overview_of_pressure_vessel_design__asme_career_series_9667

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I. INTRODUCTION

Pressure vessels are used in many industries (e.g., hydrocarbon processing, chemical, power, pharmaceutical, food and beverage). The mechanical design of most pressure vessels is done in accordance with the requirements contained in the ASME Boiler and Pressure Vessel Code, Section VIII. Section VIII is divided into three divisions. This course provides an overview of pressure vessel mechanical design requirements. It focuses on Division 1, highlights the differences in scope among the three divisions of Section VIII, and discusses several factors related to pressure vessel design that the ASME Code does not cover. The following summarizes the main sections of the course:

• General

- Main Pressure Vessel Components

- Primary Process Functions of Pressure Vessels

- Scope of ASME Code Section VIII

- Structure of Section VIII, Division 1

• Materials of Construction

- Material Selection Factors

- Maximum Allowable Stress

• Design

- Design Conditions and Loadings

- Weld Joint Efficiency and Corrosion Allowance

- Design for Internal Pressure

- Design for External Pressure and Compressive Stresses

- Reinforcement of Openings

- Flange Rating

- Flange Design

- Maximum Allowable Working Pressure

• Other Design Considerations

- Vessel Support

- Local Loads

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- Vessel Internals

• Fabrication

- Acceptable Welding Details

- Postweld Heat Treatment Requirements

• Inspection and Testing

- Inspection

- Pressure Testing

This course is nominally only ½ day in length. Therefore, it cannot provide an in-depth treatment of all aspects of pressure vessel design. However, the topics are covered in sufficient depth to provide participants with a general understanding of pressure vessel design requirements, to design pressure vessel components to a limited extent, or to review pressure vessel designs prepared by others. It also prepares individuals who require a more thorough understanding of pressure vessel design to attend a more in-depth course or to acquire the necessary knowledge on their own.

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II. General

This section describes the various components of pressure vessels through the use of conceptual drawings. It also describes the scope of the ASME Boiler and Pressure Vessel Code Section VIII, and the basic structure of Section VIII, Division 1.

A. Main Pressure Vessel Components

Pressure vessels are containers for fluids that are under pressure. They are used in a wide variety of industries (e.g., petroleum refining, chemical, power, pulp and paper, food, etc.)

1.0 Shell

The shell is the primary component that contains the pressure. Pressure vessel shells are welded together to form a structure that has a common rotational axis. Most pressure vessel shells are either cylindrical, spherical, or conical in shape.

• Figure 2.1 illustrates a typical horizontal drum. Horizontal drums have cylindrical shells and are fabricated in a wide range of diameters and lengths.

• Figure 2.2 illustrates a small vertical drum. Small vertical drums are normally located at grade. The maximum shell length-to-diameter ratio for a small vertical drum is about 5:1.

• Figure 2.3 illustrates a typical tall, vertical tower. Tall vertical towers are constructed in a wide range of shell diameters and heights. Towers can be relatively small in diameter and very tall (e.g., a 4 ft. diameter and 200 ft. tall distillation column), or very large in diameter and moderately tall (e.g., a 30 ft. diameter and 150 ft. tall pipestill tower).

A tower typically contains internal trays in the cylindrical shell section. These internal trays (noted in Figure 2.3) are needed for flow distribution. Several types of tower trays are available, such as the bubble -cap, valve, sieve, and packed. The particular type of tray used depends on the specific design conditions and process application.

The shell sections of a tall tower may be constructed of different materials, thicknesses, and diameters. This is because temperature and phase changes of the process fluid – two of

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the factors that affect the corrosiveness of the process fluid - vary along the tower’s length. Alloy material or a corrosion-resistant lining are sometimes used in sections of a vertical tower where corrosion is a critical factor.

• Figure 2.4 illustrates a typical reactor vessel with a cylindrical shell. The process fluid undergoes a chemical reaction inside a reactor. This reaction is normally facilitated by the presence of catalyst which is held in one or more catalyst beds.

2.0 Head

All pressure vessel shells must be closed at the ends by heads (or another shell section). Heads are typically curved rather than flat. Curved configurations are stronger and allow the heads to be thinner, lighter, and less expensive than flat heads. Figures 2.1 through 2.4 show heads closing the cylindrical sections of the subject pressure vessels. Heads can also be used inside a vessel. These “intermediate heads” separate sections of the pressure vessel to permit different design conditions in each section.

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Nozzle

ShellA

A

Head

Saddle Support(Fixed)

Saddle Support(Sliding)

Head

Section A-A

Horizontal Drum on Saddle Supports Figure 2.1

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Head

Shell Nozzle

Head

SupportLeg

Vertical Drum on Leg Supports Figure 2.2

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Trays

NozzleHead

Shell

Nozzle

Cone

Shell

Nozzle

NozzleSkirtSupport

Head

Tall Vertical Tower Figure 2.3

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InletNozzle

Head

Shell

UpperCatalyst

Bed

Catalyst BedSupport Grid

LowerCatalyst

BedOutletCollector

Head

SupportSkirt

OutletNozzle

Vertical Reactor Figure 2.4

3.0 Nozzle

A nozzle is a cylindrical component that penetrates the shell or heads of a pressure vessel. The nozzle ends are usually flanged to allow for the necessary connections and to permit easy disassembly for maintenance or access. Nozzles are used for the following applications:

• Attach piping for flow into or out of the vessel.

• Attach instrument connections, (e.g., level gauges, thermowells, or pressure gauges).

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• Provide access to the vessel interior at manways.

• Provide for direct attachment of other equipment items, (e.g., a heat exchanger or mixer).

Nozzles are also sometimes extended into the vessel interior for some applications, such as for inlet flow distribution or to permit the entry of thermowells.

Figure 2.5 shows a pressurized storage vessel with a spherical shell.

CrossBracing

SupportLeg

Shell

Spherical Pressurized Storage Vessel Figure 2.5

4.0 Support

The type of support that is used depends primarily on the size and orientation of the pressure vessel. In all cases, the pressure vessel support must be adequate for the applied weight, wind, and earthquake loads. The design pressure of the vessel is not a consideration in the design of the support since the support is not

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pressurized. Temperature may be a consideration in support design from the standpoint of material selection and provision for differential thermal expansion.

4.1 Saddle Supports

Horizontal drums (See Figure 2.1) are typically supported at two locations by saddle supports. A saddle support spreads the weight load over a large area of the shell to prevent an excessive local stress in the shell at the support points. The width of the saddle, among other design details, is determined by the specific size and design conditions of the pressure vessel. One saddle support is normally fixed or anchored to its foundation. The other support is normally free to permit unrestrained longitudinal thermal expansion of the drum.

4.2 Leg Supports

Small vertical drums (See Figure 2.2) are typically supported on legs that are welded to the lower portion of the shell. The maximum ratio of support leg length to drum diameter is typically 2:1. Reinforcing pads and/or rings are first welded to the shell to provide additional local reinforcement and load distribution in cases where the local shell stresses may be excessive. The number of legs needed depends on the drum size and the loads to be carried. Support legs are also typically used for spherical pressurized storage vessels (See Figure 2.5). The support legs for small vertical drums and spherical pressurized storage vessels may be made from structural steel columns or pipe sections, whichever provides a more efficient design. Cross bracing between the legs, as shown in Figure 2.5, is typically used to help absorb wind or earthquake loads.

4.3 Lug Supports

Lugs that are welded to the pressure vessel shell (See Figure 2.6) may also be used to support vertical pressure vessels. The use of lugs is typically limited to vessels of small to medium diameter (1 to 10 ft.) and moderate height-to-diameter ratios in the range of 2:1 to 5:1. Lug supports are often used for vessels of this size that are located above grade within structural steel. The lugs are typically bolted to horizontal structural members to provide stability against overturning loads; however, the bolt holes are often slotted to permit free radial thermal expansion of the drum.

