CHEVRON Pressure Vessel - Mechanical and Structural Design

400 Mechanical and Structural Design

AbstractThis section outlines the responsibilities of designing and fabricating vessels. It discusses establishing the design basis, covers the details of mechanical design, and gives an overview of structural design. It also discusses calculation sheets and computer programs available to assist in designing vessels.

Most of the pressure vessels used in the petroleum and chemical industries have cylindrical shells with elliptical or hemispherical heads. They are relatively simple to design, fabricate, and install in either vertical or horizontal positions. The design is usually governed by internal pressure, but a few are designed to operate below atmospheric pressure.

Most pressure vessels are designed according to the rules of the ASME Code, Section VIII, Division 1, which is therefore emphasized in this section.

The information in this section is not intended to be a substitute for the Code, but rather to clarify the requirements of, and to save time in using the Code. You must use a current Code in designing, fabricating, and testing new vessels.

Contents Page

410 Owner/User’s and Manufacturer’s Responsibilities 400-3

411 ASME Code, Section VIII, Division 1


420 Determining Design Conditions 400-6

421 Design Pressure: Overview

422 Design Pressure: ASME Code, Section VIII, Division 1


424 External Pressure

425 ASME Code, Section VIII, Division 1, Design Temperatures


Chevron Corporation 400-1 April 2000

400 Mechanical and Structural Design Pressure Vessel Manual

427 Wind and Earthquake Design

428 Corrosion Allowance

429 External and Internal Loads

430 Mechanical Design 400-18

431 Background Information and Overview

432 Design Summary

433 Design for Internal Pressure

434 Design for External Pressure

435 Example of Internal/External Pressure Design

436 Openings and Nozzle Reinforcement

437 Bolted Flanged Connections

438 Minimum Wall Thickness and Nominal Plate Sizes

439 Design of Welded Joints

440 Structural Design 400-56

441 Overview

442 Vertical Vessels: Combining Structural Loads

443 Horizontal Vessels: Structural Design

444 Allowable Stresses and Deflections

445 Shell Stresses from Wind and Earthquake

446 Internal Loads

447 Structural Supports

450 Calculation Sheets and Computer Programs 400-76

451 Overview

452 Calculation Sheet, PVM-EF-65

460 Quick Reference Guide to ASME Code, Section VIII, Division 1 400-77

April 2000 400-2 Chevron Corporation

Pressure Vessel Manual 400 Mechanical and Structural Design

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410 Owner/User’s and Manufacturer’s Responsibilities

OverviewThe ASME Code, Section VIII, assigns certain responsibilities to the owner/user, manufacturer, and authorized inspectors that, when all discharged properly, provide for the design and construction of pressure vessels that will be safe to operate under the intended service conditions. The Code is the minimum requirement for safe containment of the design pressure at the design temperature.

The Code does not give direct consideration to service conditions that can affect the performance of a pressure vessel, such as process environments and operating procedures. The owner/user is solely responsible for specifying requirements that exceed the Code minimums depending on the specific service conditions that a pres-sure vessel will be exposed to. This responsibility should be thought of as a Code requirement placed upon the owner/user beyond the explicit rules of the Code, but the Code leaves it to the judgment of the owner/user how to fulfill this responsi-bility within broad guidelines.


Owner/User’s ResponsibilitiesParagraph U-2 of ASME Code, Section VIII, Division 1, defines the owner/user’s (Company’s) and manufacturer’s responsibilities. The owner/user, or his designated agent, must “establish the design requirements for pressure vessels, taking intoconsideration factors associated with normal operation, and such other conditiostartup and shutdown.” In this manner, the Code recognizes that all of its desigrules and construction details may not be appropriate for the specific process eronment and operating conditions. Therefore, the owner/user is made responsifor specifying requirements that exceed the Code rules, or prohibiting the use osome details of construction to assure safe operation for the intended service ctions. This task can be delegated to a “designated agent,” such as the engineecontractor for a project, but the final responsibility for safe operation still resideswith the owner/user.

The need for corrosion allowance and postweld heat treatment beyond the requments of the Code are specifically mentioned in Paragraph U-2. The Companyfulfills its responsibilities under Paragraph U-2 by specifying to the manufacturethe requirements for corrosion allowance and postweld heat treatment for eachsure vessel, as discussed in Sections 500 and 600 of this manual. Additional gCompany requirements that exceed the minimum Code requirements may inclu

1. Double-V butt welds or equivalent required for all girth, longitudinal, and heto-shell welds.

2. Full penetration welds required for all nozzle and flange welds.

3. Integrally reinforced nozzles required for shell components with a thicknessgreater than 2 inches, and for operating temperatures above 650°F.



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4. Assured resistance to failure by brittle fracture during startup and shutdown, by requiring the materials of construction to have adequate CV-impact toughness at the minimum design metal temperature (MDMT).

5. Seismic and wind design practices that exceed ANSI standards.

These additional requirements significantly increase the reliability and safety of pressure vessels for the service conditions which they usually operate under at refin-eries and chemical plants. The additional expense incurred for these requirements is adequately compensated for by lower maintenance costs and reduced loss of production due to unscheduled shutdowns for repair. Exceptions can sometimes be made if the manufacturer advises that the additional expense appears unreasonably high, but these exceptions should not be granted until they are discussed with an experienced pressure vessel engineer.

Refer to the Specifications in this manual for recommendations dealing with the five factors above.

Manufacturer’s ResponsibilityParagraph U-2 of ASME Code, Section VIII, Division 1, requires the manufacturer of a pressure vessel to hold a “Certificate of Authorization” from ASME, and to satisfy an authorized inspector that all applicable requirements of the Code havbeen met before applying the “U” stamp to the vessel's nameplate. The manufaturer (i.e., Certificate of Authorization holder) is responsible for the calculationsrequired to determine the minimum thickness of all components of the vessel, regardless of mechanical design input from the Company or the dimensions onCompany's drawings submitted to the manufacturer. The manufacturer must alprepare a “Manufacturer's Data Report” that contains data on design temperatuand maximum allowable working pressure, materials of construction, corrosionallowance, thicknesses of major components, nondestructive examinations, joinefficiencies, and postweld heat treatment. This data report must be signed by bthe manufacturer and the authorized inspector, and it is usually accepted by locjurisdictions (that have a pressure vessel law requiring compliance with the ASCode), as proof that legal requirements have been met.

Authorized Inspector’s ResponsibilityAuthorized Inspectors receive a commission from the National Board of PressuVessel Inspectors to inspect pressure vessels for Code compliance, and are usemployed by an insurance carrier (i.e., Hartford or Travelers, et al.) that servesthe manufacturer's inspection agency. The Authorized Inspector cannot be direemployed by the manufacturer.

The Authorized Inspector is required to monitor the manufacturer's quality contand nondestructive examinations, verify that all required calculations have beenmade, and make all other inspections necessary to certify that the pressure veshas been designed and fabricated according to all applicable rules of the CodeAuthorized Inspector is not, however, required to verify the accuracy of the manfacturer's calculations or compliance of nondestructive examinations with Codeacceptance standards. Furthermore, he should not be relied upon to verify that



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Company’s additional specified requirements have been met, nor that the pressure vessel is satisfactory for its intended service conditions. The Company’s pressure vessel engineers and inspectors, or contractors acting for the Company, should be involved with the design and construction of vessels to assure that the Company’s requirements are properly met. It is also recommended that Company engineers and inspectors make their own verification that Code rules are complied with whenever design reviews of shop inspections are made.

412 ASME Code, Section VIII, Division 2The ASME Code, Section VIII, Division 2, is more explicit than Division 1 concerning the responsibilities of the owner/user, manufacturer, and authorized inspector. The responsibilities and duties of each party are detailed in Article G-3 of Section VIII.

Owner/User’s ResponsibilitiesThe owner/user, or an agent acting on his behalf, is required to prepare a “UseDesign Specification” that gives “the intended operating conditions in such detato constitute an adequate basis for selecting materials and designing, fabricatinand inspecting the vessel.”

A significant difference from Division 1 is that the owner/user must indicate if a fatigue analysis should be made for cyclic pressure and/or temperature operatiand the owner/user must provide sufficiently detailed information regarding thecyclic conditions to make the fatigue analysis. A fatigue analysis can complicatedesign of a pressure vessel, but the owner/user may exempt the vessel from a fanalysis based upon the successful operation of similar equipment under similaconditions.

Manufacturer’s ResponsibilityThe manufacturer is required to design and construct the pressure vessel to meconditions in the User's Design Specification as well as to comply with all appli-cable requirements of the Code. This requirement of the ASME Code, Section Division 2, differs somewhat from the requirements of Division 1: Division 2 makthe manufacturer responsible for establishing the design requirements to meet service conditions specified by the owner/user; Division 1 leaves the design reqments for the service conditions largely up to the owner/user.

A “Manufacturer's Design Report” must be prepared that includes all the calculations necessary to show that the design as shown on the fabrication drawings complies with the requirements of the Code and meets the conditions in the UsDesign Specification. Analyses of local primary membrane stresses, discontinustresses, and secondary (thermal) stresses should be included wherever they athe design of a component of the vessel. Fatigue analyses should also be incluthey are required by the User's Design Specification.



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Authorized Inspector’s ResponsibilitiesThe Authorized Inspector must review documentation concerning materials certifi-cation, nondestructive examination, the User’s Design Specification, and the Manu-facturer’s Data Report to verify that all requirements of the Code have been met before applying the “U-2” stamp to the nameplate. However, as with the requirements for Division 1, the Authorized Inspector is not responsible for verifying thcompleteness or correctness of the design calculations, or for determining that service conditions in the User's Design Specification have been appropriately addressed in the manufacturer's design. Company pressure vessel engineers ainspectors need to insure that the design and construction of the vessel is apprpriate for the service conditions. Design and construction of a Division 2 vesselusually require greater involvement of Company staff, because the accuracy ofdesign and integrity of fabrication are critically important to the reliable perfor-mance of the vessel.

420 Determining Design ConditionsThe following design conditions for a pressure vessel must be established befoactual design begins:

• Design pressure• Design temperature• Wind and earthquake loads• Corrosion allowance• External loads• Internal loads

Pressure and temperature are the factors that often govern the mechanical desa pressure vessel. The ASME Code requires that a pressure vessel be designethe “most severe condition of pressure and temperature expected in normal option.” Pressure is directly entered into the Code equations for design calculationTemperature indirectly influences the design through its effect on the maximumallowable design stresses for the materials of construction.

Wind and earthquake loads specific to the geographic location where the vessebe installed must also be considered in the design. The greatest effect of wind earthquake is usually upon the support and anchoring of a vessel, with no effecupon the design of pressure containing components, unless the shell is relativethin with respect to its diameter (i.e., a large vessel with a low internal pressurethe contents of the vessel are relatively heavy.

It is usually necessary to add a corrosion allowance to the minimum required thness calculated for each component of the vessel, or to provide a corrosion-rescladding to protect the vessel from corrosion by the process environment.

The major external loads that must be considered are normally those from pipinconnections to nozzles or from structural attachments to the shell and heads, ssupport clips for platforms and ladders.



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Piping is usually designed to have sufficient support and flexibility so the resulting loads applied to a vessel have a negligible effect on the design of nozzles and the reinforcement of the openings in the shell. Similarly, external loads attributable to structural attachments usually have a negligible effect on the design of the pressure-containing components of a vessel.

421 Design Pressure: OverviewThe following pressures must be considered in the design of a pressure vessel:

• Operating pressure• Design pressure• Maximum allowable working pressure (MAWP)

Figure 400-1 shows a typical pressure vessel with the pressures that must be cered, and the relationships among these pressures. These pressures are also itrated in Figure 400-2. The terminology used is consistent with ASME Code, Section VIII, Division 1 and 2.

Fig. 400-1 Representative Pressure Vessel with Definitions of Pressures Considered in Design. See Section 422 for further explanation.

Po(top)MAWP

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Pd = design pressure

Po = maximum operating pressure

Ph = hydrostatic head

∆P = pressure drop

P = component design pressure

P1, 2, 3 = component design pressure at heights 1, 2, & 3

P = Pd + Ph + ∆P

Pd = Po (top) + margin

MAWP = Pd + correction (when applicable)



Fig. 400-2 Schematic Illustration of Various Pressures. (Not to scale.) See Section 422 for a complete detailed explanation.



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422 Design Pressure: ASME Code, Section VIII, Division 1This section describes the various pressures in more detail, and discusses how they are determined and used for the design of a pressure vessel.

Maximum Allowable Working PressureThe maximum allowable working pressure (MAWP) is required to be displayed on a pressure vessel's nameplate, and is defined in the Code as, “the maximum prepermissible at the top of the vessel in its normal operating position at the (desigtemperature.” The MAWP is not the same as the design pressure (Pd), which provides the basis for the design of the vessel. The MAWP is determined from design of the vessel as described below, and is not used for the design.

Design Pressure and Operating PressureThe maximum operating pressure (Po), specified by process design engineers, is thmaximum internal pressure that will occur under normal process conditions. Thdesign pressure (Pd) is determined by adding a margin to the maximum operatingpressure (Po) to allow for pressure surges above Po (without lifting the pressure safety relief valve).

Po should be increased by the following minimum margins to obtain Pd.

These margins may have to be increased if the process designers anticipate hipressure surges because of the characteristics of the process.

Design Pressure and Pressure Used for Design of Individual ComponentsAlthough the design pressure (Pd) establishes the basis for the design of a vessel,is not always the pressure used for designing components of the vessel. The psure at the bottom of a vessel containing a liquid is higher than the pressure at top due to the hydrostatic head. Therefore, the hydrostatic head (Ph) from the top of the liquid to the component being designed must be added to Pd to establish the component design pressure (P), which is used in the Code design calculations that component (i.e., P = Pd + Ph for liquids).

The gas pressure at the top of a vessel that contains a liquid can be thought ofdesign pressure (Pd). If a vessel contains only gas, the hydrostatic head (Ph) is negli-gible, and the design pressure (Pd) can be used as the pressure for the design of aof the vessel's components (i.e., P = Pd + 0 for gases).

Usually, only one value for P is determined for the bottom of the vessel (i.e., formaximum hydrostatic head). That number is used for the design of all componeof the vessel from top-to-bottom. However, for tall vessels at low internal pressu

Po, psig Margin

0—170 25 psi

170—300 0.15 × Po

300—450 45 psi

450 + 0.10 × Po



(i.e., Ph is large relative to Pd), it may be advantageous to design individual compo-nents for the actual hydrostatic heads that exist at various levels in the vessel. The shell thickness can be reduced toward the top of the vessel, which in turn reduces the material and fabrication costs and the cost of the vessel support structure, because of the reduced vessel weight and wind/earthquake overturning moment.

Pressure DropThe process design may indicate a large pressure drop (∆P) through the vessel that should be considered in the design of each vessel component. If ∆P is significant, it should be added to establish the component design pressure (P):

P = Pd + Ph + ∆P(Eq. 400-1)

Calculation of Maximum Allowable Working PressureThe actual thicknesses of the various vessel components will usually be thicker than the thickness calculated using the component design pressure (P). It is usually more economic to obtain the required thickness plus corrosion allowance by purchasing the next thicker commercial size of plate, pipe, or ANSI B16.5 flange, than to have the components specially fabricated to the exact thicknesses required. Therefore, the MAWP permitted (without exceeding the maximum allowable design stress) for the material at the design temperature, will usually be somewhat higher than the design pressure (Pd). The Code allows calculating a MAWP based on this extra thickness, adjusted for the hydrostatic head (Ph), for each vessel component, using the lowest MAWP for any component as the MAWP for the vessel.

If a MAWP is not calculated for the actual component thicknesses in this described manner, the design pressure (Pd) must be used for the MAWP on the nameplate. When the design pressure (Pd) is used for the MAWP on the nameplate, the extra thickness should be added to the corrosion allowance for each component of the vessel.

The MAWP of a vessel should not normally be limited by the MAWP of a minor component, such as a flange or nozzle. For example, if an ANSI B16.5 flange has a lower pressure rating than the MAWP for the shell and head components, the flange should be upgraded to the next higher class. However, this upgrading can cause complications if the associated piping class calls for lower pressure flanges, and a nonstandard flange must be added to the pipe mating to the vessel. These factors must be evaluated for each specific circumstance.


Under Division 2, the pressures used for the design of each component are very similar to those discussed for Division 1, although terminology differs. The maximum pressure permissible at the top of the vessel at the design temperature is defined as the design pressure (Pd), and is displayed as such on the vessel’s name-plate. However, similar to Division 1 requirements, this pressure may not be the



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pressure that is used for the design of the components of the vessel. The hydrostatic head (Ph) must be added to Pd to determine the pressure (P) used for the design of each component (i.e., P = Pd + Ph). Division 2 requires that the coincident pressure (i.e., Pd + Ph) at any point should be used for the design of a component.

