid Metal

For the design considerations of these gaskets. See Figure 500-1.

Composition Asbestos GasketsComposition asbestos gaskets are flat nonmetallic gaskets. They are durable, ipensive, quick delivery gaskets which are very forgiving to gasket surface prob-lems. Asbestos is still acceptable in all except acid service. The chemically resistant “African blue asbestos” (crocidolite) is no longer available in the U.S. Substitution of “Canadian white asbestos” (chrysotile) in acid services is not sa

Some operating locations, however, are replacing composition asbestos gasketwith nonasbestos substitutes because of health hazards connected with both thhandling of asbestos fibers during manufacturing and disposal of the used gaskFor most services, flexible graphite is the best alternative to asbestos. Many othnonasbestos materials are available at less cost than flexible graphite, but theygenerally inferior mechanical properties, and lower temperature and chemical rtance.

Fig. 500-14 Nozzle Sizes for Chemical Cleaning

Chemical Cleaning Nozzle, in. Shell Side Nozzle, in. Tube Side Nozzle, in.

2 2 to 4 2 to 6

3 6 8

4 8 and larger 10 and larger



Heat Exchanger and Cooling Tower M

anual

March 1994

500-36Chevron Corporation

Fig. 500-15 Gasket Design Considerations (1 of 3)

g Surface Condition Seating Surface Type

0 rms when new. ctions less than 50% fective gasket

width can be toler-

All surfaces except with nubbin or tongue and groove.

rms when new. Grit-ish OK but not ended. Cannot imperfections ner and outer verlaps sit on urface.

Never use with nubbins.

0 rms when new. ctions less than 25% fective gasket

width can be toler-

Never use with nubbin. Must have compression ring when used with raised faced flanges. Compression rings are not needed if flanges are designed to fit metal-to-metal with tongue and groove. Make certain that gasket is crushed correctly.

rms when new. tolerate less than surfaces.

Nubbin. (Nubbin is a tooth on the gasket surface which bites into metal gasket)

Gasket Type Typical Location Process Contraints Min/Max Dimensions Seatin

CompositionAsbestos

All closures except floating head.

450°F @200 psi max. 3/4" minimum width. 125—25Imperfeof the efsurfaceated.

Double-Jacketed All closures. 650°F @600 psi max. 3/8"—3/4"; refer to TEMA. Best performing width is 1/2"—5/8" with centering tabs.

50—125blast finrecommtoleratewhere ingasket ogasket s

Spiral Wound All closures except floating head.

Based on materials used for windings and filler (refer to ASME material tables).

1/2" min., no max. Take 1/8" off O.D. when calculating flange Tmin or bolt loads. It does not provide seal.

125—25Imperfeof the efsurfaceated.

Solid Metal Floating heads. It’s common to find them at other loca-tions on high pressure units such as the feed versus effluent units at the Isomax Reactors.

Highest pressures and temperatures

1/8" min. width to nubbin. Solid metal gasket width for 1/8" nubbin is 3/8" min. Refer to ASME Section 8 for more details.

50—125Cannot perfect


anual500 Shell and Tube Exchanger Com

ponent Design Considerations

Chevron Corporation500-37

March 1994

in low temperature and pressure.

tos cannot be used.

works. Double-jacketed with two pieces es seal as well as a spiral.

face is required.

F

Gasket Type Future Concerns Pros Cons When to Use

Composition Asbestos Asbestos material may be discontinued. Nonasbestos materials are only good to 150—250°F @ 100—200 psi.

Cheap. Easy to handle. Quick to obtain. Does not require special seating surfaces. Good performer. Great for salt water service.

Can blow out if unit is over-pressured. Difficult to install because gasket is not rigid. Must use asbestos handling procedures for installation, removal, and disposal.

Whenever possible

Double-Jacketed None Reliable. Easy to obtain in most materials. Can be retorqued.

Integral ribs cause leaks. Soldered ribs break off easily. Carbon steel gaskets rust when used and cause-leaks.

Composition asbes

Spiral Wound None Expensive but has best sealing capabilities of all gasket designs. Takes longer to order and obtain.

Requires perfect flange gasket seating surfaces. Windings explode when handled roughly. Cannot be retorqued to stop leaks.

When nothing else of filler can sometim

Solid Metal None Requires less bolting and Tmin because there is less gasket being seated. Does not need wide gasket; typi-cally 3/8" min. (refer to TEMA). Can seal against high pressure and tempera-ture.