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4.4 Skirt Supports

Tall, vertical, cylindrical pressure vessels (e.g., the tower and reactor shown in Figures 2.3 and 2.4 respectively) are typically supported by skirts. A support skirt is a cylindrical shell section that is welded either to the lower portion of the vessel shell or to the bottom head (for cylindrical vessels). Skirts for spherical vessels are welded to the vessel near the mid-plane of the shell. It is normally not necessary for the skirt bolt holes to be slotted (as with lug supports). The skirt is normally long enough to provide enough flexibility so that radial thermal expansion of the shell does not cause high thermal stresses at its junction with the skirt.

Vertical Vessel on Lug Supports

Figure 2.6

B. Scope of the ASME Code Section VIII

Pressure vessels are typically designed in accordance with the ASME Code Section VIII, even for locations outside the US. Section VIII is divided into three divisions: Division 1, Division 2, and Division 3. Division 1 is used most often since it contains sufficient requirements for the majority of pressure vessel applications.

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The main objective of ASME Code rules is to establish the minimum requirements that are necessary for safe construction and operation. The ASME Code protects the public by defining the material, design, fabrication, inspection, and testing requirements that are needed to achieve a safe design. Experience has shown that the probability of a catastrophic pressure vessel failure is reduced to an acceptable level by use of the ASME Code.

The ASME Code is written to apply to many industries. Accordingly, it cannot anticipate and address every possible design requirement or service application. Therefo re, users must supplement the ASME Code by specifying additional requirements that are appropriate for their particular industry and applications.

1.0 Division 1

The ASME Code Section VIII, Division 1 applies for pressures that exceed 15 psig and through 3,000 psig. At pressures below 15 psig, the ASME Code is not applicable. At pressures above 3,000 psig, additional design rules are required to cover the design and construction requirements that are needed at such high pressures.

The ASME Code is not applicable for piping system components that are attached to pressure vessels. Therefore, at pressure vessel nozzles, ASME Code rules apply only through the first junction that connects to the pipe. This junction may be at the following locations:

• Welded end connection through the first circumferential joint.

• First threaded joint for screwed connections.

• Face of the first flange for bolted, flanged connections.

• First sealing surface for proprietary connections or fittings.

The Code also does not apply to non pressure-containing parts that are welded, or not welded, to pressure-containing parts. However, the weld that makes the attachment to the pressure part must meet Code rules. Therefore, items such as pressure vessel internal components or external supports do not need to follow Code rules, except for any attachment weld to the vessel.

The ASME Code identifies several other specific items where it does not apply. These include:

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• Fired process tubular heaters (e.g., furnaces).

• Pressure containers that are integral parts mechanical devices (e.g., pump, turbine, or compressor casings).

• Piping systems and their components.

Note that all detailed design requirements discussed in this course are based on Division 1. Refer to Divisions 2 and 3 for comparable information in those documents

2.0 Division 2, Alternative Rules

The scope of Division 2 is identical to that of Division 1; however, Division 2 contains requirements that differ from those that are contained in Division 1. Several areas where the requirements between the two divisions differ are highlighted below.

• Stress. The maximum allowable primary membrane stress for a Division 2 pressure vessel is higher than that of a Division 1 pressure vessel. The Division 2 vessel is thinner and uses less material. A Division 2 vessel compensates for the higher allowable primary membrane stress by being a more stringent than Division 1 in other respects.

• Stress Calculations. Division 2 uses a complex method of formulas, charts, and design by analysis that results in more precise stress calculations than are required in Division 1.

• Design. Some design details are not permitted in Division 2 that are allowed in Division 1.

• Quality Control. Material quality control is more stringent in Division 2 than in Division 1.

• Fabrication and Inspection. Division 2 has more stringent requirements than Division 1.

The choice between using Division 1 and Division 2 is based on economics. The areas where Division 2 is more conservative than Division 1 add to the cost of a vessel. The lower costs that are associated with the use of less material (because of the higher allowable membrane stress) must exceed the increased costs that are associated with the more conservative Division 2 requirements in order for the Division 2 design to be economically attractive.

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A Division 2 design is more likely to be attractive for vessels that require greater wall thickness, typically over approximately 2 in. thick. The thickness break point is lower for more expensive alloy material than for plain carbon steel, and will also be influenced by current market conditions. A Division 2 design will also be attractive for very large pressure vessels where a slight reduction in required thickness will greatly reduce shipping weights and foundation load design requirements.

3.0 Division 3, Alternative Rules For Construction of High Pressure Vessels

Division 3 applies to the design, fabrication, inspection, testing, and certification of unfired or fired pressure vessels operating at internal or external pressures generally above 10,000 psi. This pressure may be obtained from an external source, a process reaction, by the application of heat, or any combination thereof. Division 3 does not establish maximum pressure limits for either Divisions 1 or 2, nor minimum pressure limits for Division 3.

C. Structure of Section VIII, Division 1

The ASME Code, Section VIII, Division 1, is divided into three subsections as follows:

• Subsection A consists of Part UG, the general requirements that apply to all pressure vessels, regardless of fabrication method or material.

• Subsection B covers requirements that apply to various fabrication methods. Subsection B consists of Parts UW, UF, and UB that deal with welded, forged, and brazed fabrication methods, respectively.

• Subsection C covers requirements that apply to several classes of materials. Subsection C consists of Parts UCS (carbon and low-alloy steel), UNF (nonferrous metals), UHA (high-alloy steel), UCI (cast iron), UCL (clad and lined material), UCD (cast ductile iron), UHT (ferritic steel with properties enhanced by heat treatment), ULW (layered construction), and ULT (low-temperature materials).

Division 1 also contains the following appendices:

• Mandatory Appendices address subjects that are not covered elsewhere in the Code. The requirements that are contained in these appendices are mandatory when the subject that is covered is included

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in the pressure vessel under consideration. Examples of Mandatory Appendices are:

- Supplementary Design Formulas

- Rules for Bolted Flange Connections with Ring Type Gaskets

- Vessels of Noncircular Cross Section

- Design Rules for Clamped Connections

• Nonmandatory Appendices provide information and suggested good practices. The use of these nonmandatory appendices is not required unless their use is specified in the vessel purchase order. Examples of nonmandatory appendices are:

- Basis for Establishing Allowable Loads for Tube-to-Tubesheet Joints

- Suggested Good Practice Regarding Internal Structures

- Rules for the Design of Tubesheets

- Flanged and Flued or Flanged Only Expansion Joints

- Half-Pipe Jackets

III. Materials of Construction

This section discusses the primary factors that influence material selection for pressure vessels and the maximum allowable material stresses specified by the ASME Code. The mechanical design of a pressure vessel can proceed only after the materials have been specified. The ASME Code does not state what materials must be used in each application. It specifies what materials may be used for ASME Code vessels, plus rules and limitations on their use. But, it is up to the end user to specify the appropriate materials for each application considering various material selection factors in conjunction with ASME Code requirements.

A. Material Selection Factors

The main factors that influence material selection are:

• Strength

• Corrosion Resistance

• Resistance to Hydrogen Attack

• Fracture Toughness

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• Fabricability

Other factors that influence material selection are cost, availability, and ease of maintenance.

1.0 Strength

Strength is a material's ability to withstand an imposed force or stress. Strength is a significant factor in the material selection for a particular application. Strength determines how thick a component must be to withstand the imposed loads.

The overall strength of a material is determined by its yield strength, ultimate tensile strength, creep and rupture strengths. These strength properties depend on the chemical composition of the material. Creep resistance (a measure of material strength at elevated temperature) is increased by the addition of alloying elements such as chromium, molybdenum, and/or nickel to carbon steel. Therefore, alloy materials are often used in elevated temperature applications.

2.0 Corrosion Resistance

Corrosion is the deterioration of metals by chemical action. A material's resistance to corrosion is probably the most important factor that influences its selection for a specific application. The most common method that is used to address corrosion in pressure vessels is to specify a corrosion allowance. A corrosion allowance is supplemental metal thickness that is added to the minimum thickness that is required to resist the applied loads. This added thickness compensates for thinning (i.e., corrosion) that will take place during service.