Ph for a Division 2 vessel is usually very small in relation to the high Pd that makes the use of Division 2 requirements economic, and the highest Ph at the bottom of the vessel can be used for the design of all components, regardless of the liquid level, in the vessel, without cost penalty.

Division 2 does not refer to a MAWP. Each component of a pressure vessel will usually be fabricated from custom-produced materials close to the minimum thick-ness plus corrosion allowance, with the exception of ANSI B16.5 flanges. There-fore, no benefit will be obtained by calculating a MAWP for the actual thicknesses.

424 External PressureMost vessels are designed to contain positive internal pressure, but it is possible to develop a partial vacuum inside of the vessel during steam-out cleaning, when draining liquids, or during abnormal process conditions. Therefore, it is general practice to check all vessels that are designed for an internal pressure for their resis-tance to collapse under 7.5 psi external pressure at 450°F. This will normally afonly relatively large diameter vessels that are designed for relatively low internapressures. If the minimum required thickness for the internal pressure does notprovide adequate resistance to collapse under a partial vacuum, it is usually moeconomic to add stiffening rings to the shell rather than to increase the thicknesthe other hand, if a vessel design exceeds the required minimum external presrating of 7.5 psi, the vessel should be stamped for the higher pressure, up to 15 psi.

425 ASME Code, Section VIII, Division 1, Design TemperaturesMany pressure vessels are designed to operate at high temperatures, but are urequired to be under pressure at ambient temperature during startup and/or shudown conditions. A few vessels are required to operate at temperatures below ambient temperatures, and others can be subjected to “autorefrigeration” resultfrom operational upsets or gas leaks.

Both Divisions 1 and 2 of ASME Code, Section VIII, require the nameplate of avessel to display both a maximum temperature and a minimum temperature onvessel's nameplate. This has long been a practice for the Company and a requment of Division 2, but it has only recently become a requirement of Division 1,beginning with the 1986 edition.

The maximum temperature controls the vessel design by establishing the maxiallowable design stresses for the selected materials of construction. Selecting mrials based upon maximum temperature assures that the stresses developed insure vessel at the MAWP will not cause failure by ductile bursting or gross yieldduring continuous operation at the maximum temperature.



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The minimum temperature does not directly affect the design of a pressure vessel. However, the minimum temperature can affect the selection of the materials. The materials must have sufficient toughness at the minimum design temperature to prevent failure by brittle fracture during startup and shutdown.

Division 1 defines both maximum and minimum design temperatures, which in effect, establish a temperature range over which the vessel can be operated at the MAWP.

The maximum design temperature as defined in Division 1 “shall not be less ththe mean metal temperature (through the thickness) expected under operatingconditions.” It is displayed on a vessel's nameplate as the temperature for whicthe MAWP is determined. It is the maximum metal temperature permitted at theMAWP.

Establishing the Maximum Design TemperatureThe maximum design temperature is based on the normal maximum operatingtemperature obtained from the process design. The recommended practice is ta margin of at least 25°F above the maximum operating temperature, to assureoperation in the event of minor temperature excursions. Exothermic reactions oother process characteristics that may cause higher temperature excursions wirequire greater margins, and should be discussed with the process designers bthe maximum design temperature is established.

When the maximum operating temperature is below 650°F, it is usually desirabincrease the maximum design temperature to 650°F to provide greater flexibilitfuture changes in operating requirements. This is possible without significantly altering the design of the vessel (i.e., the minimum required thickness for majorcomponents), because the maximum allowable design stress for most of the mrials of construction is the same for any design temperature up to 650°F.

However, the pressure ratings for ANSI B16.5 flanges does change at temperabelow 650°F, and increasing the temperature to 650°F may necessitate upgradthe flanges to the next higher class. If an increase in cost is unacceptable or a lflange rating is desired for matching piping flanges, the maximum design tempeture can be decreased to the rated temperature for the lower flange class.

Maximum design temperatures above 650°F significantly affect the design. Mulower stresses are permitted by the ASME Code for temperatures above 650°Fmost materials of construction. Therefore, the required thickness for pressure-containing components increases with increasing design temperature above 65

Minimum Design Metal TemperatureThe minimum design metal temperature (MDMT) is defined in Division 1 as “thelowest (temperature) expected in service.” This temperature is essentially the sas the minimum pressurizing temperature (MPT) in Company terminology. For mvessels, the minimum design metal temperature is the lowest metal temperaturpermitted at the MAWP. The Company’s current practice is to limit the operatingpressure to 40% of MAWP at temperatures below MDMT.



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This practice (allowing pressurizing to 40% of MAWP when below the minimum design metal temperature) has changed for vessels built in 1999 and later. This is due to the fact that ASME increased the allowable stresses for Division 1 vessels in 1999. For the newer vessels, limit operating pressure to 35% of MAWP at tempera-tures below MDMT. Refer to Section 342 of the Corrosion Prevention and Metal-lurgy Manual for more information, and consult a specialist to confirm if this practice applies to your situation.

The minimum design temperature should never be above 50°F, unless the circustances are discussed with an experienced pressure vessel engineer.

The minimum design metal temperature can be established for a vessel in sevedifferent ways depending upon how the vessel will be operated, as discussed bIt is necessary to assure resistance to brittle fracture at the minimum design temature. This is accomplished by either selecting materials that are known to haveadequate toughness at the minimum design temperature, or requiring CV-impatesting in order to establish that adequate toughness exists. Preventing brittle fture is discussed in Section 500, and the Specifications section. The model specations contain recommendations for procuring equipment with adequate resistto brittle fracture.

It is preferable to use materials with assured adequate toughness as designatethe Code curves described in Section 500. Materials that require CV-impact tescan be substituted if the vessel fabricator advises that they are more economicprocurement will be expedited. But the substitution should be discussed with anexperienced pressure vessel engineer before it is permitted.

Whenever CV-impact testing is required, the tests must be conducted at the minimum design temperature (for the base metal, weld metal, and heat affectedzone) for all weld procedure qualifications. In addition, production test plates mneed to be prepared during fabrication of the vessel to verify that the materials welding consumables actually used will provide the required CV-impact toughn

Operation Below Ambient TemperaturePressure vessels that will be in continuous operation at temperatures below noambient temperatures are considered to be in “critical” service because of the rbrittle fracture. The minimum design metal temperatures for these vessels are bon their normal minimum operating temperatures. The minimum design temperture will usually be designated at approximately 10°F below the minimum oper-ating temperature.

Startup and Shutdown at Ambient TemperaturePressure vessels that are installed at geographic locations with severe winter climates may be required to be at maximum operating pressure at low ambienttemperatures during transient startup and/or shutdown conditions. The minimumdesign metal temperature for this situation should be the lowest ambient wintertemperature that would be expected for a startup or shutdown during unusuallysevere winter conditions. Figure 2-2 of API 650 (available in the Tank Manual) can be used as a guideline for establishing minimum design metal temperature;



however, somewhat warmer temperatures are usually satisfactory for normal start-ups and shutdowns, especially in locations with temperate climates. Engineering judgment should be used to determine the probable minimum temperature during vessel pressurization. The operating pressure must not exceed 40% of the MAWP until the vessel is warmed up to the minimum design metal temperature (35% of the MAWP for vessels built in 1999 and later).

It is necessary to provide adequate resistance to brittle fracture during startup and shutdown at low ambient temperatures. All of the new Code requirements must be adhered to for the minimum design metal temperature. However, transient startup and shutdown conditions are not considered to be critical service, unlike continuous operation at low temperatures.

Restricted Startup and ShutdownMost pressure vessels in locations with a mild winter climate will not require special materials selection or CV-impact testing to assure adequate resistance to brittle frac-ture at their operating pressure during startup and shutdown. The minimum design metal temperature for each component can be conveniently established by deter-mining the minimum temperature permitted by the Code for the material of construction and nominal thickness. The maximum temperature determined in this manner would be designated as the minimum design metal temperature for the vessel. The operating pressure during startup and shutdown should not be allowed to exceed 40% of MAWP at lower temperatures (35% of MAWP for vessels built in 1999 and later), but this should not place prohibitive restrictions upon startup and shutdown procedures.

AutorefrigerationThe autorefrigeration temperature is defined as the temperature that the contents of a vessel would reach if the vessel is depressured to 40% of the MAWP (35% of MAWP for vessels built in 1999 and later). However, the autorefrigeration tempera-ture need not always be used as the minimum design temperature. Autorefrigeration is not considered a critical service condition equivalent to continuous operation at a low temperature, because the loss of pressure that causes the contents of a vessel to autorefrigerate also reduces the operating stresses during autorefrigeration. Further-more, cooling is likely to be highly localized at the source of a leak, or lag appre-ciably behind the temperature of the vessel’s contents during a general system loss of pressure.

Pressure vessels subject to autorefrigeration should be fabricated from SA 516 normalized or SA 537 plate (with equivalent pipe or forging grades). See Section 500 for typical selections. These are materials that have inherently good toughness at low temperatures, and should provide adequate resistance to brittle fracture during a transient condition that causes autorefrigeration. CV-impact testing is not mandatory for the autorefrigeration temperature.

HydrotestingHydrotesting a pressure vessel should never be performed at an ambient tempera-ture below the minimum design metal temperature. The high stresses developed



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Maximum TemperatureDivision 2 requires that the design temperature “shall be based upon the actuametal temperature expected under operating conditions,” and should be determfrom the maximum operating temperature in the same manner as described abfor Division 1 vessels. The design temperature would normally be the maximumtemperature permitted coincident with the design pressure, and is displayed ason the nameplate.

Minimum TemperatureA minimum permissible temperature is required to be displayed on the nameplaa Division 2 vessel. Unlike the nameplate for a Division 1 vessel, a coincident psure is not displayed for the minimum permissible temperature. However, the psure during startup or shutdown is explicitly not permitted to exceed 25% of thedesign pressure (defined in the Code as 20% of hydrotest pressure) at temperabelow the minimum permissible temperature.

The minimum permissible temperature can be determined for a Division 2 vessthe same way the minimum design temperature is determined for a Division 1 vessel, as discussed above. However, the curves in Division 2 concerning CV-impact test requirements for the various materials of construction differ somewhfrom those in Division 1. The curves in Division 2 tend to require CV-impact testat slightly higher temperatures for the same material of construction and compothickness, which is probably related to the greater toughness required to resist fracture at the higher design stresses in Division 2.

427 Wind and Earthquake DesignNote that structural design, including for wind and earthquake, is discussed in Section 440.

The potential wind and earthquake loadings on a pressure vessel are specific tgeographic location where the vessel is installed. Design parameters for the UnStates can be obtained from maps published by API. Refer also, to the Civil and Structural Manual. Section 440 of this manual describes how these design paramters are used to determine the maximum potential loads for design of the vesse

Both wind and earthquake loads create overturning moments that develop longdinal stresses in the shell of vertical vessels. The weight of the internal contentsvessel will amplify the overturning moment resulting from earthquake loading, amust be taken into consideration when calculating the longitudinal stresses. It inecessary to design a pressure vessel for the simultaneous occurrence of maxwind and earthquake loads.



The longitudinal stresses developed in the vessel’s shell by the wind and earthquake loads must be added to the longitudinal stresses attributable to the internal pressure. However, the longitudinal stress attributable to the internal pressure is nominally one-half of the hoop stress, which is the maximum principal stress that governs the design of the vessel for internal pressure. Furthermore, the combined stresses for wind or earthquake and internal pressure are allowed to exceed by 50% the maximum allowable design stress for the material of construction given in the Code.

Higher stresses are permitted because a vessel is only intermittently subjected to severe wind and earthquake loads. Consequently, wind and earthquake loads will not usually affect the design of a pressure vessel shell. The major exceptions would be vessels designed for low pressures with relatively thin shells and a high weight of internal contents. Wind and earthquake loads can have a very significant effect on the design of the support and anchoring for a vertical vessel.

428 Corrosion AllowanceThe internal process environment that a pressure vessel is exposed to during opera-tion can frequently cause the material to corrode. Therefore, a corrosion allowance must be added to the calculated minimum thicknesses required for each component of the vessel. The Code assigns to the owner/user of the vessel the responsibility for specifying the corrosion allowance. This is necessary to prevent corrosion from reducing the thicknesses below the required minimums during operation. Recom-mendations for determining corrosion allowances, based on operating experience, are discussed in Section 500.

An alternative to a corrosion allowance, especially when the process environment would result in a very high corrosion rate, is to employ a corrosion-resistant clad-ding or to apply a corrosion-resistant coating.

External corrosion is rarely significant. Pressure vessels usually either operate at high temperatures that prevent the condensation of moisture, or weather shielding is provided. Therefore, an external corrosion allowance is not usually necessary.

429 External and Internal LoadsSee Section 440 for a more detailed discussion of structural loads.

External LoadsExternal loads applied to a pressure vessel are usually of a local nature. WRC Bulletin 107 provides the almost universally accepted methodology for calculating the stresses that these loads develop in the shell of a vessel, and for determining the effect that they will have upon the design (i.e., the minimum required thicknesses for the affected components of the vessel).

The stresses developed in the vessel’s shell by local external loads must be added to the stresses attributable to the internal pressure. It is important to recognize that they are normally either local primary membrane stresses, or bending stresses. There-fore, as explained in Section 100, the total stress obtained by adding the stresses



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ca- the for tion re

developed by local external loads to those attributable to the internal pressure is permitted to reach 1.5 times the maximum allowable design stress given for the material of construction in the Code. Consequently, only relatively high external loads are likely to affect the design of a vessel.

Piping ConnectionsStresses can be developed in a pressure vessel shell due to forces and moments that result from piping connections to nozzles. The magnitude of these forces and moments applied to a vessel are relatively insignificant and need not be considered for the design of a vessel. This is especially true for vessels designed according to Section VIII, Division 1, where extensive experience has shown that the required Safety Factor of 4 is sufficient to compensate for these relatively small loads without detailed analyses. Heavy equipment attached to nozzles (such as valves and bridles) should be supported to minimize the external loads acting on the nozzle.

The forces and moments attributable to piping connections can be calculated using the computer program CAESAR (see the Piping Manual), if it is suspected that they are high enough to affect the design of the vessel.

Structural AttachmentsPlatforms and ladders are frequently supported by direct attachment to a pressure vessel. The normal practice is to provide a sufficient number of clips welded to the vessel’s shell such that the local load transmitted to the shell by any individual clip is not great enough to affect the design of the vessel. This may not always be prac-tical for vessels with a low design pressure and a relatively thin shell. For these vessels, the external load can usually be distributed over a larger area to reduce the stresses developed in the shell, by providing a reinforcing pad on the shell for attachment of the clip. However, it is not advisable to use a reinforcing pad if the operating temperature of the vessel will exceed 450°F, and other design approashould be investigated.

Attached EquipmentOther equipment, such as reboilers or valves, occasionally are connected direcpressure vessel nozzles without providing separate support. The weight of this equipment results in forces and moments on the nozzles that develop local strein the vessel's shell similar to the stresses caused by piping connections. If thestresses are great enough to affect the design of the nozzle and reinforcement opening, when added to those attributable to internal pressure, additional suppothe attached equipment should be considered.

Lifting LugsLifting lugs must be provided for moving and erecting a pressure vessel. The lotion of these lugs and the loads that will be applied to them depend on how thevessel will be moved and erected. The details are usually worked out between fabricator of the vessel and the construction contractor who will be responsible erection. A rigging diagram should be provided to the fabricator by the construccontractor. The static loads on the lugs resulting from the weight of the vessel ausually multiplied by a factor, depending on the construction contractor's lifting



procedure and experience, to accommodate dynamic loads that can be developed during the lifting.

There are no Code criteria for maximum allowable stresses that can be used for design of the lifting lugs. They are designed to prevent damage to the vessel during installation only, and they have no effect on the integrity and reliability of the vessel during operation.

Internal LoadsPressure vessels normally contain various internal components that are attached directly to a vessel’s shell, such as distributor trays, catalyst support grids, baffles, and demister pads, etc. These internal components apply loads to the shell, and thereby develop stresses that must be added to those resulting from the internal pres-sure. The weight of the internal components plus the weight of liquid or catalyst supported by the component must be considered. In addition, the pressure drop across the component will apply an additional load to the shell that must be consid-ered separately from the influence of the pressure drop on the design pressure.