Nubbins require more main-tenance. Gasket must be centered on nubbin perfectly. Marks or gouges across nubbin must be repaired. Nubbin edges require renewing every 2nd or 3rd assembly. Tends to leak if temperatures are not circumferentially uniform.

When a nubbin sur

ig. 500-15 Gasket Design Considerations (2 of 3)



anual

March 1994

500-38Chevron Corporation

ThicknessEffects What to do if Gasket Leaks

uggested; thicker less bolting (see “y” values in ASME

8). Thicker is easier out.

d asbestos filled is Use double s filler for more ke effect; this can modate some cycling (typical or some pump head ).

can handle more ssion and higher g pressures.

Retorque. Most leaks cannot be stopped and a gasket change-out is neces-sary.

and nubbin will cut Too thick and gasket deform correctly vide good seal. at determined by aterial.

Retorque. Leaks are usually from damaged nubbins and require nubbins to be inspected and repaired.

Fig. 500-15 Gasket Design Considerations (3 of 3)

GasketType

TorquingProcedures

OrderingDescription

SpecialConsideration

CompositionAsbestos

Not necessary for seating the gasket. Torque if leaker only.

O.D. × I.D., 1/16" tk, Durable (include rib configuration).

Gasket can extend beyond seating surfaces for centering in gasket surface.

1/16" is srequires“m” andSectionto blow

Double-Jacketed Normally required. Channel to shell closure most critical because of different process environments. Retorque when unit is hot.

O.D. × I.D., 3/32 tk (include rib configuration), 304ss clad asbestos with silver solder rib(s) to ring I.D., add (#) centering tabs to ring O.D. or I.D. 1/8" × 1/4". Order spares.

Gasket cannot extend past seating surfaces. Use double shell double-jack-eted design if bolting is too great. Use double filler thick-ness when thermal cycling is constant.

Standar3/32" tk. asbestospring liaccom-thermaldesign fgaskets

Spiral Wound Required for all spiral wound gasketed closures — raised faced or metal-to-metal designed flanges.

O.D. × I.D., ring dimensions if present, thickness of wind-ings and rings, materials for all parts, and always include the operating pressure the gasket must seal against. Order spares because of handling problems.

Must have compression ring on O.D. if not contained. I.D. ring helps protect windings from process and makes easier to handle. Filler mate-rials like ceramic or flexite are rock like and very hard to seal.

Thicker compreoperatin

Solid Metal Required to ensure equal load throughout gasket.

O.D. × I.D., 1/16" tk. Typically no ribs for solid metal design.

Many closures are designed with nubbins and clad gaskets. Be sure to look at both a clad with no nubbin, and a solid metal with no nubbin before nubbin is removed. Gasket material must be softer than nubbin material.

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The following are recommended for asbestos substitute gasket materials:

1. Use flexible graphite sheet, graphite-filled spiral-wound gaskets or double-jeted gaskets for all hydrocarbon and steam services.

2. Use reinforced elastomer bound sheet gaskets for AWSI Class 150/300°F maximum water service.

Use PTFE (“Teflon”) filled spiral-wound gaskets (first choice) or sheet gaskets (second choice) in sulfuric acid service. See Figure 500-16 for other chemical service recommendations.

Double-Jacketed Gaskets (See Figure 500-17)Double-jacketed gaskets have greater compressibility and resilience than solid-metal gaskets. Even compression is achieved by the use of the overlapped jackthe inside and outside diameters. Double-jacketed gaskets are generally reliabhowever, they are much less forgiving to gasket surface alignment or surface cotion problems than composition asbestos.

Spiral-Wound Gaskets (See Figure 500-18)Spiral-wound gaskets provide the best sealing capabilities. However, they toleraless flange face misalignment and require more care in handling than either comsition asbestos or double-jacketed gaskets. They are custom-designed to meetcompression requirement of body flange bolting. Spiral-wound gaskets are fullyseated when the flanges are pulled up snugly against the compressing guide riThis ring also prevents gasket crushing by over-tightening of bolts.

(1) This table gives conservative recommendations for materials resistant up to at least 200°F. Please consult with the Materials and Engineering Analysis Division when selecting gaskets for a new chemical service.

(2) All concentrations.(3) PTFE (Teflon) is a suitable replacement material for “Blue African” asbestos for all the chemical services

listed above. PTFE sheet gaskets are not fire safe, whereas PTFE filled spiral-wound gaskets (SWG) are often considered fire safe. Thus the first choice for most acid applications will be Teflon-filled SWG.