The corrosion resistance of carbon steel could be increased through the addition of alloying elements such as chromium, molybdenum, or nickel. Alloy materials, rather than carbon steel, are often used in applications where increased corrosion resistance is required in order to minimize the necessary corrosion allowance.

3.0 Resistance to Hydrogen Attack

At temperatures from approximately 300°F to 400°F, monatomic hydrogen diffuses into voids that are normally present in steel. In these voids, the monatomic hydrogen forms molecular hydrogen, which cannot diffuse out of the steel. If this hydrogen diffusion

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continues, pressure can build to high levels within the steel, and the steel can crack.

At elevated temperatures, over approximately 600°F, monatomic hydrogen not only causes cracks to form but also attacks the steel. Hydrogen attack differs from corrosion in that damage occurs throughout the thickness of the component, rather than just at its surface, and occurs without any metal loss. In addition, once hydrogen attack has occurred, the metal cannot be repaired and must be replaced. Thus, it is not practical to provide a corrosion allowance to allow for hydrogen attack. Instead, materials are selected such that they are resistant to hydrogen attack at the specified design conditions.

Hydrogen attack is a potential design factor at hydrogen partial pressures above approximately 100 psia. Material selection for these hydrogen service applications is based on API 941, Steels for Hydrogen Service at Elevated Temperatures and Pressures in Petroleum Refineries and Petrochemical Plants. API 941 contains a family of design curves (the Nelson Curves) that are used to select appropriate material based on hydrogen partial pressure and design temperature.

4.0 Fracture Toughness

Fracture toughness refers to the ability of a material to withstand conditions that could cause a brittle fracture. The fracture toughness of a material can be determined by the magnitude of the impact energy that is required to fracture a specimen using Charpy V-notch test. Generally speaking, the fracture toughness of a material decreases as the temperature decreases (i.e., it behaves more like glass). The fracture toughness at a given temperature varies with different steels and with different manufacturing and fabrication processes.

Material selection must confirm that the material has adequate fracture toughness at the lowest expected metal temperature. It is especially important for material selection to eliminate the risk of brittle fracture since a brittle fracture is catastrophic in nature. It occurs without warning the first time the necessary combination of critical size defect, low enough temperature, and high enough stress occurs.

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4.1 ASME Code and Brittle-Fracture Evaluation

The following pressure vessel components must be considered in brittle fracture evaluations:

• Shells

• Manways

• Heads

• Reinforcing pads

• Nozzles

• Tubesheets

• Flanges

• Flat cover plates

• Backing strips that remain in place

• Attachments that are essential to the structural integrity of the vessel when welded to pressure-containing components (e.g., vessel supports)

The Minimum Design Metal Temperature (MDMT) is the lowest temperature at which the component is designed to have adequate fracture toughness. It is a function of the component’s material specification and thickness. The Critical Exposure Temperature (CET) is the minimum metal temperature that can occur at the same time as a significant membrane stress in the vessel (e.g., at a pressure that is greater than 25% of the design pressure). The CET is determined by either ambient conditions or process conditions, whichever results in the lowest metal temperature. While the terms MDMT and CET are often used interchangeably, they are separate parameters.

Each component must be evaluated separately for impact test requirements based on its material, thickness, and MDMT. In all cases, the MDMT must be no greater than the CET.

Division 1 contains a simplified approach to evaluate the potential for brittle fracture in carbon and low-alloy steel. Material specifications are classified within Material Groups A through D for the purpose of brittle fracture evaluation (See Table 3.1, excerpted from Figure UCS-66 of Division 1). The Code contains exemption curves for these Material Groups that

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identify the acceptable MDMT versus thickness (0 in. through 6 in.) where impact testing (Charpy V-notch) is not required. The curves shown in Figure 3.1 are excerpted from Figure UCS-66. If the design conditions do not permit exemption in accordance with this basis, then material impact testing at the specified CET is required to permit its use. The Code specifies the necessary impact test procedure and acceptance criteria.

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Material Group Applicable Materials

Curve A • All carbon and low alloy steel plates, structural shapes, and bars not listed in Curves B, C, and D.

• SA-216 Gr. WCB and WCC, SA-217 Gr. WC6, if normalized and tempered or water-quenched and tempered.

Curve B • SA-216 Gr. WCA if normalized and tempered or water-quenched and tempered

• SA-216 Gr. WCB and WCC for maximum thickness of 2 in., if produced to fine grain practice and water-quenched and tempered

• SA-217 Gr. WC9 if normalized and tempered

• SA-285 Gr. A and B

• SA-414 Gr. A

• SA-515 Gr. 60

• SA-516 Gr. 65 and 70 if not normalized

• SA-612 if not normalized

• SA-662 Gr. B if not normalized

• Except for cast steels, all materials of Curve A if produced to fine grain practice and normalized which are not included in Curves C and D

• All pipe, fittings, forgings, and tubing which are not included in Curves C and D

• Parts permitted under Para. UG-11 shall be included in Curve B even when fabricated from plate that otherwise would be assigned to a different curve

Curve C • SA-182 Gr. 21 and 22 if normalized and tempered

• SA-302 Gr. C and D

• SA-336 Gr. F21 and F22 if normalized and tempered

• SA-387 Gr. 21 and 22 if normalized and tempered

• SA-516 Gr. 55 and 60 if not normalized

• SA-533 Gr. B and C

• SA-662 Gr. A

• All material of Curve B if produced to fine grain practice and normalized which are not included in Curve D

Curve D • SA-203

• SA-508 Cl. 1

• SA-516 if normalized

• SA-524 Cl. 1 and 2

• SA-537 Cl. 1, 2, and 3



• SA-738 Gr. A

Bolting and Nuts See Figure UCS-66 of Division 1for impact test exemption temperatures for specified material specifications.

Material Groups for Impact Test Exemptions Table 3.1

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Nominal Thickness, in.(Limited to 4 in. for Welded Construction)

0.394 1 2 3 4 5

140

120

100

80

60

40

20

0

-20

-40

-55-60

-80

Min

imum

Des

ign

Met

al T

empe

ratu

re,

F

Impact testing required

D

C

BA

Impact Test Exemption Curves for Carbon Steels Figure 3.1

A capital letter that designates the corresponding Material Group appears above each curve in Figure 3.1. If the CET of a pressure vessel is equal to or above that shown by the intersection of the Material Group curve and component thickness, then impact testing is not required. For example, a Group B material that is 1.5 in. thick does not require impact testing as long as the CET of the vessel is approximately 50°F or higher.

Division 1 has additional impact test requirements, some of which are highlighted below. Refer to the code for additional information.

• Impact testing is required for all welded construction that is over 4 in. thick if the MDMT is below 120°F.

• Impact testing is required for non-welded construction (e.g., a seamless, bolted heat exchanger cover plate) if the component is over 6 in. thick and the MDMT is below 120°F.

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• Impact testing is not required for ASME B16.5 or B16.47 ferritic steel flanges if the design metal temperature is no colder than -20°F.

• Unless specifically exempt by Fig. UCS-66, materials with a minimum yield strength greater than 65 ksi must be impact tested.

• Low temperature grades of steel that are impact tested to conform to the particular material specification (e.g., SA-333 or SA-350) may be used at design metal temperatures as low as the impact test temperature.

• If PWHT is done on P-1 material when it is not required by ASME Code rules, its impact test exemption temperature may be reduced by 30°F from that provided in Fig. UCS-66 (Ref. Para. UCS-68), as long as the resulting exemption temperature is no lower than -55°F. This recognizes the fact that a material’s fracture toughness is improved after stress relief.

• The MDMT of a vessel component may be further reduced if the general primary membrane stress in the vessel component is less than the design allowable stress. This could occur in situations where the nominal thickness of the component is greater than that required for the design conditions plus corrosion allowance (Ref. Fig. UCS-66.1).