The internal loads in vertical pressure vessels are usually downward, developing a compressive stress in the vessel’s shell that counteracts the longitudinal tensile stress developed by the internal pressure. Therefore, it is rare that internal loads affect the design of a vertical vessel. An exception could be encountered with an upflow vertical vessel, if a high pressure drop occurs across an internal component. This would develop a tensile stress in the shell that would add to the longitudinal stress developed by the internal pressure. Note that the weight of the internal contents of a vessel (i.e., internal components, catalyst, and fluids, etc.) will affect the design of the vessel’s support, both by directly increasing the compressive stress and indi-rectly by amplifying the overturning moment in the event of an earthquake.

430 Mechanical Design

431 Background Information and Overview

Use of the ASME CodeVessels must be designed to the latest edition of the ASME Boiler and Pressure Vessel Code, Section VIII, and addenda. This section discusses only Section VIII, Division 1 rules. For application of Division 2 rules, consult an experienced pres-sure vessel engineer.

Note This section will help you understand the Code and how to apply it. It does not replace the Code; therefore, you should not attempt designing a vessel without the Code.

There are two general approaches to designing and ordering a pressure vessel. One way is to specify only the required service conditions, nozzle sizes, and orientations to the fabricator and to rely on him to determine all of the design details of the vessel construction.



Another approach is for the owner to determine many of the construction details prior to ordering the vessel. Although this approach causes the owner more up-front design engineering, it is often used when vessel weights are required for estimating purposes, or when other design work requires some detailed knowledge of the vessel construction, before the bidding phase.

PVM-EF-65, in the Standard Drawings and Forms section, is useful for tracking the basic computations required for pressure vessel design.

All Company pressure vessels are designed to the ASME Code, Section VIII, even in states or foreign countries which have not adopted the ASME Code. Further-more, you must also comply with any additional requirements of the jurisdiction(s) governing the location where the vessel will be installed.

The ASME Pressure Vessel Code, Section VIII, covers all pressure vessels other than those required to be in accordance with Sections I, III, or IV. Fired process tubular heaters are specifically excluded. See Section 200 for more information on the application of the Code.

Vessels excluded from Section VIII are those where the internal design pressure does not exceed 15 psig with no limitation on size, and those having an inside diam-eter not exceeding 6 inches, with no limitation on pressure.

Piping Versus Pressure VesselsFrequently it is difficult to decide if a unit being incorporated into a piping system should be classed as piping or as a pressure vessel. No precise differentiation exists; hence the classification must be left to engineering judgment.

The unit should be called piping, however, if:

1. As a part of the piping system, its primary function is to transport fluid from one location to another within the system. A header or manifold for the distri-bution of fluid would fall into this category. Special design features or accesso-ries added to permit secondary functions would not change its classification. For example, enlargement of any part or all of a header to provide a degree of pulsation dampening, or to accumulate and remove liquid from a gas in connec-tion with its primary function, would not, in itself, classify the piping system or any part of it as a pressure vessel.

2. The element under consideration is available from, and is classified by recog-nized piping equipment suppliers as a piping component or accessory. Exam-ples would include certain types of strainers, filters, steam traps, steam separators, expansion joints, and metering devices. Units which are normally constructed in accordance with the Code, however, should not be included in this category, but classified as pressure vessels.



Even if fabricated exclusively from pipe and fittings, a unit other than a commercial piping accessory should be classified as a pressure vessel, if:

1. Its primary purpose is not to transport fluid, but to process fluids by distilla-tion, heat exchange, separation of fluids, or removal of solids.

2. Its primary function is to store fluids under pressure.

Pressure/Temperature LimitationsFigure 400-3 shows the design pressure and design temperature range through which the ASME Code, Section VIII, Division 1, applies. Design codes outside of this range are also shown.

Design ConsiderationsIt is important to understand that Code rules are formulated to provide minimum requirements for safety and service, and that the Code holds the owner/user and the designer/manufacturer responsible for meeting Code rules as well as the service needs. This concept is discussed in detail in Section 410.

An ASME Code stamp on a pressure vessel only provides that new and cold, the pressure vessel is good for the pressure and temperature indicated and recorded on the manufacturer’s ASME data report. The stamp also permits the legal operation of the pressure vessel in a jurisdiction that has adopted the ASME Code as a part of the law.

The ASME Code cannot include rules for all types of pressure vessels and all types of services. This is particularly true for Section VIII, Division 1, since it covers a wide scope of pressure vessels from simple water and/or air receivers constructed of low carbon steel to process pressure vessels constructed with nonferrous, high-alloy,

Fig. 400-3 Scope of Various Design Codes and Standard



de of e

and/or high-strength materials. Section VIII, Division 2, provides for the design of pressure vessels subjected to external and internal pressure without any limit on the pressure. The design temperature is limited by the design stress for the material listed in the applicable stress table or Code case covering the construction.

Because of the range of materials and potential services for pressure vessels constructed under Section VIII, Division 1, it is understandable that the Code must hold the owner/user and the manufacturer responsible for constructing a pressure vessel for a specific service, as discussed in Section 410.

The owner/user’s designated agent may be an engineer on his staff, a design agent specifically engaged by the owner/user, the manufacturer and/or designer of a system for a specific service, an organization which offers pressure vessels for sale or lease for a specific service, or the engineer of the stamp holder that will manufacture the pressure vessel. Design considerations must include, but need not be limited to, the following:

1. The need for corrosion allowance beyond that specified by the rules. Except for a limited number of vessels constructed of low carbon steel containing air, steam, or water, corrosion allowance is not a requirement of the Code.

Corrosion allowance is the extra wall thickness available beyond the thickness computed by Code rules. This thickness is assumed to be lost to corrosion at the end of the vessel design life. (Section 500)

2. The contents defined with particular attention to toxic substances in order that the mandatory requirements of the Code can be met.

3. The need determined for impact testing of base materials and deposited weld metal if the vessel will operate at low temperature. This is to assure adequate toughness. (Section 500)

4. The need for postweld heat treatment beyond Code mandatory requirements that might be necessary for resistance to corrosion or because of the service environment. (Section 600)

5. Proper selection of the materials for the service conditions. (Section 500)

6. Nozzle locations and the external piping reactions that will occur on the nozzles during normal operation.

7. The installed orientation of the pressure vessel and the type of support required.

8. For pressure vessels generating steam, the need for piping, valves, instrumenta-tion, and fittings required for the service as set forth in Section I, ParagraphsPG-59 through PG-61.

9. The number of openings and the size for the required safety relief devices.

10. Selection of vessel heads—Very high compressive stresses can exist in theknuckle region of torispherical heads under internal pressure. The magnituthe stress is primarily a function of the knuckle radius, the diameter, and thdiameter-to-thickness ratio. (Section 100)



r on of

r

er rit-

ll age

d

s, re. ngth

nd

11. Openings in knuckles—Caution should be exercised in locating openings oother attachments that may produce large stress concentrations in the regithe knuckle of torispherical heads or the shell-to-cone junction of conical transitions.

12. Structural, wind and earthquake—The Code does not have design rules fowind or earthquake loads. Therefore, design for wind and earthquake loadsshould be in accordance with the UBC or CIV-EN-100, the recommended Company design standard for wind and earthquake. (Structural design is discussed in Section 440.)

Recommended PracticesCode rules establish minimum construction standards for a pressure vessel. Because of the extremely wide scope of Section VIII, Division 1, however, highquality standards are usually justified, except for the simplest of vessels in noncical service:

• UM vessels (these pressure vessels have a volume and pressure limit).

• Utility pressure vessels, such as the air receivers supplied as part of a smanoncritical air compressor package, and small noncritical refrigeration packequipment.

The principal areas where Company model specifications contain recommendeprovisions above Code requirements are:

• Protection against brittle fracture• Sour service requirements• Types of welded joints• Nozzles• Limits on using reinforcing pads• Limits on using torispherical heads• Limits on certain welding procedures• Radiography• Stress relief of cold-formed parts• Preheat and postweld heat treatment requirements• Hydrostatic tests

Protection Against Brittle FractureDesign for brittle fracture is covered in detail in Section 500. At low temperaturestrength is less a consideration than the increasing susceptibility to brittle fractuBrittle fractures generally occur without warning at stresses below the yield streof the material.

Carbon and low-alloy steels are subject to brittle fracture at low temperatures aeven at temperatures significantly above ambient, depending on the thickness,chemistry, method of manufacture, and other factors. The ASME Code, Section



“wet

nts ecifi-ing ion, 750 a-

rrant es, epti-

w-s

This

ip er- of

gura-, or

ult in fter

ding the essel

o be

VIII, Division 1, rules regarding material toughness for carbon steels were imple-mented in January 1988.

Wet H2S ServiceWet H2S cracking can occur at carbon steel equipment exposed to aqueous or H2S” service. Wet H2S is defined as service conditions of at least 50 ppm of H2S dissolved in a liquid water phase. Streams with higher H2S or with cyanides are considered more severe. In addition to specific repair and inspection requiremefor this service, the construction of pressure vessels for this service requires spcations formulated to minimize the problems associated with this service includpost weld heat treatment and wet fluorescent magnetic particle testing. In additit is possible to use HIC resistant steels for these services. Specifications EG-4and EG-4749 provide additional requirements for sour service. For more informtion, refer to the Wet H2S Notebook available from CRTC.

Postweld Heat Treatment for Service ConditionsSee Section 600 for specific recommendations where service environments wapostweld heat treatment (PWHT). PWHT is performed to reduce residual stresshold dimensional tolerances, improve resistance to brittle fracture, reduce suscbility to stress corrosion cracking, or provide added safety in certain hazardousservices.

In addition to thickness requirements, the Code requires PWHT for carbon or loalloy vessels operating below minus 50°F, for unfired steam boilers at pressureabove 50 psig, and for lethal-service vessels. Where service conditions might promote stress corrosion cracking, the need for PWHT should be investigated. subject is discussed in the Corrosion Prevention and Metallurgy Manual.

Hydrostatic TestsThe purpose of hydrostatic testing is to reveal defects in the vessel workmanshand to detect the presence of leaks. Note that most weld defects cannot be detmined by the use of hydrostatic testing and must be determined by other formsnondestructive testing.

Code rules permit an initial hydrostatic test pressure to be set at 1.5 times the maximum allowable working pressure (MAWP), or the design pressure to be stamped on the vessel. This value may be adequate for vessels of simple confition used in noncorrosive services. However, for large or other complex vesselswhen a significant corrosion allowance is provided, this pressure rating can resoperating the vessel at stresses above those induced by the hydrostatic test (acorrosion has reduced the thickness of the vessel wall).

Hydrostatic test pressure should be calculated to stress the full thickness, inclucorrosion allowance and cladding of the strongest Category A weld to 150% of Code allowable stress, at the test temperature, provided no component of the vis stressed above 90% of the minimum yield strength. The pressure should alsreduced as necessary to avoid overstressing weaker Category A welds.



Industry experience indicates that, for test purposes, an upper limit on stress equal to 90% of the minimum specified yield strength of the base metal can be used provided no excessive stress raisers are present. Common stress raisers are: improp-erly reinforced openings; poorly designed transitions between shell sections and between shells and heads (especially in the case of cone-to-cylinder intersections); or torispherical heads. Designs for especially large or thick vessels or unusual details should be reviewed by a stress analyst, especially in cases where test stresses approach the yield strength. Hydrostatic testing must also be conducted above the minimum design metal temperature (to avoid brittle fracture).

If certain Category A welds cannot be fully tested as outlined above without over-stressing other parts, the following alterations should be made: first, heads and tran-sition sections should be thickened; and second, Category A welds which still cannot be fully tested should be 100% radiographed.

☞ Warning Hydrostatic or pneumatic pressure testing is dangerous and has caused many injuries. Testing, therefore, must be conducted by those trained to do so. The following guidelines are important to observe:

1. Avoid being within the vicinity of any pressure test. Let the authorized inspec-tors and those with direct responsibility for testing the vessel read the pressure gages and check for leaks.

2. Stay away from the vessel when it is being pressurized.

3. Never make a close-up visual examination for leaks, particularly when the vessel is being pressurized. Authorized personnel can make leak checks, after the vessel has been depressurized to design pressure, using safe methods. High pressure leaks eject high velocity jets of gas/liquid mixtures that can penetrate deep into the skin or eyes. Many persons have been blinded by visual leak checking without following proper precautions.

4. Most fabricators locate the hydro pump and its pressure gage at some nominal distance away from the vessel during the testing procedure so they need not get too close to the vessel. Stay away from the pump and its gage and tubing as well. Sometimes fittings break off or the gages leak, and anyone in the path of the leak or fittings may be injured.

5. Catastrophic failure during pressure testing is a possibility. There is a common belief that because the liquid compressibility is low, the stored energy and the energy released during a vessel failure is low. On the contrary, significant energy in hydrostatic tests comes from the compressibility of the water, the expansion of the vessel, energy stored in dissolved gases, and most signifi-cantly, from trapped air that may not have been removed. Photographs of hydrostatic tests intentionally carried to destruction indicate that nearby objects or buildings would be significantly damaged.

6. The energy of pneumatic testing is far greater than hydrotesting because of the high compressibility of air. Therefore, avoid being within the vicinity of any vessel under pneumatic testing.



Design TemperatureSelection of design temperature should provide: (1) maximum flexibility for future changes in operating requirements at a minimum cost premium; and (2) minimum penalty to structural thickness requirements to provide this flexibility. Refer to Section 425 for guidance on establishing design temperatures.

432 Design SummaryThe following is a summary of the mechanical design procedure:

1. Determine the service, contents, and conditions for the vessel.

2. Determine the design pressures and temperature per Section 420.

3. Divide the vessel into its Code components. These include the shell, heads, transitions, and nozzles.

4. Identify the weld efficiency for each weld in each component (discussed later in this section).

5. Establish the material’s allowable stresses at design temperature.

6. Establish a corrosion allowance. See Section 500 for guidance.

7. Use Code procedures to size components for all loads. Code equations are summa-rized in Figure 400-4 for frequently used components under internal pressure.

8. Review the appropriate design considerations, and the need for higher-than-Code provisions.

9. Determine adequacy of external pressure. The procedure for determining adequacy for external pressure is more complex than for internal pressure, but should be performed for all vessels subject to partial vacuum.

10. Integrate and document the design. Design Data Sheets PVM-EF-65 and PVM-EF-66, included in this manual, may be used for this purpose. They are also useful as checklists to help ensure covering all the needed details.

433 Design for Internal PressureFigure 400-4 gives equations for various components under internal pressure. Internal or external pressure produces a longitudinal-seam stress that is two times larger than that on the circumferential seams. For this reason the equations are shown only for longitudinal seams. However, if widely varying joint efficiencies for the circumferential and longitudinal seams are chosen, or other stresses, such as structural stresses, come into play, then the circumferential seam may govern. This circumstance should be checked individually for each vessel under consideration. For thick wall vessels where the wall thickness exceeds one half of the inside radius, or P exceeds 0.385 SE (S and E are defined in Figure 400-4), the equations given by Code Appendices 1 and 2 should be applied.



Example: Cylindrical Shell Under Internal PressureASME Code, Section VIII, Division 1, uses an approximate equation for deter-mining the required thickness of a cylinder subjected to internal pressure:

(Eq. 400-2)

where:t = Required thickness of shell, in

P = Internal design pressure, psi

tPR

SE 0.6P–------------------------=



Fig. 400-4 Equations for Vessels under Internal Pressure From Pressure Vessel Handbook by E. F. Megyesy. Courtesy of Pressure Vessel Handbook Publishing, Inc.

a = Half Apex Angle of Cone, Deg.

D = Inside Diameter, Inches

Do = Outside Diameter, Inches

E = Efficiency of Welded Joints

L = Inside Crown Radius, Inches

Lo = Outside Crown Radius, Inches

M = Factor, See Table Above

P = Design Pressure or Maximum Allowable Pressure, psig

R = Inside Radius, Inches

Ro = Outside Radius, inches

S = Stress Value of Material, psi

t = Thickness, inches

Formulas for Vessels Under Internal PressureNotation



spot

rd . The ing

S = Allowable stress, psi

R = Inside radius of cylinder, in

E = Joint efficiency factor, dimensionless

Consider a vessel with an inside diameter D = 96 inches. The design pressure P = 100 psig.

Material: SA 515-70 plate

Design Temperature: 100°F

Corrosion Allowance (c.a.): 0.125 inch (corroded I.D. = 96.250 inches.)

All circumferential and longitudinal seams are double-welded butt joints and are radiographed. The vessel is to be built per ASME Code, Section VIII, Division 1. From Table UCS-23 of the Code, for SA 515-70 at temperatures up to 650°F, S = 17,500 psi. For spot radiographed joints, from Table UW-12, E = 0.85.