(4) Flexible graphite is “fire safe” and suitable for most chemical services except those that are highly oxidizing, such as nitric acid or concentrated sulfuric acid.

Fig. 500-16 Suitability of Materials in Sheet- or Spiral-Wound

Gaskets(1) Service(2) PTFE(3) Suitable?Graphite(4) Suitable?

H2SO4 Yes No Alloy 20

HNO3 Yes No T-304 SS

HF Yes Yes Monel

H3PO4 Yes No T-316 SS

HCI Yes Yes Hastelloy C

Note Metal for foil reinforcement of graphite sheet gaskets, or for windings in SWG.



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Solid-Metal GasketsThese gaskets, are prone to leakage and are no longer recommended.

Solid-metal gaskets come in many shapes. They have good strength and are retant to corrosion. They are effective at higher temperature and pressures than tother types of gaskets. Solid-metal gaskets require an excellent seating surfacecondition and alignment. They have been used with nubbin-seating surfaces. Anubbin is a very small (1/4 inch wide) seating surface on the face of the flange.Because the nubbin is small, less force is needed to seat the gasket.

Fig. 500-17 Double-Jacketed Gasket

Fig. 500-18 Spiral-Wound Gasket



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550 Insulation

551 Reasons for InsulatingExchanger shells, channels, and flanges are insulated for the following reasons

• To minimize heat loss and consequently save fuel. Obviously, insulating manot be appropriate in “cooling services,” such as for cooling water exchang

• To protect personnel working where surfaces are over 140°F. Exchangers which are not readily accessible need not necessarily be insulated to protepersonnel. Exchanger shells which are accessible but should not be insulamay use alternative means to protect personnel, such as guard posts and s

552 Types of InsulationCalcium silicate, fiber glass, and mineral wool are the common types of insulatiused on exchangers. Calcium silicate is generally preferred, especially in areashigh-foot traffic or where flammability is a concern. See Section 100 of the Insula-tion and Refractory Manual for more information on the types of insulation avail-able. See IRM-MS-1381 for installation requirements.

553 WeatherjacketingIn general, 3/16 inch pitch cross-crimped aluminum weather jacketing should bused on exchanger shells, and flat aluminum or mastic weather jacketing used exchanger heads. See Section 100, Model Specification IRM-MS-1381, and Stdard Drawing GD-N99785 in the Insulation and Refractory Manual for more detailed information.

554 Flange InsulationIn general, body flanges over 100°F should be insulated for the following reasons:

• To save heat

• To protect personnel

• To prevent large thermal gradients across the flange during inclement weatconditions. Large thermal gradients across the flange can cause distortion the flange and ultimately cause the flange to leak.

If the flanges and insulation are improperly designed, the following problems moccur:

• If flange and bolts are not of similar materials, differential thermal expansioat the operating temperature can unseat the gasket or cause the bolts or flato yield.



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• The bolts will relax and stretch if the internal operating temperature is abovthe creep stress limit of the bolts. This problem can cause leakage and posauto-ignition because of the high temperatures.

• Improperly designed insulation will “soak up” leakage and may cause auto-ignition.

To prevent these problems, the following criteria should be used for design andlation of flanges:

• The flanges and bolts should be of similar material, i.e., B7 or B16 studs focarbon steel or low alloy flanges.

• Flange and bolt materials should be designed for the maximum internal destemperature and corrosive nature of the process fluid. In other words, the sin flange and bolt material must be kept below the creep stress limits at themaximum internal design temperature.

Temperature limits for commonly used studs are as follows:

• Use insulation covers designed for safe leakage. See Section 100 of the Insula-tion and Refractory Manual and Model Specification IRM-MS-4197 for the design of leak-safe, removable insulation covers.

• Apply insulation when the flange is cold (after hydrotest and before startup)minimize startup stresses. Insulation may be temporarily removed after stato inspect for leaks.

A practical problem in the plants is the mixing of B7 and B16 studs, especially during plant turnarounds when a great deal of bolting and unbolting is being doIf a location cannot guarantee that these studs can be totally segregated, then another option is to leave flanges over 750°F uninsulated with a weathercover overthe flange to protect against wind and rain.

Less than 750°F A193 B7

750°F to 950°F A193 B16

Above 950°F Consult CRTC’s Heat Exchanger specialists, Fuels and Processing Unit on a case-by-case basis.


CHEVRON - Shell and Tube Exchanger Component Design Considerations

Documents

cooling tower

model specification

breech lock

tube rupture

differential

include rib

allowable

minimum pressurizing