Division 1 also contains impact-testing procedures and impact-energy requirements for cases that are subject to impact testing. Refer to Division 1 for details.

5.0 Fabricability

Fabricability refers to the ease of construction and to any special fabrication practices that are required to use the material. Of special importance is the ease with which the material can be rolled or otherwise shaped to conform to vessel component geometry requirements.

Pressure vessels commonly use welded construction. Therefore, the materials used must be weldable so that individual components can be assembled into the completed vessel. The material chemistry of the weld area must be equivalent to that of the base material so that the material properties and corrosion resistance of the weld area will be the same as those of the base material.

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B. Maximum Allowable Stress

One of the major factors in the design of pressure vessels is the relationship between the strength of the components and the loads (i.e., pressure, weight, etc.) imposed upon them. These loads cause internal stresses in the components. The design of a pressure vessel must ensure that these internal stresses never exceed the strength of the vessel components.

Pressure vessel components are designed such that the component stresses that are caused by the loads are limited to maximum allowable values that will ensure safe operation. Maximum allowable stress is the maximum stress that may be safely applied to a pressure vessel component. The maximum allowable stress includes a safety margin between the stress level in a component due to the applied loads and the stress level that could cause a failure.

1.0 Maximum Allowable Stress Criteria

The ASME Code Section II, Part D, Appendix 1 discusses the basis used to establish maximum allowable stress values for materials other than bolting for Division 1 vessels. A similar discussion is contained in Section II, Part D, Appendix 2 for bolting, and Section VIII, Division 1, Appendix P for low-temperature, cast or ductile iron materials. Refer to these appendices for the specific safety margins and other considerations used in determining the maximum allowable stresses.

Two sets of allowable stress values are provided in Division 1 for austenitic materials and for specific non-ferrous alloys. The higher alternative allowable stresses exceed two-thirds but do not exceed 90% of the minimum yield strength of the material at temperature. The higher allowable stress values should be used only where slightly higher deformation of the component is not in itself objectionable (e.g., for shell and head sections). These higher allowable stresses are not recommended for the design of flanges or other strain-sensitive applications. In the case of flanges, for example, the larger deformation that would be expected if the higher allowable stresses were used could cause flange leakage problems even though a major flange failure would not occur.

2.0 ASME Maximum Allowable Stress Tables

Tables in the ASME Code Section II, Part D contain the maximum allowable tensile stresses of materials that are acceptable for use in

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ASME Code Section VIII pressure vessels. The maximum allowable stress varies with temperature because material strength is a function of temperature.

Figure 3.2 (adapted from Table 1A of the ASME Code Section II, Part D) shows examples of maximum allowable Division 1 tensile stress for three different material specifications.

The first part of Figure 3.2 identifies the Spec. No. (i.e., material specification number), the grade (a material specification may have multiple strength grades), the nominal chemical composition, the P-No. and Group No., and the minimum yield and tensile strengths in thousands of pounds per square inch (ksi). This first part of Figure 3.2 also helps identify similarities that may exist among the material specifications (e.g., nominal alloy composition, yield strength, and tensile strength). In some cases, these similarities may help select the material to use for pressure vessel fabrication, given specific process conditions. The maximum allowable stress va lues as a function of temperature are presented in the second part of Figure 3.2.

The information that is contained in the ASME Code Table 1A has been condensed and reorganized in Figure 3.2 in two parts to help Participants compare the material types and to note variances in maximum allowable stress that are determined by temperature and alloy composition.

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ALLOWABLE STRESS IN TENSION FOR CARBON AND

LOW-ALLOY STEEL Spec No. Grade Nominal

Composition P-No. Group No. Min. Yield

(ksi) Min. Tensile

(ksi) Carbon Steel Plates and Sheets

SA-515 55 C-Si 1 1 30 55 60 C-Si 1 1 32 60 65 C-Si 1 1 35 65 70 C-Si 1 2 38 70 SA-516 55 C-Si 1 1 30 55 60 C-Mn-Si 1 1 32 60 65 C-Mn-Si 1 1 35 65 70 C-Mn-Si 1 2 38 70

Plate - Low Alloy Steels SA-387 2 Cl.1 ½ Cr-½ Mo 3 1 33 55 2 Cl.2 ½ Cr-½ Mo 3 2 45 70 12 Cl.1 1Cr-½ Mo 4 1 33 55 12 Cl.2 1Cr-½ Mo 4 1 40 65 11 Cl.1 1 ¼ Cr-½Mo-Si 4 1 35 60 11 Cl.2 1 ¼ Cr-½Mo-Si 4 1 45 75 22 Cl.1 2 ¼ Cr-1Mo 5 1 30 60 22 Cl.2 2 ¼ Cr-1Mo 5 1 45 75

ASME Maximum Allowable Stress (Table 1A Excerpt) Figure 3.2

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ALLOWABLE STRESS IN TENSION FOR CARBON AND LOW ALLOY STEEL

Max Allowable Stress, ksi (Multiply by 1,000 to Obtain psi) for Metal Temperature, °F, Not Exceeding

650

700

750

800

850

900

950

1000

1050

1100

1150

1200

Spec No.

Carbon Steel Plates and Sheets 13.8 13.3 12.1 10.2 8.4 6.5 4.5 2.5 -- -- -- -- SA-515 15.0 14.4 13.0 10.8 8.7 6.5 4.5 2.5 -- -- -- -- SA-515 16.3 15.5 13.9 11.4 9.0 6.5 4.5 2.5 -- -- -- -- SA-515 17.5 16.6 14.8 12.0 9.3 6.5 4.5 2.5 -- -- -- -- SA-515

13.8 13.3 12.1 10.2 8.4 6.5 4.5 2.5 -- -- -- -- SA-516 15.0 14.4 13.0 10.8 8.7 6.5 4.5 2.5 -- -- -- -- SA-516 16.3 15.5 13.9 11.4 9.0 6.5 4.5 2.5 -- -- -- -- SA-516 17.5 16.6 14.8 12.0 9.3 6.5 4.5 2.5 -- -- -- -- SA-516

Plate-Low Alloy Steels (Cont'd)

13.8 13.8 13.8 13.8 13.8 13.3 9.2 5.9 -- -- -- -- SA-387 17.5 17.5 17.5 17.5 17.5 16.9 9.2 5.9 -- -- -- -- SA-387 13.8 13.8 13.8 13.8 13.4 12.9 11.3 7.2 4.5 2.8 1.8 1.1 SA-387 16.3 16.3 16.3 16.3 15.8 15.2 11.3 7.2 4.5 2.8 1.8 1.1 SA-387 15.0 15.0 15.0 15.0 14.6 13.7 9.3 6.3 4.2 2.8 1.9 1.2 SA-387 18.8 18.8 18.8 18.8 18.3 13.7 9.3 6.3 4.2 2.8 1.9 1.2 SA-387 15.0 15.0 15.0 15.0 14.4 13.6 10.8 8.0 5.7 3.8 2.4 1.4 SA-387 17.7 17.2 17.2 16.9 16.4 15.8 11.4 7.8 5.1 3.2 2.0 1.2 SA-387

ASME Maximum Allowable Stress (Excerpt), cont'd

Figure 3.2, cont'd

Note that the allowable stresses at temperatures between -20°F and 650°F are the same as the allowable stress at 650°F for each material presented in Figure 3.2 (except for SA-387, Grade 22 Cl. 2). The allowable stress increases for SA–387, Grade 22 Cl. 2 material at temperatures below 650°F to a maximum of 18.8 ksi at 100°F and below. Note that each material specification has different Types, Grades, and/or Classes within it. In some cases, these differences are due to different chemical compositions, while in other cases they may be due to the particular steel making process that is employed. Higher strength grades of a particular material specification have higher maximum allowable stresses.