(Eq. 400-3)

Use 0.5 inch plate.

Spherical Shells and Hemispherical Heads Under Internal PressureThe ASME Code equation for spherical shells under internal pressure is:

(Eq. 400-4)

The calculated minimum thickness of formed heads is not rounded up to standaplate because of the thinning that occurs in portions of the head during formingcalculated value should be the minimum thickness at any point on the head. Usthe same vessel example as above:

(Eq. 400-5)

The calculated thickness should be increased by corrosion allowance:

t = 0.162 + 0.125 (c.a.) = 0.287 in.

t100 48.125×

17 500, 0.85 0.6 100×–×---------------------------------------------------------------=

0.325 in. + 0.125 in. (corrosion allowance)=

= 0.425 in.

tPR

2SE 0.2P–---------------------------=

t100 48.125×

2 17 500, 0.85 0.2 100×–××------------------------------------------------------------------------ 0.162 in.= =



Semiellipsoidal Heads Under Internal PressureFor 2:1 semiellipsoidal heads, the ASME Code equation is:

(Eq. 400-6)

where D is the inside diameter of the head.

Example: Using the same vessel as in previous examples: E = 1.0 for a seamless head.

= 0.275 + 0.125 (c.a.) = 0.400 in.(Eq. 400-7)

Torispherical Head Under Internal PressureFor torispherical heads where the knuckle radius is 6% of the inside crown radius, the ASME Code equation is:

(Eq. 400-8)

Using the same example:

L = 96 inches

E = 1.0 (seamless head)

= 0.243 + 0.125 (c.a.) = 0.368 in.(Eq. 400-9)

434 Design for External PressureFigure 400-5 shows the general procedure for determining the maximum allowable external pressure on a vessel. This flow chart is based on the requirements of Code Appendix 5.

Cylindrical Shell Under External PressurePer Code Paragraph UG-28(f), vessels intended for service under external working pressures of 15 psi or less and which are to be stamped with the Code symbol are

tPD

2SE 0.2P–---------------------------=

t100 96.25×

2 17 500, 1 0.2 100×–××-----------------------------------------------------------------=

t0.885LM

2SE 0.2P–---------------------------=

t100 96 0.885××

2 17 500, 1 0.2 100×–××-----------------------------------------------------------------=



Fig. 400-5 Maximum Allowable External Pressure on Cylinder



e

designed for a maximum allowable external pressure of 15 psi, or 25% more than the maximum possible external pressure, whichever is smaller.

The ASME procedure for cylinders under external pressure is described in detail in Paragraph UG-28. It has two different alternatives, for ratios Do/t ≥10 and Do/t <10.

where:Do = External diameter of the vessel

t = Minimum required thickness of the cylindrical shell

L = Design length of a vessel section determined as indicated in Paragraph UG-28(b)

Usually, the thickness is first determined from the internal pressure then the external pressure calculation is used to determine if this thickness is adequate for the external design pressure, or if additional stiffening of the shell is required. If the thickness of the shell has to be determined to start the procedure, a value for t must be assumed.

The Code procedure then goes through several steps for each alternative in order to find the value of factor “B,” determined by using a general geometric chart (Figure 5—UGO-28.0 in Appendix 5) and a material chart for maximum design metal temperature (Appendix 5).

For Do/t ≥10, the maximum allowable external pressure is determined using theequation:

(Eq. 400-10)

For Do/t <10, two values, Pa1 and Pa2 should be determined and the lower value isused:

(Eq. 400-11)

(Eq. 400-12)

where S is the lesser of two times the allowable stress in tension or 0.9 times thyield strength of the material.

Example. Using the same vessel as in the internal pressure calculation:

• Tangent to tangent length: 36 ft 0 in. = 432 in.• Two 2:1 semiellipsoidal heads• External design pressure: 15 psig at 500°F

Pa4B

3 Do/t( )------------------=

Pa1

2.167Do/t( )

--------------- 0.0833– B=

Pa2

2SDo/t( )

--------------- 1 1/ Do/t( )–[ ]=



or A

e

s is

es-or this load alcu-

L = 448 inches (length of shell plus one-third of the depth of each head, 16 in.)

L/Do = 448/97 = 4.62

Do/t = 97/0.375 (corroded) = 258.67

From geometrical chart Figure 5, UGO-28.0, A = 0.00007. From Figure 5, UCS-28.2 for SA 515-70 at 500°F, the modulus of elasticity of material, E = 27,000,000 psi; and f= 0.00007, the value falls to the left of the applicable temperature line. Then Pa = 2AE/3 (Do/t) = 2 × 0.00007 × 27 × 106/3 × 258.67 = 4.87 psi.

The vessel is good for only 4.87 psi external pressure, so stiffening rings are required.

Try two stiffening rings equally spaced between tangent lines.

L = 144 (length of shell between rings) + 8 (1/3 depth of head) = 152 in.

L/Po = 152/97 = 1.56

A = 0.00022 from chart.

B = 2800 from Figure 5, UCS-28.2

(Eq. 400-13)

The vessel is still not good for 15 psi, so another ring should be added. Try threstiffening rings equally spaced between tangent lines.

L = 108 + 8 = 116 in.

L/Do = 116/97 = 1.19

A = 0.00027 from chart

B = 3700 from chart Figure 5, UCS-28.2

(Eq. 400-14)

Since Pa is greater than the design pressure, the vessel with three stiffening ringgood for full vacuum (15 psi).

Design of Stiffening RingsThe equations for the maximum strength of a cylindrical shell under external prsure are established under the assumption that the shell is simply supported. Fto be true, stiffening rings are used as lines of support, assumed to carry all thethat the shell carries due to external pressure. The size of the stiffening ring is c

Pa4B

3 Do/t( )------------------

4 2800×3 258.67×------------------------- 14.4psi= = =

Pa4 3700×

3 258.67×------------------------- 19.07psi= =



lated by using the equation for buckling of a circular ring under uniform external pressure.

The required moment of inertia of the stiffening ring is:

(Eq. 400-15)

where:Ls = The sum of one-half the distances on both sides of the stiffening

ring from the centerline of the ring to the (1) next stiffening ring, (2) to a circumferential line on the head line at one-third its depth, or (3) to a jacket connection.

As = Cross-sectional area of the stiffening ring

The ASME Code allows a portion of the shell to be considered as contributing to the moment of inertia. The width of the shell section is 1.1 Dot, and is considered as laying one-half on each side of the centroid of the ring.

In this case, the required moment of inertia of the combined area is:

(Eq. 400-16)

The Code procedure for determining the size of the ring is as follows:

Example. Using the same vessel as in the previous example, select a 5 x 3 x 3/8 angle ring with As = 2.86 in2 and I = 7.37 in4

(Eq. 400-18)

IDo

2/Ls t As/Ls+( )A

14-------------------------------------------------=

Step 1. Assume a ring size. Determine As

Step 2. Calculate Factor B

(Eq. 400-17)

Step 3. From the material chart in Appendix A, find the value of A.

Step 4. Calculate the value of Is. The moment of inertia of the ring or the combined section (ring plus shell section) has to be larger than the required moment of inertia.

IDo

2Ls t As/Ls+( )A

10.9------------------------------------------------=

B 3/4( )PDo

t As/Ls+----------------------=

B 3/4( ) 15 97×0.375 2.86/116+----------------------------------------- 2 730,= =



From Figure UCS-282, A = 0.0002

= 6.23 in4

(Eq. 400-19)

Since the available moment of inertia is larger than the required moment of inertia, the vessel is adequately stiffened.

Spherical Shell and Hemispherical Head Under External PressureThe ASME Code procedure is as follows:

1. Calculate the value of A using the equation:

(Eq. 400-20)

where:Ro is the outside radius of the sphere.

2. Find the value of B from the Code material/temperature chart in Appendix 5.

3. Calculate the maximum allowable external pressure:

Pa = B / (Ro / t)(Eq. 400-21)

or for values of A falling to the left of the applicable temperature line:

Pa = 0.0625E / (Ro/t)2

(Eq. 400-22)

Using the same example:

Ro = 48.5 in.

t = 0.125 in.

(Eq. 400-23)

From Figure UCS-28.2 B = 10,500

(Eq. 400-24)

The hemispherical head is good for the external design pressure of 54 psi.

Is97

2116 0.375 2.86/116+( ) 0.0002×××

14---------------------------------------------------------------------------------------------------=

A0.125Ro/t

-------------=

A0.125

48.5/0.187-------------------------- 0.00013= =

Pa10 500,

48.5/0.187-------------------------- 54.13= =



Semiellipsoidal Head Under External PressureAccording to the ASME Code, the required thickness for a semiellipsoidal head under external pressure should be the greater of the following:

1. The thickness as calculated by the equation given for internal pressure using a design pressure 1.67 times the external pressure and joint efficiency, E = 1.00.

2. The thickness by the equation Pa = B/(R/t) where R = 0.9 Do and B to be deter-mined for a sphere.

Torispherical Head Under External PressureThe required thickness is computed by the procedures given for ellipsoidal heads using a value for R = Do.

435 Example of Internal/External Pressure DesignFigure 400-6 illustrates how to use the Code to check a vessel for internal and external pressure.

Fig. 400-6 Example: Internal/External Pressure (1 of 3)

Problem

Determine the minimum required thickness of a cylindrical shell and hemispherical heads of a welded pressure vessel designed for an internal pressure of 100 psi at a design temperature of 250°F. There is no corrosion. The shell, which contains a longitudinal butt weld, is also butt welded to seamless heads. All Category A butt joints are Type (1) with full radiography (RT). E = 1.00 for all calculations. The shell has a 5 foot 0 inch inside radius and is 30 foot 0 inch long from tangent to tangent.

Also determine the minimum required thicknesses of the same vessel designed for an external pressure of 15 psi at 100°F without stiffening rings. What is the stiffening ring spacing if the required thickness of internal pressure is used?

Solution

1. For SA 515 Gr. 60, the allowable tensile stress from Table UCS-23 at 100°F is 15.0 ksi, and the external pres-sure chart is Figure 5-UCS-28.2.

2. As is generally the case for internal pressure on a cylinder, when E = 1.00 for all butt joints, UG-27(c)(1) for circumferential stress (hoop stress) controls over UG-27(c)(2) longitudinal stress by:

Check for applicability of using UG-27(c)(1):

Is t < R/2? 0.400 in. < 30 o.k.

Is P < 0.385 SE? 100< 0.385 (15,000)(1) = 5,775 o.k.

3. For internal pressure on hemispherical heads, use UG-32(f):

Is t < 0.365 L? 0.200 < 0.365 (60) = 21.9 o.k.

t PRSE 0.6P–----------------------- 100( ) 60( )

15 000, 1.× 0( ) 0.6 100×( )–------------------------------------------------------------------- 0.401 in.= = =

t PL2SE 0.2P–-------------------------- 100( ) 60( )

2 15 000 1.0×,( ) 0.2 100×( )–---------------------------------------------------------------------- 0.200 in.= = =



Is P < 0.665 SE? 100 < 0.665 (15,000)(1) = 9,975 o.k.

4. For external pressure on cylinder, use UG-28 and Appendix 5:

For cylindrical shells with formed heads on the end the length of the shell plus 1/3 of the depth of each head is used to determine the effective lengths (L) (see UG-28).

Determine the effective length without stiffening rings = 1/3 of each head depth plus straight length = (1/3)(2)(60) + 360 = 400 in.

Assume tmin. for internal pressure of 0.400 in. and Do = 120 + 2(0.4)

L/Do = 400/120.8 = 3.31

Do/t = 120.4/0.4 = 301

a. Enter Figure 5-UGO-28.0 with L/Do= 3.31 and read across to sloping line of Do/t= 301. Read A = 0.000075

b. Enter Figure 5-UCS-28.2 with A = 0.000075 and the modulus of elasticity E = 29.0 × 106 which is off the left side and cannot be read. Following Step (7) of UG-28(c):

Assume t = 5/8 in. = 0.625 in. and Do = 120 + 2(0.625) = 121.25 in.

L/Do = 400/121.25 = 3.30; Do/t = 121.25/0.625 = 194

a. From Figure 5-UGO-28.0, A = 0.00014

b. Recalculate Pa:

Assume t = 11/16 in. = 0.6875 in. and Do = 120 + 2(0.6875) = 121.375 in.

L/Do = 400/121.375 = 3.30

Do/t = 121.375/0.6875 = 177

A = 0.00017

Further calculations show that tmin. = 0.64 in. for 15.0 psi external pressure.

5. For external pressure on hemispherical head, use UG-33(c), UG-28(d), and Appendix 5.

First assumption, use tmin. for internal pressure of t = 0.200.

Assume t = 0.200 in. and Ro = 0.5(120 + 2 × 0.2) = 60.2 in.


Pa2AE

3 Do/t( )------------------

2 0.000075( ) 29.0 106×( )

3 301( )--------------------------------------------------------- 4.817 psi < 15.0 psi MAWP. Increase thickness.= = =

Pa2 0.00014( ) 29.0 10

6×( )3 194( )

------------------------------------------------------ 14.0 psi < 15.0 psi MAWP. Increase thickness.= =

Pa2 0.00017( ) 29.0 10

6×( )3 177( )

------------------------------------------------------ 18.6 psi > 15.0 psi MAWP. o.k.= =



436 Openings and Nozzle Reinforcement

Vessel OpeningsOpenings are required to attach piping, mechanical equipment and instrumentation, and to permit inspections. The physical boundary between the jurisdiction of the ASME Code and the appropriate piping code is one of the following interfaces:

• Welded pipe—between the first circumferential joint of pipe and the nozzle• Screwed connections—first threaded joint• Flanged connections—first flange face• Other connections—first sealing surface

a. Calculate A:

b. Enter Figure 5-UCS-28.2 with A = 0.0004 and read B = 5,800

c. Determine Pa:

Of interest is the fact that for 100 psi internal pressure the minimum required thickness of the cylinder is 0.401, while for 15.0 psi external pressure the minimum required thickness is 0.636 in. For the head, the minimum required thickness is only 0.200 in. for internal pressure, while for external pressure the minimum required thickness is less than 0.200 inches.

If a thickness between 0.400 in. and 0.636 in. is desired for the cylinder, stiffening rings are required on the cylinder to obtain a smaller value of L to use in the calculation of Pa. By “trial-and-error,” the approximate maximum stiffening ring spacing with the minimum thickness required for internal pressure of 0.400 is 120 in. as follows:

Assume t = 0.400 in. and L = 120 in.

L/Do = 120/120.8 = 0.993

Do/t = 302

a. Enter Figure 5-UCS – 28.0 and A = 0.00025

b. Enter Figure 5-UCS – 28.2 and B = 3500

c. Determine Pa

This indicates that the optimum design would be one where the shell was thickened above 0.400 in. with stiff-ening rings being placed at a spacing larger than 120 in. center-to-center. The optimum design would be obtained by “trial-and-error.” After the “best” thickness and stiffening ring spacing is determined, the design of the stiffening ring is developed according to UG-29.


A 0.125Ro/t( )

--------------- 0.12560.2/0.2( )

---------------------- 0.0004= = =

PaB

Ro/t( )--------------- 5 800,

60.2/0.2( )---------------------- 19.3 psi > 15 psi MAWP. o.k.= = =

Pa4B

3 Do/t( )------------------ 4 3 500,( )

3 302( )--------------------- 15.45 psi > 15.0 psi MAWP. o.k.= = =



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Opening Shapes and SizesThe Code permits a pressure vessel to be designed with openings of any shape or size. Circular and elliptical openings are recommended because vessel manufacturers are familiar with these. An “obround” opening is formed by two parallel sides and semicircular ends. This shape provides good access for maintenance personnebecause the headroom of this opening can be made large. Elliptical openings ausually found when attached piping joins the shell or head at an oblique angle. When a hillside nozzle is required, the opening in the shell or the head is ellipticAlthough the Code does not restrict the size of a vessel opening, large openingrequire reinforcement, which may cause cost and fabrication problems. Good dpractice dictates that when an opening becomes more than one-half the inside eter of the shell, the design should employ a shell-reducing section instead of anozzle.

Inspection OpeningsInspection openings are required by the Code for all pressure vessels containinprocess environments that cause corrosion, erosion, or mechanical abrasion. Mpressure vessels will require access manways and inspection openings which alarger and more numerous than required by the Code. Vessels which contain innals usually require maintenance of the internals during the life of the unit. Theinspection openings in the vessel should be designed to permit reasonable entpersonnel, welding equipment, and internals components. The Code requires a15-inch manway in vessels having an inside diameter over 36 inches. The folloguidelines may be followed in most cases:

1. For vessels 12-inch nominal diameter and smaller, means of inspection areoften omitted.

2. Vessels from 12-inch up to 18-inch nominal diameter should have two insption openings of 2½-inch minimum diameter. (The Code requires two 1½-insize openings.)

