EXH500 shell and tube exchanger component design considerations

This section discusses the mechanical design of shell and tube heat exchangers and their components. Emphasis is placed on company practices which differ from industry standards. Contents: Design Pressure and Temperature Bundle Design Channel and Shell Design heat exchanger gasket heat exchanger insulation

525 Retrofitting Floating Head Bundles with U-tubes

510 Design Pressure and Temperature

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

External Pressure

Exchangers operating at less than atmospheric pressure should be designed for an external pressure (vacuum) of 15 psig All exchangers designed for internal pres-sure 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 corro-sion allowance.)

512 Design Temperature

The design temperature for any part of a heat exchanger is the maximum allowable operating 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

the fluid on that side (Note that tubes are exposed both to shell side and tube side 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 exchanger

is the temperature at which the ASME Code allowable stress for the component is determined, and must be above the maximum expected operating temperature Normal operating temperature is only occasionally related to design temperature For example, tubes exposed to treated cooling tower water, irrespective of metal-lurgy, 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 func-tional range However, the name plate design temperature is usually the highest temperature at which the specific material maintains its maximum allowable stress For carbon steels, this temperature is 650°F

Minimum pressurizing temperature, an important design parameter, is discussed

in detail in the Pressure Vessel Manual In brief, the reason for establishing a

minimum pressurizing temperature is to avoid a brittle fracture Ordinary carbon steels, for example, become brittle at low temperatures The ductile-to-brittle transi-tion 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 material which will not suffer brittle fracture under the design operating conditions of the exchanger This includes hydrotest, which should be done at a temperature above the minimum pressurizing temperature

513 Relief Valves

Pressure Relief

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

Consequently, many heat exchangers do not have pressure relief valves directly on the vessel, but are rather part of an overall hydraulic system which does have protec-tion 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.

3 The component is not protected by a PSV.

To minimize the possibility of a TRV releasing during a PSV relief, the TRV can be set at 110% of design pressure as allowed by ASME Code One thermal relief valve can serve as the protective device for multiple exchangers in series if there are no block valves between them

514 Rupture Surge Pressure

All exchangers with liquid or vapor-liquid mixtures on the low pressure side should

be designed for tube rupture safety This is accomplished by setting the design sure on the low pressure side equal to the maximum normal operating pressure plus the initial surge pressure due to the complete break of one tube Long term (e.g., 2 + seconds) pressure transients should be prevented with relief devices in the piping Tubesheets, shells, shell covers, and channels should be designed to this surge pres-sure Body flanges should also meet ASME Code requirements but not leak tight-ness requirements at this design pressure

pres-Tube rupture is particularly a problem in high pressure gas/low pressure cooling water applications Appendix F gives a detailed procedure and examples for deter-mining rupture surge pressure and rupture flow rate

520 Bundle Design

521 Tubesheet Design

This section covers the applicable codes and industry practices for establishing tubesheet design and tubesheet thicknesses Tubesheets separate the shell side and tube side fluids and provide the anchor point for tube ends TEMA standard rules for calculating tubesheet thickness are used in the industry extensively

TEMA

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

ASME Code, Section VIII, Division 1

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

ASME Code, Section VIII, Division 1, Appendix A, tells how to calculate 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 provide

allow-an expallow-ansion joint in the shell of allow-an exchallow-anger Appendices A allow-and AA of Section VIII are both nonmandatory and therefore do not have to be followed by a vendor unless required by the Company

Waste Heat Boiler–Fixed Tubesheet Exchanger Type

Waste 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, which

account for staying capacity of the tubes Tubesheet thickness is governed by the largest unstayed area, which is usually the annular space between the bundle and the shell

Tubesheets designed by TEMA rules would be much thicker and are unacceptable for 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 thickness and API 660, Paragraph 7.8, for flatness tolerance on new tubesheets, although this is generally not a problem

• Some Company locations add a “maintenance” allowance (usually 1/8 inch) onto the channel side tubesheet thicknesses beyond TEMA minimum require-ment to compensate for any surface repairs required due to maintenance activi-ties