91

Exercise 1

Material Selection Based on Fracture Toughness

A new horizontal pressure vessel is being designed for an application where the CET is -2°F. The material being used for the shell and heads is SA-516 Gr. 70 plate. The heads are hemispherical in shape and are ½ in. thick. The cylindrical shell is 1.0 in. thick. The supplier has not specified any impact testing for the shell and head plate. Is this correct? If this is not correct, what should be done to correct the situation?

92

IV. Design

A. Design Conditions and Loadings

The mechanical design of a pressure vessel begins with specification of the design pressure and design temperature. Pressure imposes loads that must be withstood by the individual vessel components. Temperature affects material strength and, thus, its allowable stress, regardless of the design pressure. Some pressure vessels have multiple sets of design conditions that correspond to different modes of operation. For example, during its operating cycle, a reactor may have a high pressure and moderate temperature during normal operation, but it may operate at a much lower pressure and a very high temperature during catalyst regeneration. Both sets of design conditions must be specified because either one or the other may govern the mechanical design.

All pressure vessels must be designed for the most severe conditions of coincident pressure and temperature that are expected during normal service. Normal service includes conditions that are associated with:

• Startup.

• Normal operation.

• Deviations from normal operation that can be anticipated (e.g., catalyst regeneration or process upsets).

• Shutdown.

Pressure vessels must also be designed for other loading conditions and service factors that may apply in particular situations. These are highlighted later.

93

1.0 Pressure 1.1 Operating Pressure

The operating pressure must be set based on the maximum internal or external pressure that the pressure vessel may encounter. The following factors must be considered:

• Ambient temperature effects.

• Normal operational variations.

• Pressure variations due to changes in the vapor pressure of the contained fluid.

• Pump or compressor shut-off pressure.

• Static head due to the liquid level in the vessel.

• System pressure drop.

• Normal pre-startup activities or other operating conditions that may occur (e.g., vacuum), that should be considered in the design.

1.2 Design Pressure

Generally, design pressure is the maximum internal pressure (in psig), that is used in the mechanical design of a pressure vessel. For full or partial vacuum conditions, the design pressure is applied externally and is the maximum pressure difference that can occur between the atmosphere and the inside of the pressure vessel. Some pressure vessels may experience both internal and external pressure conditions at different times during their operation. The mechanical design of the pressure vessel in this case is based on which of these is the more severe design condition.

The specified design pressure is based on the maximum operating pressure at the top of the vessel, plus the margin that the process design engineer determines is suitable for the particular application. A suitable margin must also be provided between the maximum operating pressure and the safety relief valve set pressure. This margin is necessary to prevent frequent and unnecessary opening of the safety relief valve that may occur during normal variations in operating pressure. The safety relief valve set pressure is normally set equal to the design pressure.

94

Pressure vessels, especially tall towers, may have liquid in them during normal operation. The maximum height of this liquid normally does not reach the top of the vessel. The liquid level that is required for design is specified by the process design engineer.

The hydrostatic pressure that is exerted by the liquid must be considered in the design of vessel components upon which it acts. Therefore, the pressure that is used to design a vessel component is equal to the design pressure at the top of the vessel, plus the hydrostatic pressure of the liquid in the vessel that is above the point being designed (i.e., PBH = PT + γH). See Figure 4.1.

95

PT = Design Pressure atTop of Vessel

PBH = Design Pressure ofBottom Head

H = Height of Liquid

= Weight Density of Liquid in Vessel

γ

Design Pressure Figure 4.1

2.0 Temperature 2.1 Operating Temperature

The operating temperature must be set based on the maximum and minimum metal temperatures that the pressure vessel may encounter. The operation and vertical length of tall towers, and the presence of liquid in the bottom section, sometime result in large temperature reductions between the bottom and top of the vessel. It is permissible to specify different operating

96

temperatures at different elevations of such a pressure vessel, as long as the temperatures can be accurately predicted. This approach results in dividing the vessel into sections along its vertical length. Each section is designed for the temperature that it will encounter, rather than for the most severe condition at the bottom of the vessel. Figure 4.2 illustrates this concept.

Support Skirt

Grade

Section 2(T-X)

Section 3(T-Y)

Section 4(T-Z)

Section 1(T) F

Temperature Zones in Tall Vessels Figure 4.2

2.2 Design Temperature

The design temperature of a pressure vessel is the maximum fluid temperature that occurs under normal operating conditions, plus an allowance for variations that occur during operation.

2.3 Critical Exposure Temperature (CET)

The CET must also be specified for pressure vessel design to ensure that materials that have adequate fracture toughness are selected for construction (i.e., MDMT ≤ CET). Fracture toughness was previously discussed.

97

3.0 Other Loadings

Paragraph UG-22 of Division 1 specifies the loadings that must be considered to determine the minimum required thicknesses for the various vessel components. These design loadings are as follows:

• Internal or external design pressure.

• Weight of the vessel and its normal contents under operating or test conditions.

• Superimposed static reactions from the weight of attached equipment (e.g., motors, machinery, other vessels, piping, linings, insulation).

• Loads at attached of internal components or vessel supports.

• Wind, snow, and seismic reactions.

• Cyclic and dynamic reactions that are caused by pressure or thermal variations, or by equipment that is mounted on a vessel, and mechanical loadings.

• Test pressure combined with hydrostatic weight.

• Impact reactions such as those that are caused by fluid shock.

• Temperature gradients within a vessel component and differential thermal expansion between vessel components.

B. Weld Joint Efficiency and Corrosion Allowance

The weld joint efficiency and corrosion allowance are additional design parameters that are required to calculate vessel component thicknesses.

1.0 Weld Joint Efficiency

Weld joint efficiency (E) accounts for the quality of a welded joint and for the concentration of local stress. This higher local stress is due to local material or structural discontinuities.

Paragraph UW-12 of Division 1 specifies weld joint efficiencies to be used to calculate component thicknesses. Figure 4.3 identifies weld joint categories, Figure 4.4 identifies weld types, and Figure 4.5 defines weld joint efficiencies based on the type of weld and degree of radiographic examination.

98

The majority of pressure vessel welds use a Type 1 joint design. A Type 1 joint has an efficiency of either 0.85 or 1.00, corresponding with spot or full radiographic examination, respectively.

B AC

D

D B

C

A

B

A

B

D

D CA

A

C

AC

C

BD

Weld Joint Categories

Figure 4.3

2.0 Corrosion Allowance

Corrosion, erosion, or abrasion causes vessel components to thin during their operating life. To compensate for this thinning, components must have their thicknesses increased over those that are calculated using the ASME Code design formulas. Internal corrosion/erosion-resistant linings are sometimes used as an alternative to the use of greater component thicknesses.

Process design and materials engineers typically specify the corrosion allowance. The corrosion allowance is based on the expected corrosion rate for the vessel material in the anticipated process environment. The corrosion rate is multiplied by the nominal design life of the vessel (normally 20 years) to determine the corrosion allowance.

C. Design for Internal Pressure

1.0 Cylindrical Shells

The idealized equations for the calculation of hoop and longitudinal stresses, respectively, in a cylindrical shell under internal pressure are as follows:

tPr

=σθ

and t2

Pr1 =σ

99

These equations assume a uniform stress distribution through the thickness of the shell. Note that the longitudinal stress is half the hoop stress. Since this is an idealized state, the ASME Code formulas (See Figure 4.6) have been modified to account for non-ideal behavior.

1

2

3

4

5

6

Butt joints as attained by double-welding or by othermeans which will obtain the same quality of depositedweld metal on the inside and outside weld surface.

Backing strip, if used, shall be removed aftercompletion of weld.

Single-welded butt joint with backing strip whichremains in place after welding.

For circumferentialjoint only

Single-welded butt joint without backing strip.

Double-full fillet lap joint.

Single-full fillet lap joint with plug welds.

Single-full fillet lap joint without plug welds.