3. Vessels from 18-inch through 36-inch nominal diameter should have two flanged 4-inch openings or a manhole. (The Code only requires two 2-inchopenings.)

4. If replacement of internals is necessary, one end of the vessel should havediameter flange for vessel sizes through 24-inch diameter. From 24-inch through 36-inch vessel diameter, one of the following alternatives can be considered:

– A conical section and a 20-inch diameter flanged end (generally the leaexpensive and most satisfactory alternative)

– Full diameter flanged end– A shell or head manway

5. Vessels larger than 36 inches in diameter:

All should have at least one manway. Recommended minimum diameter is18 inches, although the Code allows a 15-inch diameter.



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In relatively large vessels, two manways are frequently provided to facilitate maintenance and to improve ventilation.

6. The following practices assume that access through internal tray manways is not restricted by appurtenances or by staggering of the tray manways:

Columns having 10 trays or less should be equipped with a manway below the bottom tray and above the top tray.

Columns having more than 10 trays but less than 30 trays should generally be equipped with a manway below the bottom tray, above the top tray, and near the middle of the column, preferably near the feed tray.

Columns of more than 30 trays present access problems which must be given special review. Manways should be provided below the bottom tray and above the top tray. Intermediate manways should, in general, be placed no further than 30 feet (about 15 trays) apart and located as close as economical to points of high corrosion, such as feed trays. If possible, project or plant design specifica-tions should establish manway locations for each column.

You should note the relation between limitations on maximum ladder height without a break (30 feet), and the normal limit for number of trays between manholes, i.e., 15, and seek the economic balance between requirements for safety and requirements for maintenance. For example, in columns having between 10 and 15 trays, it may be possible to arrange the ladders and plat-forms so that a platform is not needed except at top and bottom. If this is the case, the cost of an intermediate platform to serve a “mid-point” or “feed tramanway should be added to the cost of the manway when evaluating econjustification for the “third” manway. If two columns are served by common ladders and platforms, it may be possible to secure a more desirable locatimanways and comply with the ladder length limitation if the number of traysbetween manways exceeds 15. The normal 15-tray limitation should not beexceeded without thorough considerations, but when better over-all designwould result, it may be increased to a maximum of 20 trays if ladders can barranged to meet the safety standard. Thus, in the final design of a columnproblem of manway location is one of balancing the 30-foot ladder limit withdesired manway locations at points of high corrosion.

7. Normally, manways should be at least 18-inch nominal diameter. This diamis sufficient to accommodate routine inspection and equipment for minor mtenance. However, if internal materials, equipment, and appurtenances, surefractory lining, etc., require extensive maintenance or internal staging, a 24-inch manway should be provided to facilitate access of personnel and equipment.

Nozzle Neck ThicknessThe wall thickness of a nozzle neck or other connection should not be less thanthickness computed for the applicable loadings plus the thickness added for cosion allowance on the connection. Except for access openings and openings foinspection only, it should not be less than the smaller of the following:



1. The minimum thickness of standard wall pipe plus the corrosion allowance on the connection.

2. The required thickness (assuming E = 1)1 of the shell (or head) to which the connection is attached plus the corrosion allowance provided in the shell (or head) adjacent to the connection; but for a welded vessel, in no case, less than 1.16 inches.

The minimum thickness of a pipe is the nominal wall thickness less the 12.5% allowable tolerance.

Concept of Reinforcing OpeningsWhen an opening is made in a pressure vessel shell or head, the ability of the nearby wall to retain pressure is significantly reduced. Reinforcing pressure vessel open-ings maintains the pressure retaining capabilities of the shell by the addition of wall thickness near the opening.

The basic rule of the Code is that the wall section around the opening of the vessel must be reinforced with an area of metal equal to the area of metal removed to create the vessel opening. The replaced area of material is called the opening or nozzle reinforcement. The reinforcement may be incorporated into the vessel wall, the nozzle wall, or an attached pad surrounding the nozzle.

The simple rule, however, needs further amplification, as follows:

1. The Code says it is not necessary to replace the removed amount of metal, but the amount of wall thickness required to resist the internal pressure. This required thickness at the openings is usually less than at other points of the shell or head, because of corrosion allowance and nominal size plates yielding extra thickness. In computing the allowable pressure for an existing vessel, most engineers follow the practice of using the lesser wall thickness at the openings, but when designing new vessels, openings can often be reinforced to full as-built shell thickness so vessels may be used up to the limits of their strength whenever required. In the design of new vessels, there will be no appreciable extra cost if the openings are reinforced for the full thickness of the new plate.

2. The plate actually used and nozzle neck are usually thicker than would be required according to calculation. According to the Code, the excess plate in the vessel wall (A1) and nozzle wall (A2) serves as reinforcements. (See Figure 400-7.) Likewise, the inside extension of the opening (A3) and the area of the weld metal (A4) can also be taken as credit for reinforcement. The recom-mended practice is to assume that A is equal to zero; i.e., no excess in vessel wall is credited to reinforcement. Instead, the excess should be assigned to corrosion allowance.

3. The reinforcement must be within certain dimensional limits.

1. E = 0.80 when the opening in a vessel is not radiographed.



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ing

4. The area of reinforcement must be proportionally increased if its stress value is lower than that of the vessel wall.

5. The area requirement for reinforcement must be satisfied for all planes through the center of opening and normal to the vessel surface.

Figure 400-7 presents the equations and terminology for determining nozzle reinforcement.

Calculations for Reinforcement of OpeningsTo determine whether an opening is adequately reinforced, it is first necessary to determine whether the areas of reinforcement available will be sufficient without the use of a pad. The total cross-sectional area of reinforcement required (in square inches) is indicated by the letter A, which is equal to the diameter (plus corrosion allowance) times the required thickness. The area of reinforcement available without a pad includes:

• The area of excess thickness in the shell or head, A1 (Usual practice is to set A1 = 0 and allocate it to corrosion allowance.)

• The area of excess thickness in the nozzle wall, A2

• The area available in the nozzle projecting inward, A3

• The cross-sectional area of welds, A4

If A 1 + A2 + A3 + A4 ≥ A, the opening is adequately reinforced. If A1 + A2 + A3 + A4 < A, a pad is needed.

If the reinforcement is found to be inadequate, then the area of pad needed (A5) may be calculated as follows:

A5 = A − (A1 + A2 + A3 + A4)(Eq. 400-25)

See Figure 400-7 for nozzle and shell data, equations, and nomenclature for calating the area of pad needed for reinforcement.

If a pad is used, the factor (2.5t) in the equation for A2 in Figure 400-7 is measured from the top surface of the pad and, therefore, becomes (2.5tn + Tp). The area A2 must be recalculated on this basis and the smaller value again used. Then:

If A 1 + A2 + A3 + A4 + A5 ≥ A, the opening is adequately reinforced.

The other symbols in Figure 400-7 for the area equations (all values except E1, and F are in inches) are as follows:

d = Diameter in the plane under consideration of the finished openin its corroded condition

t = Nominal thickness of shell or head, less corrosion allowance



Fig. 400-7 Nozzle Reinforcement Calculations

P = Pressure, psi

S = Allowable stress, psi

Rn = Inside radius of nozzle, in.

E1 = Joint efficiency, %

tn = Actual thickness of nozzle (minus corrosion)

trn = Calculated thickness of nozzle

tn–trn = Excess thickness

t = Actual thickness of shell or head (minus corrosion)

tr = Calculated thickness of shell or head

E1t–Ftr = Excess thickness in shell or head

trnPRn

SE1 0.6P–-------------------------=

Nozzle Data(1) Shell Data

Area of reinforcement required(1)

Area of excess thickness in shell or head (use greater value) (Chevron practice is to set A1 = 0)

Area available in nozzle projecting outward, use smaller value

Area available in nozzle projecting inward

Cross-sectional area of welds

Area of reinforcement available without pad

If A1 + A2 + A3 + A4 ≥ A

If A1 + A2 + A3 + A4 < A

A = dtrF

A1 = (E1t–Ftr)d, or A1 = 2(E1t–Ftr)(t + trn)

A2 = 2(2.5tn)(tn–trn)fr (2), or A2 = 2(2.5t)(tn–trn)fr

A3 = 2(tn–c)fr × h

(A1 + A2 + A3 + A4)

Opening is adequately reinforced.

Opening is not adequately reinforced.Element must be added or thickness increased.

A4 2W1( )2

W2( )2+

2------------------------------------- fr×=

Area is pad A5 = 2WpTp =

Total area available =



tr = Required thickness of shell or head as defined in Code Paragraph UG-37

trn = Required thickness of a seamless nozzle wall

Tp = Thickness of reinforcement pad

Wp = Width of reinforcement pad

tn = Nominal thickness of nozzle wall, less corrosion allowance

W1 = Cross-sectional area of weld

W2 = Cross-sectional area of weld

E1 = 1 when an opening is in the plate or when the opening passes through a circumferential point in a shell or cone (exclusive of head-to-shell joints), or

Joint efficiency obtained from Code Table UW-12 when any part of the opening passes through any other welded joint

F = Correction factor which compensates for the variation in pressure stresses on different planes with respect to the axis of a vessel. A value of 1.00 is used for F in all cases except when the opening is integrally reinforced. If integrally reinforced, see Code Figure UG-37.1

fr = Strength reduction factor. Ratio of material stresses: ≤ 1.0

h = Distance nozzle projects beyond the inner surface of the vessel wall before corrosion allowance is added.

To correct for corrosion, deduct the specified allowance from shell thickness, t, and nozzle thickness, tn, but add twice its value to the diameter of opening, d.

Note Total load to be carried by attachment welds may be calculated from the equation, W = S(A-A1), in which S equals the allowable stress specified by Code Section C and Code Paragraph UW-15(b). For large openings in cylindrical shells, see Code Appendices 1-7. For openings in flat heads, see Code Paragraph UG-39. For weld sizes required, see Code Paragraph UW-16 and Code Figure UW-16.1.

On all welded vessels built under column C of Code Table UW-12, 80% of the allowable stress value must be used in design equations and calculations.

Notes: 1. For cases when the allowable stress of the nozzle of reinforcing element is less than the allowable stress of the vessel, refer to Code Par. UG-41 (A) and Appendix L, latest Addenda, for consideration of this effect.

2. If reinforcing pad is used, the factor 2.5tn becomes (2.5tn + Tp)

Fig. 400-7 Nozzle Reinforcement Calculations

1. Reinforcement is considered integral when it is inherent in the shell plate or nozzle. Reinforcement built up by welding is also considered integral. Installation of a reinforcement pad is not considered integral.



See Figures 400-8, 400-9, and 400-10 for examples of reinforcement calculations.

Design for External Pressure on OpeningsThe reinforcement required for openings in single-walled vessels subject to external pressure need be only 50% of that required for internal pressure where tr is the wall thickness required by the rules for vessels under external pressure [Code UG-37 (c)(1)].

Fig. 400-8 Example 1: Reinforcement Calculations (1 of 2)

Problem

Determine the reinforcement requirements of an 8-in. I.D. nozzle which is centrally located in a 2:1 ellipsoidal head. The inside diameter of the head skirt is 41.75 in. The head material is SA 516 Gr 70 and the nozzle SA 106 Gr C. The design pressure is 700 psig. The design temperature is 500°F. There is no corrosion and the weld joint efficiency is 1.0.

Solution

1. Since the allowable tensile stress for both SA 516 Gr 70 and SA 106 Gr C at 500°F is 17.5 ksi, the material strength reduction factor is fr = 1.0.

2. The minimum required thickness of a 2:1 ellipsoidal head without an opening is determined from UG-32(d) as:

3. According to Rule (3) for tr in UG-37(a), when an opening and its reinforcement are in an ellipsoidal head and are located entirely within a circle the center of which coin-cides with the head and the diameter is equal to 80% of the shell diameter, tr is the thickness required for a seamless sphere of radius K1D, where D is the shell I.D. and K1 is 0.9 from Table UG-37. For this head, the opening and reinforcement shall be within a circle with a diameter of 0.8d = (0.8)(41.75) = 33.4 in.

4. Following (3), above, the radius R = K1D = 0.9(41.75) = 37.575 in. is used in UG-27(d) to determine the tr for reinforcement as:

5. Using UG-27(c)(1), determine the required nozzle thickness:

t PD2SE 0.2P–-------------------------- 700( ) 41.75( )

2 17 500 1.0×,( ) 0.2 700( )–--------------------------------------------------------------- 0.838in. Actual thickness used is 1.0 in.= = =

trPR

2SE 0.2P–-------------------------- 700( ) 37.575( )

2 17 500 1.0×,( ) 0.2 700( )–--------------------------------------------------------------- 0.755 in.= = =



Reinforcement of Openings for External PressureThe cross-sectional area (A) of reinforcement required for openings in vessels subject to external pressure:

(Eq. 400-26)

where:d = Diameter in the given plane of the opening in its corroded condi-

tion, inches

tr = Wall thickness required for external pressure, inches

6. Limit parallel to head surface = X = d or (d/2 + t + tn), whichever is larger. X = 8 in. or (4 + 1 +1.125 = 6.125), use X = 8 in.

7. Limit perpendicular to shell surface = Y = 2½t or 2½tn, whichever is smaller. Y = 2½(1) = 2.50 in. or 2½(1.125) = 2.81 in., use 2.50 in.

8. Limits of 2X = 2(8) = 16 in. is less than 33.4 in.; therefore, provision to use spherical head rule is valid.

9. Reinforcement area required according to UG-37(c) is:

Ar = dtrF + 2tntrF(1 – fr1) = (8)(0.755)(1) + 0 = 6.040 in.2 when fr1 = 1.0

10. Reinforcement available in head is:

A1 = d(Et – Ftr) – 2tn(Et – Ftr)(1 – fr1)

When fr1 = 1.0, the second term becomes zero; therefore, for E = 1.0 and F = 1.0

A1 = d(t – tr) = (8)(1.0 – 0.755) = 1.960 in2

11. Reinforcement available in nozzle is:

A2 = 2Y(tn – trn) = (2)(2.5)(1.125 – 0.164) = 4.805 in2

12. Total reinforcement available in head and nozzle is:

At = A1 + A2 = 1.960 + 4.805 = 6.765 in2

Area available of 6.765 in2 is greater than area required of 6.040 in2.

13. Determination of weld strength and load paths.

According to UW-15(b), strength calculations for welds for pressure loading are not required for nozzles like that shown in Figure UW-16.1(c). Since this nozzle is similar to that detail, no load path calculations are required.


trn

PRnSE 0.6P–----------------------- 700( ) 4( )

17 500 1.0×,( ) 0.6 700( )–------------------------------------------------------------ 0.164 in.= = =

Ad tr F××

2----------------------=



ey

F = Factor for computation of the required reinforcement area on different planes (as the pressure-stress varies) when the opening is in cylindrical shell or cone and integrally reinforced. For all other configurations the value of F = 1.

437 Bolted Flanged ConnectionsThe Code covers the design of flanges in (mandatory) Appendix 2. The scope of the rules apply to gaskets contained entirely within the bolt circle. According to the rules, acceptable flanged nozzles may be attained by the use of either standard rated flanges or by flange calculations. Normally, the standard ANSI B16.5 “Pipe Flanges” or API 605 “Large Diameter Carbon Steel Flanges” will be used, as th


Design Data:

Inside Diameter of Shell, D = 48 in.Design Pressure, P = 300 psi at 200°FShell material, t = 0.500 in. SA 516-70 plate, S = 17,500 psiThe vessel spot examined, E = 0.85There is no allowance for corrosion.Nozzle nominal size 6 in. Rn = 2.88Nozzle material, tn = 0.432 in. wall, SA 53B,Seamless pipe S = 15,000 psiExtension of nozzle inside the vessel, 1.5 in.h = 2.5tn = 2.5 × 0.432 = 1.08 in.The nozzle does not pass through the main seams.Fillet weld size 0.500 in.

Wall thickness required:

= 0.416 in.

= 0.058 in.