Clad Tubesheets

For clad tubesheets with rolled tube-to-tubesheet joints, the nominal cladding ness should be 1/2 inch minimum, and one of the grooves or serrations in each tube hole should be completely within the cladding The cladding thickness may be less for welded tube-to-tubesheet joints

thick-Roll-clad is the preferred method of cladding or overlay However, explosion ding is sometimes used, especially for small pieces like tubesheets where roll clad-ding is not economical For other requirements on cladding, refer to EXH-MS-

clad-2583, included in this manual

Bundle Pull Hole Design

Removable 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 located at tube 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 account tubesheet material and be designed for a maximum pulling load equal to twice the bundle weight The threads should be National Course Series below 1 inch and eight-pitch series for 1 inch and above

Bundles smaller than 20 inches in diameter may have two tapped holes for pulling eyes Small pre-engineered exchangers are usually supplied without pulling holes.These guidelines may have to be modified or waived for special construction, such

as for thin clad tubesheets Pull holes should be protected in service by threaded plugs

522 Tube-to-Tubesheet Connection

The 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 tubesheet by the following methods

• Expanding tube inside tubehole

• Welding tubes to tubesheet

Expanding Tubes Inside Tubeholes

Expanded tube-to-tubesheet joints are industry standard In this case tubes are expanded inside tubeholes by such methods as rolling or applying hydraulic pres-sure directly to the tube end Properly rolled joints have uniform tightness to mini-mize tube fractures, stress corrosion, tubesheet ligament enlargement, and dishing

of the tubesheet Rolling to 95% of tubesheet thickness is recommended Rolling at

or beyond the tubesheet thickness is not recommended—for it may damage the tubes

For moderate general process requirements (less than 300 psi and less than 350°F) tubesheet holes without grooves are standard For all other services with expanded tubes at least two grooves are machined (1/8 inch wide by 1/64 inch deep) in each tube hole See Figure 500-1

Fig 500-1 Rolled Tube - Tubesheet Connection

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 tube hole

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 times used as an indicator of adequate rolling

some-4 Hydrotest the shell side after the final rolling

Welding the Tubes to Tubesheet

Additional tightness beyond that of the tube rolling is sometimes required in the following 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 tubesheet Seal welding is defined as a very small bead of weld around the tubes where no credit can be taken for strength of that weld for calculation of tube-to-tubesheet joint 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 exercised during all steps of assembly not to contaminate cleaned parts A chronic problem especially in sour services is contamination of the weld with sulfur or iron sulfide coming 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 much like new 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 a hot alkaline 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

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

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

9 Turn tube bundle in horizontal position so tube ends are in vertical rows and weld 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 desired, both ends may be welded at once

12 Give tubes full roll

13 Place bundle in shell and test Repair as required and repeat dye-checking ations

oper-Note It is important that a specific weld procedure be developed for the work and the actual materials used The shop doing the work must demonstrate qualification

to use this procedure Consult with a local welding specialist or the Material sion Welding Specialist for help in developing the weld procedure.

Divi-523 Longitudinal Shell Baffles

In the design of heat exchangers, it is sometimes advantageous to use a TEMA Type “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 prevent bypassing, the seal between the longitudinal baffle and the shell is most commonly

a “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 Thickness

The Company recommends that the longitudinal baffle thickness be the largest of these three: (1) 1/4 inch, or (2) the thickness required by TEMA, for transverse baffles, or (3) the thickness for differential pressure loading

susceptible to damage by corrosion However, only thin strips can flex adequately

to 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 hydropro-cessing systems)

mate-The angle of contact between shell and flexible strips should be small so that tion during installation is minimized and the differential pressure has the greatest effect in causing the strips to seal To this end, it is recommended that dimensions

fric-“A” and “B” in Figure 500-3 should be about equal, with both in the range of 1/2 to 3/4 inch

Fig 500-2 Lamiflex Baffle

Fig 500-3 Lamiflex Baffle Dimensions

Protection during installation The lamiflex baffle must be protected with

crib-bing to avoid damage during rigging operations

Fiber Fabric—“Thermo Ceram” Baffle

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

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

is 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, phone 654-3274)

tele-Fig 500-4 Ceramic Fiber-Type Seal

Transverse Baffles The transverse baffles must be notched to provide clearance

for the fiber fabric seal (See Figure 500-4) The clearance area at this notch should

be minimized, for it adds to the leakage through the transverse baffle

by corrosion, erosion, or vibration However, use of an impingement device when it

is not needed increases exchanger diameter and cost

Chevron Practices

TEMA recommends impingement plates for most services Impingement plates have 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 the projection 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 less than