Types of Welded Joints

Figure 4.4

100

Joint Type

Acceptable Joint Categories

Degree of

Radiographic Examination Full Spot None 1 A, B, C, D 1.00 0.85 0.70

2 A, B, C, D (See ASME Code for limitations) 0.90 0.80 0.65

3 A, B, C NA NA 0.60

4 A, B, C (See ASME Code for limitations) NA NA 0.55

5 B, C (See ASME Code for limitations) NA NA 0.50

6 A, B, (See ASME Code for limitations) NA NA 0.45

Maximum Weld Joint Efficiency Figure 4.5

Longitudinal stress can govern the design of a cylindrical section when loadings other than internal pressure induce longitudinal stresses that are greater than one half of the hoop stress due to internal pressure. One example where this could occur is in the lower section of a ta ll tower where wind or earthquake loading could cause high longitudinal stresses. In these cases, the longitudinal stress that is due to these other loads is added to the longitudinal stress due to internal pressure. The total combined longitudinal stress is then limited to the maximum allowable stress.

Figure 4.6 summarizes the ASME Code equations used to calculate the minimum required thickness for common pressure vessel components. The equations have also been rearranged to calculate pressure and stress as a function of thickness.

101

Part

Thickness, tp, in.

Pressure, P, psi

Stress, S, psi

Cylindrical shell

P6.0SEPr

1 − t6.0rtSE1

+

( )1tE

t6.0rP +

Spherical shell

P2.0SE2Pr

1 − t2.0rSEt2

+

( )tE2

t2.0rP +

2:1 Semi -Elliptical head

P2.0SE2PD−

t2.0D

SEt2+

( )

tE2t2.0DP +

Torispherical head

with 6% knuckle P1.0SEPL885.0

−

t1.0L885.0SEt

+

( )tE

t1.0L885.0P +

Conical Section

(α = 30°) ( )P6.0SEcos2PD

−α α+α

cost2.1DcosSEt2

( )

αα+

costE2cost2.1DP

Summary of ASME Code Equations Figure 4.6

Where:

P = Internal design pressure, psig. When used in the pressure calculation equations, this is the MAWP.

r = Internal radius, in. Add corrosion allowance to specified uncorroded internal radius.

S = Allowable Stress, psi. When used in the thickness calculation equations, this is the allowable stress for the material used. When used in the stress calculation equations, this is the calculated stress for the given pressure and thickness.

E1, E = Longitudinal weld joint efficiency

tp = Required wall thickness for internal pressure of the part under consideration, in.

t = Actual wall thickness (less corrosion allowance) of the part under consideration, in.

D = Inside diameter, in. Add twice the corrosion allowance to specified uncorroded inside diameter.

DL = Cone inside diameter at large end, in. Add twice the corrosion allowance to specified uncorroded inside diameter.

102

DS = Cone inside diameter at small end, in. Add twice the corrosion allowance to specified uncorroded inside diameter.

L = Inside crown radius of torispherical head, in. Add corrosion allowance to specified uncorroded inside crown radius.

α = One half of the apex angle of the cone at the centerline, degrees.

)LengthCone()DD(5.0

tan sL1 −=α −

2.0 Heads

Figure 4.7 shows typical types of closure heads. Elliptical, hemispherical, and torispherical are the most commonly used head types. Note in Figure 4.7 that all head types but the conical have a straight flange (sf) section, which simplifies welding the head to the adjacent cylindrical shell section. The elliptical and torispherical heads have an indicated head depth (h), which is measured from the straight flange to the maximum point of curvature on the inside surface.

103

t

sf

ID

Flanged

t

sf

R

ID

Hemispherical

sf

t

h

Flanged and Dished(torispherical)

t

sf

h

Elliptical

tα

IDConical Toriconical

IDr

sf

tα

Typical Formed Closure Heads Figure 4.7

As with shells, the internal head dimensions that are used to calculate the required thicknesses must first be increased to account for the corrosion allowance. The corrosion allowance must then be added to the calculated thicknesses. See Figure 4.6 for the ASME Code equations that are used to calculate the wall thickness of each head type.

104

2.1 Elliptical Heads - The 2:1 semi-elliptical head is the most commonly used head type. Half of its minor axis (i.e., the inside depth of the head minus the length of the straight flange section) equals one-fourth of the inside diameter of the head. The thickness of this type of head is normally equal to the thickness of the cylinder to which it is attached.

2.2 Hemispherical Heads - The required thickness of a hemispherical head is normally one-half the thickness of an elliptical or torispherical head for the same design conditions, material, and diameter. Hemispherical heads are normally fabricated from segmented sections that are welded together, spun, or pressed. Hemispherical heads are an economical option to consider when expensive alloy material is used. In carbon steel construction, hemispherical heads are generally not as economical as elliptical or torispherical heads because of higher fabrication cost. Carbon steel hemispherical heads may be economical for thin, very large-diameter vessels, or in thick, small-diameter vessels.

The thickness transition zone between the hemispherical head and shell must be contoured to minimize the effect of local stress. Figure 4.8 shows the thickness transition requirements that are contained in the ASME Code.

2.3 Torispherical Heads - A torispherical (or flanged and dished) head is typically somewhat flatter than an elliptical head and can be the same thickness as an elliptical head for identical design conditions and diameter. The minimum permitted knuckle radius of a torispherical head is 6% of the maximum inside crown radius. The maximum inside crown radius equals the outside diameter of the head.

105

Length of required taper, l,may include the width

of the weld

y y

tsts

Tangent Line

th th

y3l ≥y3l ≥

Thin

ner

Par

t

Thin

ner

Par

t

Thickness Transition Between Hemispherical Head and Shell Figure 4.8

2.4 Intermediate Heads – An intermediate head may be installed inside a pressure vessel to separate two sections that can have different design conditions. Most head types can be used as intermediate heads. Intermediate heads are evaluated for internal pressure in the same way as external heads.

3.0 Conical Sections

Tall towers may have sections with different diameters along their length. The transition between the different diameters is made in a conical section. The most common design for a conical transition does not have formed knuckles at the ends of the cone. The cylindrical sections of different diameter are welded to each end of the cone. The required thickness for internal pressure of a conical shell without transition knuckles is calculated using the equation shown in Figure 4.6. This equation assumes that half of the cone-apex angle is no greater than 30°.

Formed knuckles are sometimes used at the cone-to-cylinder transition in order to reduce localized stresses. When knuckles are used, the transition is called toriconical. The use of knuckles is

106

mandatory when the cone half-apex angle exceeds 30°. Knuckles are also sometimes used for smaller angles when there is concern about potentially high local stresses at the cone-to-cylinder junction. The ASME Code has design procedures for toriconical sections.

107

Sample Problem 1 – Design for Internal Pressure

The geometry and design data of a vertical cylindrical pressure vessel are specified in Figure 4.9. Cost estimates are being prepared for this vessel. It is your job to estimate the required component thicknesses. A. What are the minimum required thicknesses for the two cylindrical sections?

Hemispherical

4' - 0"

60' - 0"

10' - 0"

6' - 0"30' - 0"

2:1 Semi-Elliptical

DESIGN INFORMATIONDesign Pressure = 250 psigDesign Temperature = 700° FShell and Head Material is SA-515 Gr. 60Corrosion Allowance = 0.125"Both Heads are SeamlessShell and Cone Welds are Double Welded and will be Spot RadiographedThe Vessel is in All Vapor ServiceCylinder Dimensions Shown are Inside Diameters

Sample Problem 1 Figure 4.9

108

Solution

1. The required wall thickness for internal pressure of a cylindrical shell is calculated using the following equation from Figure 4.6:

P6.0SEPr

t1

p −=

2. Since the welds are spot radiographed, E = 0.85 (from Figure 4.5)

3. S = 14,400 psi for SA-515/Gr. 60 at 700°F (from Figure 3.2)

4. P is given as 250 psig.

5. For the 6 ft. - 0 in. shell, calculate r (including corrosion allowance)

r = 0.5D + CA = 0.5 x 72 + 0.125 = 36.125 in.