Area of reinforcement required:

A = dtr = 5.761 × 0.416 = 2.397 in2

Area of reinforcement available:

A1(1) = (excess in shell) Larger of following:

(t – tr)d = (0.500 – 0.416) 5.76 = 0.484 in 2

or (t – tr)(tn + t)2 = (0.500 – 0.416) (0.432 + 0.500) 2 = 0.156

A2 = Excess in nozzle neck) Smaller of following:

Shell trPR

SE 0.6P–----------------------- 300 24×

17 500 1 0.6– 300××,------------------------------------------------= =

Nozzle trn

PRnSE 0.6P–----------------------- 300 2.88×

15 000, 1 0.6 300×–×-----------------------------------------------------= =



(tn – trn) 5t = (0.432 – 0.058) 5 × 0.500 = 0.935

or (tn – trn) 5tn = (0.432 – 0.058) 5 × 0.432 = 0.808

Since the strength of the nozzle is lower than that of the shell, a decreased area shall be taken into consideration: 15,000/17,500 = 0.857, 0.857 × 0.808 = 0.692 in2

A3 = (inside projection) tn × 2h = 0.432 × 2 × 1.08 = 0.933 area decreased 0.933 × 0.857 = 0.800 in2

A4 = (area of fillet weld) 2 × 0.5 × 0.5002 × 0.857 = 0.214 in2

A5 = (area of fillet weld inside) 2 × 0.5 × 0.5002 × 0.857 = 0.214 in2

Total area available: = 2.404 in2

Since this area is greater than the area required for reinforcement, additional reinforcement is not needed.

(1) The recommended Company practice would assign a value of 0 to A1. This example illustrates minimum Code requirements.


Problem

Determine the reinforcement requirements for an 8-inch SA 516 Gr 70 seamless schedule 80 radial SA-53B seamless pipe nozzle projecting 1.25 inches into the vessel I.D., in a 0.500-inch thick cylindrical pressure vessel shell of inside radius of 24 inches with a design pressure of 300 psig at 200°F. The vessel is fully radiographed and E = 1.0 for all joint efficiencies. There is no corrosion and the nozzle does not intersect any main seams. It has 0.375 inch fillet welds.

1. The required solution wall thicknesses are:

= 0.416 in.

= 0.077 in.

2. Area of reinforcement required:

A = d × tr = 7.625 × 0.416 = 3.172 in2

3. Limits of reinforcement:

Parallel to shell

X = d or rn + tn + t, the larger

= 7.625 or 3.8125 + 0.5 + 0.5

Use X = 7.625

Perpendicular to shell


Shell trPR

SE 0.6P–----------------------- 300 24×

17 500 1 0.6 300×–×,-------------------------------------------------= =

Nozzle trnPRn

SE 0.6P–----------------------- 300 3.8125×

15 000 1 0.6 300×–×,-------------------------------------------------= =



nge

and

can be selected with very little design effort, “off the shelf.” When a standard flais selected from these specifications, no additional calculations are required.

The following are typical flange standards:

• MSS SP-44, Classes 300, 400, 600, and 900 in sizes 26-36 inches

• Standard 605, “Large Diameter Carbon Steel Flanges,” 75, 150, and 300-pound rating in sizes 26-60 inches inclusive

• Taylor Forge Standard, Classes 75, 175, and 350 in sizes 26-72 inches, 9296 inches

Y = 2½ t or 2½ tn, the smaller

= 1.25 or 1.25

Use Y = 1.25

4. Area of reinforcement available:

A1(1)= (excess in shell) Larger of following:

(t – tr)d = (0.500-0.416) 7.625 = 0.641 in2

or (t – tr)(tn + t)2 = (0.500 – 0.416) (0.500 + 0.500) 2 = 0.168

A2 = (excess in nozzle neck) Smaller of following:

(tn – trn) 5t = (0.500 – 0.077) 5 × 0.5 = 1.058

or (tn – trn) 5tn = (0.500 – 0.077) 5 × 0.5 = 1.058

Since the strength of the nozzle is lower than that of the shell, a reduced area shall be taken into consideration.

15,000/17,500 = 0.857, 0.857 × 1.058 = 0.907 in2

A3 = (inside projection) tn × 2h = 0.5 × 2 × 1.25 = 1.25Area decreased 0.857 × 1.25 = 1.071 in2

A4 = (area of fillet weld) 2 × 0.5 × 0.3752 = 0.141 in2

(The area of pad-to-shell weld disregarded)

Total area available = 2.76 in2

This area is less than the required area; therefore, the difference shall be provided by a reinforcing element. It may be heavier nozzle neck, larger extension of the nozzle inside of the vessel, or reinforcing pad. Using a reinforcing pad, the required area of pad: 3.172-2.760 = 0.412 in2 using 0.250 in SA 516-70 plate for reinforcing pad the width of the pad

The outside diameter of reinforcing pad:

Outside diameter of pipe: 8.625

Width of reinforcing pad: 1.64810.273 in.

(1) The recommended practice would assign a value of 0 to A1. This procedure illustrates minimum Code requirements.


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es

• AWWA (American Water Works Association) Standard C207-55, Classes Band E, in sizes 6-96 inches

Follow the rules of Appendix 2 for the following conditions: (1) when it is neces-sary to calculate a flange because one of the standard flanges in the correct siznot available; (2) when the pressure temperature ratings are not adequate; or (3) when special design considerations are to be addressed. The design of flanlargely a trial and error process where the flange thickness is varied. Stresses iflange and hub are calculated, and if they exceed the allowable tensile stress athickness is chosen until the stresses are within the allowed amount. The codecovers integral, loose, and optional flange designs. Integral means the pipe, huring are one continuous piece. An example of this type of flange is a welding neflange. Loose means no attachment of the assembly to the pipe, such as slip-ojoint, and threaded flanges. Optional flanges are integral flanges by constructiobut analysis is permitted by the simpler method for loose flanges.

The calculation of flanges requires:

• Selection of materials for flange, bolts, and gaskets

• Determination of facing and gasket details so that the bolt loading may be determined and bolt sizes selected

• Determination of the bolt circle and the loads, moment arms and moments to gasket seating and operating conditions

• Determination of stresses

Manufacturers like Taylor Forge provide manuals devoted to the design of flangand include a one page summary giving step by step calculation sheets.

An example calculation for a weld neck flange is shown in Figure 400-11.

Fig. 400-11 Sample Calculations for a Bolted Flange (1 of 4)

Problem

What is the minimum required flange thickness of a weld neck flange with the following flange and gasket details and design conditions?

Appendix 2 rules are to be followed to design the flange.

Design pressure, p = 2,500 psi

Design temperature = 250°F

Bolt-up and gasket seating temperature = 70°F

Flange material SA 105

Bolting material SA 325 Grade 1

No corrosion.

Allowable bolt stress at design = Sb = 20,200 psi

Allowable flange stress at bolt-up and design = Sf = 17,500 psi

Gasket Details: Spiral-wound metal, fiber filled, stainless steel, inside diameter = 13.75 in. and width, N, is 1.0 in.



Gasket dimensions: bo = N/2 = 0.5 in. and b = 0.5 = 0.3535 in. (Appendix 2, Table 2-5.2) G = 13.75 + (2 × 1) – (2 × 0.3535) = 15.043 in.

Solution

1. Determine bolt loadings and size of bolts:

N = 1; b = 0.3535; y = 10,000; m = 3.0 (from Appendix 2, Table 2-5.1)

H = πG2p/4 = π (15.043)2(2500)/4 = 444,320 (hydrostatic force)

Hp = 2b π Gmp = 2(0.3535) π (15.043)(3.0)(2500) = 250,590 (joint contact surface compression load)

Wm1 = H + Hp = (444,320) + (250,590) = 694,910 (operating bolt load)

Wm2 = πbGy = π(0.3535)(15.043)(10,000) = 167,060 (gasket seating bolt load)

Am = the greater of Wm1/Sb = (694,910)/(20,200) = 34.4 in.2 or Wm2/Sa = 167,060)/(19,200) = 8.3 in.2 (total required bolting cross-sectional area)

Ab = actual bolt area = 36.8 in2 for 16 bolts at 2 in. diameter (choose bolts in multiples of 4)

W = 0.5(Am + Ab)Sa = 0.5(34.4 + 36.8)(20,200) = 719,120 lb (gasket seating)

W = WM1 = 694,910 lb (operating condition)

2. Calculate the total flange moment for the design condition:


bo



Flange Loads

HD = (π/4) B2p = (π/4)(10.75)2(2500) = 226,910 lb (hydrostatic end force based on I.D. of flange)

HG = Hp = 250,590 lb

HT = H-HD = (444,320) – (226,910) = 217,410 lb

Lever ArmshD = R + 0.5g1 – g1 + 0.5g1 – 0.5 × 3.375 = 4.1875 inch

hG = 0.5(C – G) = 0.5(22.5 – 15.043) = 3.7285 inch

hT = 0.5(R + g1 + hG) = 0.5(2.5 + 3.375 + 3.7285) = 4.8018 in. (see Appendix 2, Table 2-6.)

Flange MomentsMD = HD × hD = (226,910)(4.1875) = 950,190 in-lb

MG = HG × hG = (250,590)(3.7285) = 934,320 in-lb

MT = HT × hT = (217,410)(4.8018) = 1,043,960 in-lb

M = MD + MG + MT = 2,928,470 in-lb

3. Calculate the total flange moment for bolt-up condition:

Flange Load at Bolt Up

HG = W = 719,120 lb

Lever ArmhG = 0.5(C-G) = 3.7285 in.

Flange MomentMbolt up = HG × hG = (719,120)(3.7285) = 2,681,240 in-lb

4. Mo = greater of Mdesign or Mbolt up(Sh/Sc) = 2,928,470 in-lb.

5. Shape constants for flange:

K = A/B = (26.5)/(10.75) = 2.465

From Appendix 2, Figure 2-7.1: T = 1.35; Z = 1.39; Y = 2.29; U = 2.51

g1/go = (3.375)/(1.0) = 3.375

h/ho = (6.25)/(3.279) = 1.906

From App. 2, Figure 2-7.2: F = 0.57

From App. 2, Figure 2-7.3: V = 0.04

From App. 2, Figure 2-7.6: f = 1.0

e = F/ho = (0.57)/(3.279) = 0.1738

d = (U/V)hogo2 = (2.51/0.04)(3.279)(1)2 = 205.76

6. Calculation of stresses:


C B–2

-------------= 22.5 10.75–2

---------------------------=

ho Bgo 10.75( ) 1.0( ) 3.279= = =



Blind FlangesBlind flanges are covered in Code Paragraph UG-34. The computations determine the thickness required for both operating conditions and gasket seating (Equation UW-34). The more severe case is chosen. The basic equations for these two condi-tions are:

1. Operating Conditions

(Eq. 400-27)

2. Gasket Seating Conditions

Assume t = 4.5 in.

L = (te + 1)/ T + t3/d = (1.3201) + (0.4429) = 1.7630

Longitudinal hub stress,

SH = fMo/Lg12B = (1)(2,928,470)/(1.7630)(3.375)2(10.75)

SH = 13,570 psi

Radial flange stress,

SR = (4/3te + 1) Mo/Lt2B = (2.0428)(2,928,470)/(1.763)(4.5)2(10.75)

SR = 15,590 psi

Tangential flange stress,

ST = (YMo/t2B) – ZSR = (2.29)(2,928,470)/(4.5)2(10.75) – 1.39(15,590)

ST = 9,140 psi

7. Allowable stresses:

SH ≤ 1.5 Sf: 13,570 < (1.5)(17,500) = 26,250 psi; o.k.

SR ≤ Sf: 15,590 < 17,500 psi; o.k.

ST ≤ Sf: 9,140 < 17,500 psi; o.k.

0.5 (SH + SR) ≤ Sf: 14,580 < 17,500 psi; o.k.

0.5 (SH + ST) ≤ Sf: 11,350 < 17,500 psi; o.k.

Since all calculated stresses are below the allowable stresses, the selection of t = 4.5 in. is adequate. If an optimum minimum thickness of flange is desired, calculations must be repeated with a lesser value of t until one of the calculated stresses or stress combinations is approximately equal to the allowable stress even though other calculated stresses are less than their allowable stress.


t d1.3PSE

-----------1.9W hG

SEd3

----------------------+ 0.5

=



(Eq. 400-28)

438 Minimum Wall Thickness and Nominal Plate Sizes

Minimum Thickness—The ASME Code permits various minimum thicknesses of plate. See Figure 400-12. All vessels, regardless of alloy content, should have a minimum thickness of 3/16 inch, in addition to a 1/16 inch minimum corrosion allowance. In effect, this sets the minimum wall thickness required for pressure at 1/8 inch.

t G1.9Wa hG

SaEG3

------------------------ 0.5

=

Fig. 400-12 Minimum Thickness of Plate

Material Minimum Thickness Code Reference

Carbon and low-alloy steel 3/32 in. for shells and heads used for compressed air, steam, and water service

Par. UG-16Par. UCS-25

Minimum thickness of shells and heads after forming shall be 1/16 in. plus corrosion allowance

Par. UG-16

¼ in. for unfired steam boilers

Heat-treated steel ¼ in. for heat-treated steel Par. UHT-16

Clad vessels Same as for carbon and low-alloy steel based on total thickness for clad construction and the base plate thickness for applied-lining construction

Par. UCL-20

High-alloy steel 3/32 in. for corrosive service1/16 in. for noncorrosive service

Par. UHA-20

Nonferrous materials 1/16 in. for the welded construction in noncorrosive service

3/32 in. for welded construction in corrosive service

Par. UNF-16

Single-welded butt joint without use of backing strip may be used only for circumferential joints not over 24 in. outside diameter and mate-rial not over 5/8 in. thick

Table UW-12

Low-temperature vessels: 5, 8, 9 percent nickel steel

3/16 in. Par. ULT-16



f r

late

of ich

i-, the e kup to at

e r

.

joint

ese

-15

Utilization of Available Metal and Commercial Thickness Plate—After the minimum thickness has been determined, including corrosion allowance, vessels are often brought to the next highest commercial plate thickness. This raises the ques-tion of what to do with the thickness gained by specifying standard commercial plate. The usual practice is to call this extra plate thickness corrosion allowance, for two reasons:

• Increasing the allowable pressure of a vessel might necessitate redesign oseveral other pieces of equipment in its system in order to prevent the othepieces from limiting the system.

• The ability to predict corrosion rates is not as precise as the ability to calcupressure and structural requirements.

439 Design of Welded JointsSee Section 600 for additional information on welding of pressure vessels.

Overall ConsiderationsThe choice of weld design depends on a number of factors including:

• The conditions of welding. Accessibility has a definite influence on the typeweld joint. In small diameter vessels, the openings may have weld joints whare not accessible on both sides.

• The ASME code itself limits the design and use of weld joints based upon service, material, and location of weld.

• Company practices. Certain types of welds have been found through experence to be more reliable and less prone to failure than others. For exampleuse of backup rings in circumferential welds leads to many problems, yet thCode allows this. The Company has found the following problems with bacrings: (1) they are difficult to check for weld quality by radiography; (2) theyinvariably contain root weld defects; (3) they greatly increase susceptibility some form of fracture or cracking; and (4) they are not worth the savings thmight be achieved.

• Economic considerations. If the previous three choices allow freedom to usdifferent welds, then economic consideration may be the deciding factor; foexample, V-edge preparation of a weld joint which can be made by flame cutting. This is usually more economical than J- or U-type edge preparation

Terminology and Determination of Joint EfficiencyThe Code bases the joint efficiencies required in wall thickness calculations on category and joint type.

Joint category is the location of the weld on the vessel. Figure 400-13 shows thfour main categories.

Joint type defines the configurations of a welded joint and not its location on thevessel. Figure 400-14 also shows an example of a Type 2 weld joint. Figure 400



shows the various Code weld types. The Code places requirements on the permis-sible weld joints to be used. Figure 400-21 shows these limitations.

Note Figure 400-21 is a foldout appearing at the end of this section.

Fig. 400-13 Categories of Welded Joints: See Figure 400-21 for special requirements based on service material, thickness, and other design conditions.

Fig. 400-14 Terminology of Joint Category and Joint Efficiency

Joint category defines the location of a joint in a vessel and does not define the type of joint.

Categories of Joints

1. Longitudinal welds in main shell or attached nozzles, welds in a sphere, head or flat-sided vessel, circumferential joints connecting hemi-heads to a vessel or part of a vessel.

2. Circumferential welds in a shell or nozzle or connecting heads other than hemi-heads or part of a vessel.

3. Welds connecting flanges, tube sheets or flat heads to a vessel or part of a vessel and welds connecting flat-sided vessels.

4. Welds connecting nozzles to a vessel or part of a vessel.

Note A butt weld is defined as a weld connecting two pieces of material with surfaces that are 30 degrees or less from being in the same plane.