20 inches) and are usually not provided

Impingement Rods

Impingement rods are preferred to an impingement plate for several reasons First, the plate creates a dead space directly beneath it, lowering the heat transfer in those tubes Also, if the plate blocks too much of the inlet area, then the fluid may accel-erate 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 row is

at 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 diameter

of the shell inlet nozzle The actual length of the rods may extend beyond the effective 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

TEMA Guideline

The TEMA Standard provides a minimum guideline for determining when an impingement device should be used This guideline is appropriate for Company use also Impingement protection underneath the shell inlet nozzle is recommended for the 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 bundle entrance 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 underneath

the 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 either side of shell inlet nozzle The preferred construction is to attach the plate to the bundle

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

Distribution Belt A distribution belt consists of a collar that fits around the shell at

the inlet and/or the outlet The shell nozzle attaches to this collar The fluid enters through the nozzle and flows through the annulus between the belt and the shell The fluid enters the tube bundle through windows cut in the shell, with a reduced velocity Distribution belts are not widely used in the Company They are expensive and have maintenance problems

525 Retrofitting Floating Head Bundles with U-tubes

It is often advantageous to change a floating head bundle to a U-tube bundle This change may be warranted because of excessive leaks between floating head flange and the tubesheet Recent progress in U-tube bundle cleaning methods allows use

of U-tubes in many more services

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 the existing exchanger performance requirement or new data as specified by the process engineer based on information from the field about the opera-tion 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-evalua-tion 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 bundle The only 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 by using the Company/HTRI Programs or by using an exchanger design contractor

– Compare cost of retrofit to cost of new exchanger: extensive modifications

to channel or shell may justify purchase of a complete new exchanger.– Consider the possibility that it may be necessary to remove the channel pass partition plates on the existing unit and install new ones This is not considered extensive modification

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

Stan-530 Channel and Shell Design

This section covers mechanical design of the channel and shell on a shell and tube exchanger Refer to EXH-MS-2583 for more details on channel and shell construc-tion

531 General

Channel and Shell Thickness

• The channel and shell contain the two separated fluids in the exchanger They 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

of the tube bundles

532 Body Flanges

Body flanges are used to permit disassembly and removal or cleaning of internal parts of a heat exchanger Integral flanges (hub or weld neck) are flanges that are integral with the exchanger wall or neck This type of flange is recommended on all services and pressures except in water service for pressure up to 150 pounds Loose 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 as gral construction (no flanges), welded diaphragm seals, or breech lock closures Welded diaphragm and breech lock closures are discussed in Section 533

inte-The ASME Boiler and Pressure Vessel Code establishes the minimum requirement for a flange design and provides a method of calculation (Section VIII, Division 1, Mandatory Appendix 2 and Non-mandatory Appendix S) Deficiencies in ASME Code designed flanges, from a leakage standpoint, have been recognized for some time Although records are not routinely kept, a recent Company survey found that about half of the heat exchanger body flanges were chronic leakers For services below 250°F, ASME Code flanges are normally adequate

Chevron has developed a flange design method which corrects the deficiencies of the ASME CODE Appendix G presents the Chevron and ASME Code design methods for heat exchanger body flanges The Chevron design method is recom-mended for all heat exchanger body flanges with design temperatures above 250°F

ASME Code vs Chevron Design Methods

The current Code formulas are deficient in two ways: (1) They are based on design pressure and ignore bolt loads and flange stresses required to pass hydrotest, and

(2) they ignore the hydrostatic end force due to operating (design) pressure in the bolt 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 as the larger of W1 and W2

The ASME Code defines W1 as the bolt load required to balance the sum of gasket

reaction 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 gasket

reaction and the hydrostatic end force due to hydrotest pressure.

W1 = 0.785 G2 Ph + 2 b (3.14 G + Lp) m PhThe ASME Code defines W2 as the bolt load required to seat the gasket at zero pres-sure

W2 = 3.14 b G yCompany 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 the defined gasket seating force

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

Fig 500-5 Body Flange Configuration

Using the Chevron design method will increase flange thickness by approximately 50% (or more) and increase the number of bolts, depending on size, geometry, gasket material and design pressure The benefits are no leakage and lower mainte-nance costs (Section 1000).