2506.085.0400,14125.36250

P6.0SEPr

t1

p ×−××

=−

= = 0.747 in.

t = tp + c = 0.747 + 0.125

t = 0.872 in. required including corrosion allowance

6. For the 4 ft. - 0 in. shell, calculate r (including corrosion allowance)

r = 0.5 x 48 + 0.125 = 24.125 in.

2506.085.0400,14125.24250

t p ×−××

= = 0.499 in.

t = 0.499 + 0.125

t = 0.624 in. required (including corrosion allowance)

109

B. For the same vessel, what are the minimum required thicknesses for the top and bottom heads?

Solution

1. Since both heads are seamless, E = 1.0.

2. Top Head - Hemispherical head (Equation from Figure 4.6)

r = 24 + 0.125 = 24.125 in.

2502.01400,142125.24250

P2.0SE2Pr

t1

p ×−×××

=−

= = 0.21 in.

t = tp + c = 0.21 + 0.125


3. Bottom Head - 2:1 Semi-Elliptical Head (Equation from Figure 4.6)

D = 72 + 2 x 0.125 = 72.25 in.

in. 0.628 2502.01400,142

25.72250P2.0SE2

PDt p =

×−×××

=−

=

t = 0.628 + 0.125


110

D. Design for External Pressure and Compressive Stresses

Pressure vessels are subject to compressive forces such as those caused by dead weight, wind, earthquake, and internal vacuum. Pressure vessel components behave differently under these compressive forces than when they are exposed to tensile forces (e.g., from internal pressure). This difference in behavior is due to elastic instability, which makes shells weaker in compression than in tension. In failure by elastic instability, the vessel is said to collapse or buckle. The paragraphs that follow discuss buckling of cylindrical shells due to external pressure. These basic principles also apply to other forms of shells as well as to heads and to compressive loads other than external pressure.

1.0 Overview

The critical pressure that causes buckling is not a simple function of the stress that is produced in the shell, as is true with tensile loads. An allowable stress is not used to design pressure vessels that are subject to elastic instability. Instead, the design is based on the prevention of elastic collapse under the applied external pressure. This applied external pressure is normally 15 psig for full vacuum conditions.

The maximum allowable external pressure can be increased by welding circumferential stiffening rings (i.e., stiffeners) around the vessel shell. The addition of stiffening reduces the effective buckling length of the shell, and this length reduction increases the allowable buckling pressure. These stiffener rings may be welded on either the inside or the outside of the shell. Figure 4.10 illustrates the use of stiffeners on a pressure vessel cylinder.

Other factors also affect the design of a pressure vessel for external pressure since they also influence its resistance to buckling.

• At elevated temperature, the material stress-strain curves are nonlinear with no definite yield point and with a variable modulus of elasticity.

• The shell diameter and thickness are additional geometric parameters that affect shell stiffness.

Paragraphs UG-28 and UG-33 of Division 1 contain procedures to calculate the allowable external pressure on cylindrical shells and

111

heads, respectively. These calculation procedures use an iterative approach.

LLLLL

LLLLL

Moment Axis of Ring

h/3

h/3

h = Depth of Head

Stiffener Rings on Pressure Vessel Cylinders Figure 4.10

The maximum allowable compressive stress in a pressure vessel component that is due to loads other than external pressure is limited to the lower of the following:

• The allowable tensile stress, or

• A value, Factor B (See Figure 4.13), determined using the external pressure calculation procedure.

2.0 Shells

The allowable external pressure of a cylindrical shell is a function of material, design temperature, outside diameter, corroded thickness, and unstiffened length. See Division 1 for procedural details.

3.0 Heads

The allowable external pressure of a head is a function of material, design temperature, outside radius, head depth, and corroded thickness. Stiffening rings a re not used to increase the allowable external pressure of heads. The head thickness is increased as required to achieve the required external pressure. When an intermediate head is installed inside a pressure vessel, it may be

112

necessary to design it for an external pressure that is higher than 15 psig. See Division 1 for procedural details.

4.0 Conical Sections

The allowable external pressure of a conical section is a function of material, design temperature, outside diameters at the small and large ends, conical section length, apex angle, and corroded thickness. The allowable external pressure may be increased by the addition of stiffener rings, or by increasing the cone thickness. See Division 1 for procedural details.

113

Sample Problem 2 - External Pressure Calculation

This Sample Problem demonstrates the external pressure design procedure for one example of a cylindrical pressure vessel shell. Refer to Division 1 for additional details and procedures to use for heads and conical shells.

A tall cylindrical tower is being supplied. The geometry and design conditions are specified in Figure 4.11. The vendor has proposed that the wall thickness of this tower be 7/16 in., and no stiffener rings have been specified. Is the 7/16 in. thickness acceptable for external pressure? If it is not acceptable, what minimum thickness is required? Round your answer upward to the nearest 1/16 in.

2:1 Semi-Elliptical(Typical)

150' - 0"

4' - 0"

DESIGN INFORMATIONDesign Pressure = Full VacuumDesign Temperature = 500° FShell and Head Material is SA-285 Gr. B, Yield Stress = 27 ksiCorrosion Allowance = 0.0625"Cylinder Dimension Shown is Inside Diameter

Sample Problem 2 - Solution Figure 4.11

Solution

1. First, calculate the unstiffened design length, L, and the outside diameter, Do, of the cylindrical shell, both in inches.

114

L = Tangent Length + 2 × 1/3 (Head Depth)

The tangent length is given as 150 ft.

Since the heads are semi-elliptical, the depth of each head is equal to ¼ the inside diameter of the shell.

Head Depth = 48 /4 = 12 in.

L = 150 × 12 + 2/3 × 12 = 1,808 in.

Calculate outside diameter Do, in.

Do = 48 + 2 × 7/16 = 48.875 in.

Next, determine the ratios L/Do and Do/t.

Accounting for the corrosion allowance, t = 7/16 – 1/16 = 6/16 = 0.375 in.

Do/t = 48.875 / 0.375 = 130

L/Do = 1808 / 48.875 = 37

2. Determine the value of A using Figure 4.12 and the calculated Do/t and L/Do.

Note: If L/Do > 50, use L/Do = 50. For L/Do < 0.05, use L/Do = 0.05.

115

2

3

4

5

6 7

8 9

50.0

40.0

35.0

30.0

25.0

20.0

18.0

16.0

14.0

12.0

10.0

9.0

8.0

7.0

6.0

5.0

4.0

3.5

3.0

2.5

2.0

1.8

1.6

1.4

1.2

.00

00

1

.0

00

1

Length + Outside Diameter = L/Do

D o/t =

400

Do/t = 100

Do/t = 125

Do/t = 150

Do/t = 200

Do/t = 250

Do/t = 300 D o/t =

500

D o/t =

600

D o/t =

800

D o/t =

1,000

Do/t = 130

L/Do = 37

A = 0.000065

Factor A Figure 4.12

3. Move horizontally to the line for the value of Do/t = 130 determined in Step 2. Use interpolation for intermediate values of Do/t. Move vertically downward from this intersection point to determine Factor A.

A = 0.000065

4. Using the value of A from Step 4, enter the applicable material chart. For this case, the applicable material chart is Figure CS-1, excerpted in Figure 4.13. Move vertically in this chart to the intersection with the correct design temperature line. Use interpolation for intermediate temperatures. Note that in this case, the value of A is to the left of all the temperature curves.

116

2 3 4 5 6 7 8 9 2 3 4 5 6 7 8 9 2 3 4 5 6 7 8 9 2 3 4 5 6 7 8 9

20,00018,00016,000

14,000

12,000

10,0009,0008,000

7,000

6,000

5,000

4,000

3,5003,000

2,500

2,000

.00001 .0001 .001 .01 .1

FACTOR A

FA

CTO

R B

GENERAL NOTE: See Table CS-1 for tabular valuesup to 300°F

900°F

500°F

700°F

800°F

E-29.0 = 106

E-27.0 = 106

E-24.5 = 106

E-22.8 = 106

E-20.8 = 106

A=0.000065

Figure CS-1 Figure 4.13

5. Calculate maximum allowable external pressure for the value of t, psi.

)t/D(3AE2

Po

a =

Where:

E = Young's modulus of elasticity at design temperature for the material, psi. Do not confuse this parameter with the weld joint efficiency, E, that is used elsewhere.