Types of Joints

Type defines the configurations of a welded joint. It does not define the location of the joint in the vessel. For example, Type 2 joint is a single-butt weld with a backing strip left in place.



d gned.

Once the weld joint category and joint type are established then the joint efficiency can be determined remembering that joint efficiency is also a function of the degree of radiography used to inspect the welds.

440 Structural DesignThis section discusses how to determine the design criteria for structural loads, including those from wind and earthquake. The engineer using this information should have a working knowledge of structural mechanics. If this is not the case, then a Company specialist or third party design contractor should be consulted.

This section discusses wind and earthquake design specific to pressure vessels, but refers to the Civil and Structural Manual for determining the design criteria. PVM-EF-66 may be used as a convenient form for many of the structural design calculations, and is a useful checklist and record for future reference.

441 OverviewASME Code, Section VIII, Division 1, Paragraph UG-22, or Division 2, Paragraph AD-110, specifically requires vessel designs to include, in addition to internal and external pressure, the following (classified in this section as structural loadings):

• Vessel weight

• Weight from normal contents

• Weight from test conditions (hydrotest including static head)

• Weight from attachments (platforms, piping, machinery, linings, insulation, etc.), plus any mechanical loads associated with those attachments

• Attachment of internals

• Vessel supports, such as lugs, rings, skirts, saddles, and legs

• Environmental loadings from wind, earthquake, and vibration

This section explains how these structural loads combine with other stresses anshows how to calculate the overall stress for a particular component being desi

Joint Efficiency

The joint efficiency (E) or % of stress value (S) to be used in calculating required thick-ness or maximum

• Type of welded joint

• Degree of radiography performed

Fig. 400-14 Terminology of Joint Category and Joint Efficiency



trays ntains

442 Vertical Vessels: Combining Structural Loads

In-Service ConditionsCircumferential stresses from internal pressure usually control the design of cylin-drical pressure vessels. However, in tall vertical vessels, five other factors may contribute to longitudinal stresses, and must be considered. These are the following:

Environmental loads (never considered to act simultaneously):

• Wind load• Seismic loads• Vibration

Other loads:• Dead weight• Operating pressure (or vacuum)

The dead weight of the vessel must include all appurtenances such as internal and catalyst, as well as external attachments such as platforms. Appendix F coweight curves to help estimate vessel weights.

Fig. 400-15 Types of Welded Joints From Pressure Vessel Handbook by E. F. Megyesy. Courtesy of Pressure Vessel Handbook Publishing, Inc.

TypesCode UW-12

Joint Efficiency, E When the Joint is:

A.Fully

Radiographed

B.Spot

Examined

C.Not

Examined

1.Butt joints as attained by double-welding or by other means which will obtain the same quality of depos-ited weld metal on the inside and outside weld surface.

Backing strip if used shall be removed after comple-tion of weld. Not recom-mended since they are crack starters.

1.00 0.85 0.70



The operating pressure used in these calculations should be the Maximum Allow-able Working Pressure (MAWP) as defined in Section 420. External pressure (vacuum) rarely controls the design, but it must be reviewed in each case. (Vacuum design is discussed in Section 430.)

The first step to calculate structural loads is to determine which environmental load (wind, seismic or vibration) contributes the greatest stress, and thereby controls the design. Section 445 below provides details on calculating environmental loads.

The greatest environmental stress is then combined with stresses from internal/ external pressures. Figure 400-16 illustrates how to combine stresses for the in-service conditions. Note that signs (+ or -) of calculated stress values must be closely monitored. For an example demonstrating how to combine in-service stress conditions, see Figure 400-17.

Determining the shell thickness is an iterative process. The design procedure is as follows:

1. Calculate shell stresses due to internal pressure.

2. Determine minimum shell thickness required for internal pressure.

2.

Single-welded butt joint with backing strip which remains in place after welding

0.90 0.80 0.65

3.Single-welded butt joint without use of backing strip — — 0.60

4.

Double-full fillet lap joint — — 0.55

5.

Single-full fillet lap joint with plug welds — — 0.50

6.Single full fillet lap joint without plug welds — — 0.45

Fig. 400-15 Types of Welded Joints From Pressure Vessel Handbook by E. F. Megyesy. Courtesy of Pressure Vessel Handbook Publishing, Inc.



.)

3. Calculate longitudinal shell stresses due to wind or earthquake, whichever controls, using the minimum shell thickness required for internal pressure.

4. Combine longitudinal shell stresses, including stress from dead weight.

If the combined stresses are greater than the allowable stresses, increase the shell thickness until the combined, calculated shell stresses are less than or equal to those allowable.

The shell stresses should be calculated at the following locations:

• At the point of attachment to the support structure (i.e., skirt, ring girder, etc• At the joint of the skirt to the head• At the bottom head-to-shell joint

Fig. 400-16 Vertical Vessel In-Service Stress Combinations

P = MAWP (psi)

T = Minimum wall thickness at location stress calculation (in) [i.e., wall thickness less corrosion allowance]

DO = Outside diameter (in)

DI = Inside diameter (in)

D = Mean diameter (in)

M = Maximum moment (wind or seismic plus any eccentric moment) at location of stress calculation (lb-in) [i.e., @ Height “Y”]

W = Weight of vessel and components above point of stress calculation (lbs)

+ = Tension

- = Compression

Do DI+

2------------------



Fig. 400-17 Example Calculation (1 of 2)

Problem During operation, what is the longitudinal stress at points A and B for a design environmental load?

Solution • A review of the environmental loads for earthquake and wind indicates EQ controls at the eleva-tion being considered.

• Mean Diameter, D, is

• Calculate weight (W) of vessel and components above point of stress. W = (2.3 K/ft) (55 ft) (1,000 lb/K) = 126,500 lb

P = 0.150 ksi = 150 psi

t =

D = 54.5 in. [Do = 55 in., Di = 54 in.]

M = 17,500 K · in. (1,000 lb/K) = 17.5 × 106 lb · in.

W = 126,500 lb

Do Di+

2------------------ 55 in. 54 in.+

2------------------------------- 54.5 in.= =

58-- 1

8--– 1

2-- in.=



he g lly o-. For

• At all changes of diameter or thickness of the vessel

Pre-Service ConditionsPre-service conditions cover fabrication, transportation, erection, and testing. Vertical vessels are typically fabricated in the horizontal position and lifted into tvertical position. Maximum bending stresses must be calculated using erectionweights, which may not include internal trays or catalyst. Furthermore, operatinpressures are typically zero during this phase of the work. Test conditions usuaconsist of a hydrostatic test, and all stress conditions must be considered. Hydrstatic testing should also be made the responsibility of the designer by contractadditional guidance, contact an engineering specialist.

At point A, internal pressure and bending cause a tension stress, and the dead load results in compression:

Pressure + Bending – Dead Load

TotalLongitudinal

= Stress

σA =+ –

=+ –

σA = 4,087.5 psi + 15,139.6 psi – 1,477.6 psi = 17,749 psi @ A (tension)

At Point B, internal pressure produces tension stress, and bending plus dead load results in compression:

Pressure – Bending – Dead Load

TotalLongitudinal

= Stress

σB =– –

=– –

σB = 4,087.5 psi – 15,139.6 psi – 1,477.6 psi = 12,530 psi @ B (compression)

Fig. 400-17 Example Calculation (2 of 2)

PD4t------- 32DoM

π Do4

Di4

–

---------------------------------4W

π Do2

Di2

–

---------------------------------

150 psi( ) 54.5 in.( )4 0.5 in.( )

------------------------------------------- 32 55 in.( ) 17.5 106

lb in.⋅×( )

π 554

544

–( )-------------------------------------------------------------------

4 126 500 lb,( )

π 552

542

–( )--------------------------------

PD4t------- 32DoM

π Do4

Di4

–

---------------------------------4W

π Do2

Di2

–

---------------------------------

150 psi( ) 54.5 in.( )4 0.5 in.( )

------------------------------------------- 32 55 in.( ) 17.5 106

lb in.⋅×( )

π 554

544

–( )-------------------------------------------------------------------

4 126 500 lb,( )

π 552

542

–( )--------------------------------



gle

tion.

e ssel

tem, sels

an

ns,

ally ting vice es

l for

443 Horizontal Vessels: Structural Design

In-Service ConditionsCircumferential stresses from internal pressure usually control the design of cylin-drical pressure vessels. In horizontal vessels, wind or seismic loads rarely affect the shell stress calculation. A horizontal vessel on a saddle support acts as a beam with the following special circumstances:

• The loading conditions are different for a full or partially filled vessel.

• The circumferential stresses in the vessel vary according to the included anof the saddle.

• The load due to the weight of the vessel is combined with other loads.

The following loads for horizontal vessels need to be considered:

• Support reaction at each saddle. Design the vessel for a full waterload reac

• Internal pressure. Since the longitudinal stress in the vessel shell is only onhalf of the circumferential stress, approximately one half of the actual platethickness in the longitudinal direction is available to resist the load of the veweight and contents.

• Wind and seismic loads. These are important for designing the support systypically consisting of a saddle, anchor bolts, and a concrete pier. Long veswith very small thickness/radius values are subject to distortion from wind pressure. However, a vessel designed to 1 psi or more external pressure csuccessfully resist external loads encountered in normal service.

Maximum longitudinal stresses are found by combining:

• Operating weight• Operating internal pressure (or vacuum)

Maximum circumferential stress must be checked at the saddle support locatioand stiffeners may be required.

Pre-Service ConditionsThe orientation of a horizontal vessel during fabrication and in service are typicthe same, so no major change in shell stresses should occur. However, if the lifpoints of the vessel are not located near the final support points, then a pre-sershear and bending moment diagram should be drawn and the maximum stressconsidered. Attaching lifting lugs near the support saddle or near the shell-headintersection is recommended in order to avoid additional stiffening of the vessepre-service loads.



psi.

her

tress

da

)

444 Allowable Stresses and Deflections

Nonpressure Parts of VesselAll structural steel not part of the actual pressure vessel should be designed in accordance with AISC Manual of Steel Construction. A one-third increase over basic allowable Code stresses is permitted for stresses resulting from wind or earth-quake loads.

SkirtsFor vessel skirts, and other cylindrical shells not subject to internal or external pres-sure, the longitudinal tensile stress under loading conditions, including wind or earthquake, should not exceed 85% of the yield stress at the operating temperature at the section under consideration. The efficiency of full-penetration, butt-weld joints should be taken as 1.0.

The compressive stress for carbon and low-allow steel shells operating at atmo-spheric temperature should not exceed the lesser of the following stresses:

• 2/3 of the minimum specified yield stress• 2.32 × 106 t/R (psi)

where:t = shell thickness, in.

R = mean Radius, in.

The expression 2.32 × 106 t/R is based on modifying the classical buckling stress equation for a cylindrical shell given as CE t/r, with C=0.08 and E = 29,000,000

For carbon and low-alloy steel shells operating at high temperatures, and for otmaterials, the allowable compressive stress should not exceed the lesser of thefollowing:

• For cylindrical shells designed before publication of the Summer 1983 Addenda of the ASME Code: 1/3 greater than the allowable compressive spermitted by the applicable edition of the ASME Code.

• For cylindrical shells designed in accordance with the Summer 1983 Addenor subsequent editions: the allowable compressive stress permitted by the ASME Code.

DeflectionsLateral deflections of columns resulting from wind loads are typically limited by practical considerations of fluid sloshing off the trays in a vertical tower. The following are typical limits, per industry practice, on calculated deflections (deltadue to wind.

• Column with trays

• Column with packing

delta < 6 inches(100 feet height)----------------------------------------

delta < 9 inches(100 feet height)----------------------------------------



in

e

s that

e

Actual displacements during earthquakes are hard to predict given the transient nature of earthquake ground motions. For this reason, strict limits are not set.

445 Shell Stresses from Wind and Earthquake

Wind LoadsWind loads produce two different effects on vessels. The primary effect is an equiv-alent static force from the maximum expected wind velocities, which causes lateral shears and overturning moments. The other effect is wind-induced vibration caused by vortex shedding in a steady state wind.

Section 100 of the Civil and Structural Manual outlines procedures for determining the static wind loads, and critical wind velocities at which vessel vibrations may occur.

The following information is required to consider the effect of wind on the design of a vessel:

• Design wind speed for the facility• Exposure category• Vessel diameter, height or length, wall thicknesses• Mean steady state wind velocity at the site• Attachments to the vessel that affect the shape coefficient• Operating weights of vessel (both pre-service and in-service)

The design wind speed, exposure category, and shape coefficients are all givenSection 100 of the Civil and Structural Manual. Using this information and following the procedure and examples will help the vessel designer calculate thforces on the vessel and supports due to wind.

Earthquake LoadsVessels in high seismic zones are subjected to lateral forces during earthquakecause large stresses in the vessel wall. Section 100 of the Civil and Structural Manual also presents methods and examples for calculating the magnitude of seismic loads and their effect in vessel design.

The following information is required to consider the effect of earthquakes on thdesign of the vessel:

• Seismic zone at the facility• Height or length and wall thicknesses of the wall• Weight distribution along length of vessel• Type of support system• Site coefficient, which is a function of soil type• Operating weights of vessel (both pre-service and in-service)• Weight of attachments and their location.



e

ith

o the n-

f 40.

The seismic zone and description of site coefficients for typical Company facilities are given in Section 100 of the Civil and Structural Manual. Using this information and following the procedure and examples, the forces of earthquake loads on the vessel and its supports can be calculated. The critical moment is the larger of the two moments for wind or seismic loadings.

Moment AmplificationCertain “slender” vessels are susceptible to increased stresses as a result of theccentricity of the vessel weight in the deflected position.

Slender vessels are defined as vessels with one of the following conditions:

1. Uniform diameter, and uniform mass distribution, with a height-to-diameter(H/D) ratio exceeding 30.

2. Non-uniform vessels with large loads concentrated in the upper portions wH/D ratio exceeding 15.

Allowance should be made for the increased stresses under lateral loads due tpotential weight eccentricity in the deflected position. The designer may compesate for this concern by adding to the base wind moment (Mw) or earthquake moment (MEQ) a value Mo as defined in Equation 400-29.

(Eq. 400-29)

(Replace W2 with W1 for computing anchor bolt tension with wind loads.)

The variation of this moment with height is

(Eq. 400-30)

where:W1 = Dead weight, kips

W2 = Total operating weight, kips

X = Height above grade, ft

H = Vessel height, ft

D = Vessel diameter, ft

MDesign = (MW or MEQ) + Mo

It is recommended that free standing uniform vessels be limited to a H/D ratio oVessels with larger H/D ratios should be guyed.

Mo

W2H2

6000D----------------=

Mx Mo 1 3XH----

2– 2

XH----

3+=



Horizontal VesselsIn horizontal vessels, wind and earthquake loads primarily affect the design of saddles and supports. The design of the shell is normally controlled by other loads such as design pressures, operating weight, saddle reactions, etc. A horizontal cylin-drical vessel supported by two saddles acts like a uniformly loaded, simply supported beam.

Vessels with hemispherical or elliptical dished heads may be treated as an equiva-lent cylinder having a beam length equal to L + 4H/3, where L is the tangent-to-tangent length of the vessel and H is the depth of the curvature of the heads. Figure 400-18 shows a horizontal vessel with two saddle supports.

Given the simple beam analysis of a horizontal vessel supported by saddles, the following stress conditions may be determined:

1. Longitudinal bending stress

2. Tangential shear stress

3. Circumferential stress

Depending on the location of the saddles with respect to overall length, the maximum longitudinal stress will occur at midspan or at each support. Generate a

Fig. 400-18 Horizontal Vessel with Saddle Supports



l” ce-

at are the . ort f o the from

r nd n of

ical. ss to hell e d to

he ulting

an-

shear and bending moment diagram for the vessel and check the maximum longitu-dinal stress, at the position of maximum bending moment.

Tangential shear and circumferential shear stress will be maximum at each saddle support. The stresses in the vessel vary according to the angle included by the saddle. Circumferential bending is caused by the tendency of a cylinder to “ovawhen supported on line supports, like a saddle. Section 447 below outlines produres for saddle and stiffener design.

446 Internal LoadsFor both vertical and horizontal vessels, internal appurtenances create loads thdirectly applied to the vessel shell. These design concerns are best handled byvessel designer, with input from engineers familiar with the process technology(See below for examples.) For vertical vessels, trays are typically used to suppcatalyst or bubble caps. These are typically horizontal projections in a stream ofluid or gas that is running vertically. These loads are usually transferred back tvessel shell. Typical loads may come from catalyst deadweight, pressure drop gas flow, and change in momentum of fluids.