Applying the Chevron modifications to cover plate design gives the following

In the above equations:

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

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.)

t = Cover plate thickness for hydrotest conditions in.)

th G

0.3Ph

Sc -

1.9 h⋅ G

ScG3 -+

1.9W h⋅ G

SdG3 -+

=

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 Flanges

Chevron personnel seldom design heat exchanger body flanges from scratch, but often evaluate vendor designs or existing flanges The PCFLANGE program, provided on a floppy in the back of this manual, automates the calculations neces-sary for the evaluation of flanges Appendix H describes the operation of the PCFLANGE program

Flange design requires decisions regarding geometry, materials, gaskets, and bolts The 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 design

of the heat exchanger The materials are dictated by the operating temperature and

Fig 500-6 Channel Cover Dimensions

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

Materials The choice of flange and bolt materials is based on design temperature

and the corrosive nature of the process fluid The stress should be below creep stress limits at design fluid temperature Flange creep is not a problem in low alloy steels 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 manufacturers often supply maximum stress values One manufacturer recommends maximum spiral wound gasket stresses of 25,000 psi for asbestos filled, 13,000 psi for TFE filled, 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 used Spiral wound gaskets that are not in a recessed groove should have an I.D compres-sion stop ring, or an O.D centering ring and an I.D compression stop ring Bolt stop rings should be on the gasket ID A bolt stop ring on a gasket OD can actually unload 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 mating flanges move relative to each other, is an important gasket parameter Solid metal gaskets have almost no resilience Double jacketed gaskets can tolerate 1 to 2 mils

of axial movement at the gasket surface Spiral would gaskets can tolerate 4 to 5 mils of axial movement at the gasket surface This makes spiral wound gaskets a good 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 accommodate

a spiral wound retrofit Ideally, a spiral wound gasket should incorporate a bolt stop ring on the gasket ID, however, this makes it even wider and harder to retrofit in place of a double jacketed gasket Manufacturers can supply gaskets with centering tabs which aid installation

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

Flanges and Bolts The flanges and bolts should be of material with similar

coeffi-cients of thermal expansion, i.e., B-7 studs for low alloy flanges If the materials of the flanges and bolts are not similar, an analysis should be done to confirm that differential thermal expansion at design fluid temperature will not unseat the gasket

or yield the bolts or flanges

Bolt relaxation (creep) is a function of both temperature and actual bolt stress The following equations are for avoiding creep in new designs or evaluating for creep in existing designs To avoid relaxation(creep), bolts should be used at temperatures below 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, recog-nize that normal bolt-up practices are not precise Actual loadings often signifi-cantly exceed design loads For example, in order to hydrotest a code designed flange, bolt stress must exceed Code allowable by about 50 % Section VIII, Divi-sion 1, rules are intended to permit this practice However, bolt and flange stresses will be below Code allowable at hydrotest for a Chevron designed flange

Use the smallest bolts that will satisfy the spacing requirements and flange sions 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 calcu-lated based on the thread root area shown on Figure G-4 (Appendix G) Bolt hole diameter should be 1/8 inch larger than bolt diameter

dimen-A flange analysis requires specification of the number of bolts and the root mean area 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 and the corresponding code allowable limits The program can be run with various flange thicknesses and hub dimensions until all the stresses are at or below code allowables The program runs both the Code and Chevron methods so the differ-ence in flange thickness for the two methods can be compared Arbitrary bolt stresses can be specified in the program to investigate the resulting flange stress at high bolt stresses

Flange Rotation As flanges are stressed by forces at the bolt circle, gasket, and

shell, they pivot, or bend, about the bolt circle and gasket This bending is called rotation All flanges rotate to some degree, even at low stresses The rotation is

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, the flanges can rotate until metal to metal contact exists between mating flanges at the flange OD Depending on the flange geometry and location of bolt stop rings, rota-tion can sometimes unload gaskets and cause leaks

To approximate flange rotation, the flange is considered a free body, disregarding metal in the hub and the restraint of the nozzle neck or shell These assumptions result 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 to be acting at the bolt circle, the hydrostatic load at the inner edge of the flange (if pres-sured conditions are being considered), and the gasket reaction at the mean gasket diameter or the bolt stop ring Then calculate the total moment on the flange rela-tive to the mean flange radius

With the rotation and the flange dimensions, the deflections at any point of interest can be calculated For example, deflection at the flange OD for rotation about the gasket 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.) Flange rotation 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 transient

temperature differences during start up, steady state temperature differences between tube passes at tubesheet and channel cover flanges, process variations during operation and, for uninsulated flanges, variations in the weather, particularly rain storms It is often necessary to re-torque uninsulated bolts after each rain storms to stop leaks