Horizontal vessels may contain vertical plates for support of exchanger tubes, obaffle plates in knockout drums. These items are usually attached to the shell, athe local effects of bending moments or shears must be considered in the desigthe shell.

447 Structural SupportsProcess vessels are normally supported by one of the following methods (See Figure 400-19):

• Skirts• Support legs• Support lugs• Ring girders• Saddles

Skirts are typically used for vertical vessels because they are the most economLeg-supported vessels are normally lightweight, and the legs provide easy accethe bottom of the vessel. A lug support system depends on the stiffness of the sand its ability to adequately resist the bending moments. This capacity should bcompletely investigated. Cross-bracing on lug-supported vessels may be needeminimize lateral and torsional movements.

Vessels supported by ring girders are usually placed within a structural frame. Tring girder has the advantage of supporting torsional and bending moments resfrom the transfer of loads from the vessel wall to the supports.

Horizontal vessels are normally supported by saddles. Stiffening rings may be required if the shell is too thin to transfer the loads to the saddles. Thermal exp



sion is typically accommodated by having one end of the vessel on a sliding support.

Vertical Vessel SkirtsThe design procedure outlined here is based on cylindrical skirts. (Tapered designs are a special problem and beyond the scope of this manual.) Design of the skirt consists of determining the operating weight of the vessel and the controlling bending moment due to wind or earthquake. The skirt thickness should be checked at three locations:

1. Top of the skirt where the allowable stress may be reduced by the temperature of the vessel, and by the efficiency, E, of the weld between the skirt and the shell. (See Standard Drawing GD-C78876, Standard Skirt and Base Details for Vertical Vessels, located in the Standard Drawings and Forms Section.)

E = 0.5 for Type A Attachment per Standard Drawing GD-C78876

E = 1.0 for Type B Attachment per Standard Drawing GD-C78876

Type B attachments are preferred for all cases: In dished heads, there is a high stress concentration at the knuckle which would be augmented by an attach-ment such as a Type B skirt. Brittle failures starting at the highly stressed inside face of the knuckle have occurred in large vessels with thick heads. For this

Fig. 400-19 Typical Vessel Supports



ch

es ment

reason, the maximum thickness for dished heads should be 5/8 inch after forming. Since the thickness of dished heads is limited, the use of the Type B skirt attachment is recommended.

2. Sections with the greatest number and/or size of skirt openings. (manways, pipe penetrations, etc.)

3. Base of skirt where the maximum load exists.

The stress in the skirt is represented by:

(Eq. 400-31)

where:σ = Longitudinal stress in skirt

W = Weight of vessel

M = Moment due to wind or earthquake forces (plus moment amplifi-cation as appropriate).

A = 2πRt

I = πR3t

R = Radius of skirt

C = R/2

t = Thickness of skirt

Use any consistent units for this stress calculation.

In most practical applications, the ratio of R/t is greater than 10. Hence, the area A and the moment of inertia I are expressed in the simplified version shown above.

The equation for the stress in a skirt then becomes:

(Eq. 400-32)

Because the compressive stress in the skirt is larger than the tensile stress, the compressive stress usually controls the design and is kept below the allowable stresses defined in Section 444 above. The thickness of the skirt should be ¼ inminimum.

The base of skirts should normally be checked on the basis of allowable stresscorresponding to ambient temperature. Special attention should be given to theadded effect of thermal stresses in those vessels employing Type A skirt attach

σ W–A

--------M C⋅

I-------------±=

σ W–2πRt-------------

M

πR2t

------------±=



s

r of s

sm

s n-

to The ch

where the vessel “design temperature” is 450°F or above, and the skirt height i3 feet or less.

If a Type B attachment is used, special attention should be given to the thermalstresses for temperatures above 650°F and skirt heights under 3 feet.

Openings in the SkirtFor skirt compression, deduct the width of openings, G (ft), around the perimetethe skirt. The following equations take into account the shift in the neutral axis awell as the reduction in cross section. The equation for maximum compressive stress f under combined axial and bending loads at openings can be written:

(Eq. 400-33)

where:D = Diameter of skirt, ft

α and ß are coefficients depending on the ratio G/D. The expression for deter-mining α and ß is complex, but both can be approximated with slight conservatifor ranges up to G/D = 0.7 (90° opening) by the following equation:

(Eq. 400-34)

Substituting these values for a and ß in Equation 400-34 gives:

(Eq. 400-35)

This equation is for a single opening and is conservative when G is the sum of several openings distributed around the circumference. When the skirt thicknesobtained using this equation is large and when there are several distributed opeings or one large opening with G > 0.7D, more exact methods of determining FB may be justified.

Vertical Vessel Support LegsSupport legs can be made of wide-flange or pipe members. They are designedresist axial loads, overturning moments, and lateral shear forces in the vessel. attachment point to the vessel may cause high concentrated shell stresses whi

fWαπDT------------

4Mβ

πD2T

--------------+=

α β 1

13GπD-------–

-----------------= =

f1

πDT------------

πDπD 3G–---------------------

W4MD

--------+ =

FB f( ) T( ) 1πD 3G–--------------------- W

4MD

--------+ = =



must be investigated by a qualified vessel designer. The allowable stress in the member may also be reduced because of high temperatures and the efficiency of the weld between the leg and the shell. The design of this joint is beyond the scope of this manual, and should be handled by an engineer familiar with this subject.

Vertical Vessel Lug SupportsThe main design consideration regarding lug-supported vessels is the stress magni-tude in the shell. Bijlaard’s method, as covered in WRC 107, is usually followed for the design. This method consists of determining the stress in the shell at the vicinity of a support lug of height 2C2 and width 2C1, as shown in Figure 400-20. These designs should also be left to qualified specialists.

Fig. 400-20 Vertical Vessel Lug Supports



rts

to

ialists.

The bending moment in the shell due to support eccentricity is given by:

Mo = F⋅e(Eq. 400-36)

and the maximum stress in the shell is calculated from WRC 107. Both membrane and bending stresses are considered. Details of the calculations are well established in WRC 107.

Horizontal Ring GirdersRing girders are typically used in elevated vessels when they are supported by a structural frame. The stress distribution in a ring girder is very complicated. Given a uniform load around the perimeter, the stresses and forces can be calculated given the following assumptions:

• Supports equally spaced

• Vertical deflection at supports is zero

• Slope of ring girder at supports is zero due to symmetry of loads and suppo

• Torsion force at supports is zero. This inherently assumes the girder is abletwist as a result of the flexibility of the shell.

Even given these simplifying assumptions, the exact design of a ring girder is beyond the scope of this manual, and the designer is referred to qualified spec

Horizontal Vessels—Saddle Supports and Stiffening RingsHorizontal vessels are typically supported on two saddles. The use of two saddles is preferred both statically and economically over a system with more supports.

The vessel designer should follow these steps:

1. Locate the position of saddle supports.

2. Calculate the maximum support reactions and corresponding allowable shell stresses.

3. Provide circumferential stiffening rings if required.

The location of the saddle is sometimes governed by the position of openings, sumps, etc., on the vessel. When no openings dictate saddle location, the statically optimal points should be chosen, i.e., where the positive and negative bending moments are nearly equal. Thin walled vessels with a large diameter are best supported near the heads, to utilize the stiffening effect of the heads. The distance between the head tangent line and the saddle should in no case be more than 20% of the tangent-to-tangent length, L. Note: one end of the horizontal vessel typically contains a sliding support to facilitate thermal expansion. See Standard Drawing GA-C99694 “Standard Details of Support Feet for Horizontal Vessels.”



-rs is ll

the ld be l

n or

stan-

The

ing

or

he

Once the locations of the supports are determined, the calculation of the maximum support reaction is a matter of simple statics. The following load conditions must be considered, along with the corresponding allowable stress.

• Erection/fabrication• Hydrotest• Operation

Use the maximum support reaction to determine whether a vessel requires stiffening at the supports. A criterion for making that decision and for sizing stiffenepresented in Appendix B, “Design of Supports and Stiffening Rings for Thin WaHorizontal Vessels.”

To use Appendix B, the designer needs the following information:

• Support reactions• Vessel wall thickness, excluding corrosion allowance• Shell material• Operating temperature

Appendix B also gives an alternative method for stiffening the shell at points of vessel support by using internal struts. However, the use of internal struts shoulimited to large vessels operating at low internal pressures, since internal strutsrestrain the radial growth of the shell. This restraint may produce circumferentiadistortion causing failure of internal weld connections.

Base Ring and Anchor BoltsVertical vessels, stacks, and towers must be fastened to the concrete foundatiostructural frame by means of anchor bolts and the base (bearing) ring.

Base Ring. Standard Drawing GD-C78876 “Standard Skirt and Base Details for Vertical Vessels” should be used to detail the base ring dimensions. To use thisdard drawing, the designer must know:

• The maximum tension per bolt• The maximum compression per foot of circumference

See the following discussion on anchor bolts for determining the above values.drawing then provides all the required fabrication information, such as base ringthickness, stiffener sizes, and weld requirements.

Anchor Bolts. In selecting the number of anchor bolts and base type, the followshould be used as a guide:

• Use bolts in increments of four with a minimum of eight

• Bolt spacing as defined on Standard Drawing GB-Q68922. “Standard AnchBolts,” (Civil and Structural Manual) should be strictly adhered to unless appropriate reductions in allowable capacities are made, as presented on tstandard drawing.



r bolt into g. No .

pres-

• Normally use Type II anchor bolts (Standard Drawing GB-Q68922) to facili-tate erection of the vessel.

Following is the development of the equations for base compression and anchotension. The factor of 0.75 applied to the equations (and on PVM-EF-66) takesaccount the 1/3 increase in allowable stress under earthquake and wind loadinfurther reduction of the load or increase in the allowable stress should be made

Note that in computing bolt tension for wind forces, the dead weight (W1) is used instead of the operating weight (W2), as the maximum bolt tension under wind loading occurs when the vessel is empty.

The equations for base compression, anchor bolt tension, and skirt or shell comsion are based on the general equation for homogeneous elastic beams:

(Eq. 400-37)

where (in any consistent units):F = Force

L = Length

f = Maximum stress at extreme fiber (F)

W = Load at section (F)

M* = Moment at section (F-L)

A = Cross-sectional area (L2)

S = Section modulus (L3)

* = Controlling moment Mw or MEQ

For anchor bolts:

(Eq. 400-38)

fWA-----

MS-----±=

A NAB=

SNABDB

4--------------------=

fBW

NAB------------

4MNABDB--------------------–=

TB fBAB–1N----

4MDB-------- W1–

= =



where:N = Number of bolts

AB = Root area of one bolt (L2)

DB = Bolt circle diameter (L)

fB = Maximum stress in bolt (F/L2)

TB = Maximum tension per bolt (F)

W1 = Dead weight

W2 = Total operating weight

For base compression:

(Eq. 400-39)

where:tB = Width of base (L)

PB = Maximum pressure under base (F/L2)

B = Maximum linear compression (F/L)

In making the final selection of anchor bolt size, choose the next commercially available size above the calculated requirement. Oversizing anchor bolts is not a good practice. Studies show that the ductility provided by anchor bolts yielding at extreme loadings (whether wind or seismic) gives additional structural damping and energy absorption which increases the structural performance.

The above equations for anchor bolt tension and base compression are not exact. For bolt tension, the values obtained using the equations are conservative. More exact methods have been published, but most of these are tedious and the assump-tions made are sometimes questionable. Because of the simplicity of the equations presented here and the empirical nature of wind and earthquake design, we recom-mend the use of these conservative equations.

A πDBtB=

SπDB

2tB

4-------------------=

f PB

W2

πDBtB-----------------

4M

πDB2tB

-------------------+= =

B PBtB1

πDB-----------

4MDB-------- W2+

= =



sign

el

n

for

ard

III, g the order

lso gned

de

ance, the

450 Calculation Sheets and Computer Programs

451 OverviewThis section describes the available methods for performing the mechanical design of pressure vessels. These design tools are used mainly to determine the shell and head thicknesses, maximum and test pressures, vacuum design, and support design. They help to insure that our equipment meets the ASME Code and Company speci-fications, including Company recommendations that exceed the ASME Code minimum requirements.

There are three methods of design available:

• Manual calculation• Mainframe computer calculation• Personal computer calculation

Forms PVM-EF-65 and PVM-EF-66 are available for assistance with manual decalculations. The mainframe computer programs include ENGR09 for simple designs, ENGR17 for horizontal vessel designs and ENGR63 for complex vessdesigns.

Generalized data sheets are available for transferring the calculated informatioto Drawings: PVM-DS-35C, PVM-EF-332, PVM-EF-338, PVM-EF-339, and PVM-EF-340. These are general pressure vessel design sheets and are usefulmost applications. They are available in printed form from CRTC Support Staff Services, or as computer (CAD) files. Reduced copies are included in the StandDrawings and Forms section of this manual.

452 Calculation Sheet, PVM-EF-65Data Sheet PVM-EF-65, Design of Pressure Vessels to ASME Code, Section VDivision 1, has been developed to aid the pressure vessel designer in performincomponent design and pressure calculations. These calculations are needed into insure that vessels meet ASME Code and Company specifications. The formeliminates the need to look up the equations every time they are needed and aprovides a consistent format so that all Company pressure vessels can be desito the same standards.

Familiarity with the ASME Code, Section VIII, Division 1, would be helpful to thedesigner. PVM-EF-65 provides the needed code equations and refers to the Cowhenever necessary. It is intended that the form be used in conjunction with theCode, and does not replace it.

The first sheet of PVM-EF-65 is a listing of design information that is needed tocomplete the remainder of the form. It is important to fill in all of the requested information in order to obtain accurate results and economical designs. For instin tall vessels, the operating hydrostatic head would have a noticeable effect on



only ss of

d for g main-ed by and me

ure, res-es the ch

ause to po-re

ad of e

re-

e of

gures

thickness of the lower shell and head. If ignored, the upper shell could be oversized or worse, the lower shell and head could be undersized.

The internal pressure designs of the shell and heads are completed by using sheet two of PVM-EF-65. This sheet includes equations for calculating thicknesses using either the inside diameter (I.D.) or the outside diameter (O.D.). The O.D. equations should only be used when the vessel is being fabricated from standard pipe. In this case, the inside diameter varies with wall thickness and the outside diameter is fixed. When using pipe, remember to include a 12½% thinning allowance when choosing a standard pipe wall. This allowance does not affect the O.D. or I.D., the thickness. Sheet two also provides the equations for determining the thicknehemispherical and torispherical heads.

External pressure design is described on sheet three. The methods are provideboth cylindrical shells and ellipsoidal heads. A procedure for designing stiffeninrings is also included. This procedure is similar to the one used by Company's frame vessel design programs and indicates the limits to size of stiffeners as usthe programs. Design and placement of stiffening rings can be time-consumingcomplicated and should only be done manually for simple vessels. The mainfracomputer programs should be used for more complicated designs.

Sheet four provides the equations for calculating the maximum allowable pressnew and cold, the maximum pressure at 90% of the yield stress, the shop test psure, and the field test pressure. The shop test pressure is calculated at 1½ timmaximum allowable working pressure, new and cold, for the full thickness of eacomponent. The highest value of all components is used provided it does not cany component to exceed 90% of yield. If any component were to limit the test 90% of its yield stress, the test pressure would equal that amount and that comnent would require full radiography of its category A welds. The field test pressuis figured at 1½ times the component design pressure minus the hydrostatic hethe test liquid, normally water at the full inside height of the vessel, including thheads.

A legend is provided on sheet five of the form including a definition of each abbviation used throughout the form, arranged in alphabetical order.

After PVM-EF-65 has been completed, the information can be transferred to onthe pressure vessel data sheets that is to be sent to the fabricator.

460 Quick Reference Guide to ASME Code, Section VIII, Division 1Figure 400-22 is a graphic presentation of the various rules, paragraphs, and fiin ASME Code, Division 1. It has been reproduced by courtesy of the Hartford Steam Boiler Inspection and Insurance Co. based on the 1986 Edition, 1986 Addenda.

Note The information in this figure is provided for quick reference and to save time and should not be used for design. Always refer directly to the latest edition of the Code, updated with the latest addenda.




Fig. 400-21 Design of Welded Joints

Pressure Vessel Manual

400 Mechanical and Structural Design

Chevron Corporation400-79

April 2000

Fig. 400-22Quick-Reference Guide to ASM

E Code, Section VIII, Division 1. Courtesy of the Hartford Steam Boiler Inspection and Insurance Co. (1 of 2)



Fig. 400-22 Quick-Reference Guide to ASME Code, Section VIII, Division 1. Courtesy of the Hartford Steam Boiler Inspection and Insurance Co. (2 of 2)