E = 27 x 106 psi from Figure CS-1 (Figure 4.13) at T = 500°F

33.13031027000065.02

P6

a ××××

=

Pa = 9 psi

117

Since the calculated Pa < 15 psi, the proposed 7/16 in. shell thickness is not sufficient.

Note: In cases where A is located under the temperature curves, determine the Factor B by reading horizontally across from the intersection point. Then determine the maximum allowable external pressure, Pa, from the following equation:

)t/D(3B4

Po

a =

6. Now determine how thick the shell must be in order to have Pa ≥15 psi. This is a trial-and-error process, by which the thickness is increased until an acceptable value is found. The intent is to use the thinnest shell that will meet the requirement. Without going through all the iterations, we will assume a new shell thickness of 9/16 in. and thus a corroded thickness of ½ in.

75.975.0875.48

tDo ==

)beforeas(37

DL

o

=

A = 0.000114

psi7.1533.1303

1027000114.02P

6

a =×

×××=

118

Exercise 2 Required Thickness for Internal Pressure

Determine the minimum required thickness for the cylindrical shell and heads of the following pressure vessel:

• Inside Diameter - 10’ - 6”

• Design Pressure - 650 psig

• Design Temperature - 750°F

• Shell & Head Material - SA-516 Grade 70

• Corrosion Allowance - 0.125”

• 2:1 Semi-Elliptical heads, seamless

• 100% radiography of cylindrical shell welds

• The vessel is in an all vapor service (i.e., no liquid loading)

119

E. Reinforcement of Openings

Calculation of the required wall thickness of a nozzle is one step in the design of openings in pressure vessels. This is done in the same manner as for any other cylindrical shell. There is more to the design of openings than calculating the nozzle thickness, cutting a hole in the vessel, and welding the nozzle in.

The ASME Code uses simplified rules to ensure that the membrane stresses are kept within acceptable limits when an opening is made in a vessel shell or head.

Dp

tn Rntrn

tr

c

hd

te

d or Rn + tn + t

Use larger value

t

2.5t or 2.5tnUse smaller value

2.5t or 2.5tn + teUse smaller value

d or Rn + tn + t

Use larger value

For nozzle wall abuttingthe vessel wall

For nozzle wall insertedthrough the vessel wall

Cross-Sectional View of Nozzle Opening Figure 4.14

When the opening is made, a volume of material is removed from the pressure vessel. This metal is no longer available to absorb the applied loads. The ASME Code simplifies the design calculations by viewing the nozzle-to-vessel junction area in cross section (See Figure 4.14). This simplification permits the nozzle reinforcement calculations to be made in terms of metal cross-sectional area rather than metal volume. The ASME Code requires that the metal area that is removed for the opening must be replaced by an equivalent metal area in order for the opening to be adequately reinforced. The replacement metal must be located adjacent

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to the opening within defined geometric limits. The replacement metal area may come from two sources:

• Excess metal that is available in the shell or nozzle neck that is not required for pressure or to absorb other loads.

• Reinforcement that is added to the shell or nozzle neck.

Figure 4.15 shows several typical nozzle design configurations including examples of inserted versus abutted nozzles, pad reinforcement versus no reinforcement, and self-reinforced nozzles. Self-reinforced nozzles are forged fittings that have extra thickness in the nozzle-to-vessel junction area to provide reinforcement.

Additional reinforcement must be provided if the vessel shell and nozzle do not have sufficient excess thickness that is not required for pressure or other loads. Additional reinforcement can be in one of the following forms:

• A reinforcement pad.

• Additional thickness in the vessel shell or head.

• Additional thickness in the nozzle near its attachment to the vessel.

The reinforcement must be located within defined boundaries in order for it to be considered effective.

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(a)Full Penetration Weld

With Integral Reinforcement (a-1) (a-2) (a-3)

Separate Reinforcement Plates Added

(b) (c) (d) (e)

Full Penetration Welds to Which Separate Reinforcement Plates May be Added

(f-1) (f-3)

(f-2)(f-4) (g)

Self - Reinforced Nozzles Typical Nozzle Design Configurations

Figure 4.15

If a reinforcement pad is used, its material should have an allowable stress that is at least equal to that of the pressure vessel shell or head material to which it is attached. No credit can be taken for the additional strength of any reinforcement that has a higher allowable stress. If reinforcement material with a lower allowable stress is used, the reinforcement area must be increased to compensate for this.

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The ASME Code specifies circumstances under which no nozzle reinforcement eva luations are needed. It also provides rules to evaluate the reinforcement of openings that are located near each other. These situations are not discussed in this course. Refer to the ASME Code for details. Sample Problem 3 illustrates the procedure used to evaluate nozzle reinforcement.

Sample Problem 3 - Reinforcement of Openings

You are reviewing the nozzle design details that are proposed by a vendor for a new drum and have selected an NPS 8 nozzle into the shell for detailed evaluation. The vendor has not provided any reinforcement for this nozzle, and he has not provided any calculations to verify that use of the nozzle without reinforcement is acceptable.

Determine if this nozzle requires additional reinforcement. If it does, assume that a 0.5 in. thick reinforcement pad of SA-516, Gr. 60 material is used. What must the minimum pad diameter be? Neglect any contribution of weld areas in these calculations since they are insignificant. The information that is needed to perform your evaluation is in Figure 4.16. Use Figure 4.14 as a reference.

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NPS 8 Nozzle(8.625" OD)0.5" Thick

0.5625" Thick Shell, 48" Inside Diameter

DESIGN INFORMATIONDesign Pressure = 300 psigDesign Temperature = 200°FShell Material is SA-516 Gr. 60Nozzle Material is SA-53 Gr. B, SeamlessCorrosion Allowance = 0.0625"Vessel is 100% RadiographedNozzle does not pass through Vessel Weld Seam

Sample Problem 3 Figure 4.16

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Solution

Calculate the required reinforcement area, A

A = dtrF

Where:

d = Finished diameter of circular opening, or finished dimension (chord length at mid surface of thickness excluding excess thickness available for reinforcement) of nonradial opening in the plane under consideration, in.

tr = Minimum required thickness of the shell using appropriate ASME Code formula and a weld joint efficiency of 1.0, in.

F = Correction factor normally equal to 1.0.

Calculate the diameter, d.

d = Diameter of Opening – 2 (Thickness + Corrosion Allowance)

d = 8.625 – 1.0 + .125 = 7.750 in.

Calculate the required thickness of the shell, tr (See Figure 4.6)

.in487.03006.01000,15

)0625.024(300P6.0SE

Prt

1r =

×−×+×

=−

=

Assume a value of 1.0 for F.

Calculate the required reinforcement area, A

A = dtrF

A = (8.625 - 1.0 + 0.125) × 0.487 × 1 = 3.775 in.2 required area

Calculate the available reinforcement area in the vessel shell, A1, as the larger of A11 or A12

A11 = (Elt - Ftr)d

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A12 = 2 (Elt-Ftr)(t + tn)

Where:

El = 1.0 when the opening is in the base plate away from the welds, or when the opening passes through a circumferential joint in the shell (excluding head to shell joints).

El = The ASME Code joint efficiency when any part of the opening passes through any other welded joint.

F = 1 for all cases except integrally reinforced nozzles that are inserted into a shell or cone at an angle to the vessel longitudinal axis. See Fig. UG-37 for this special case.

tn = Nominal thickness of the nozzle in the corroded condition, in.

A11 = (Elt - Ftr)d = (0.5625 - 0.0625 - 0.487) x 7.75 = 0.1 in.2

A12 = 2(Elt - Ftr) (t + tn)

= 2(0.5625 - 0.0625 - 0.487) (0.5625 - 0.0625 + 0.5 - 0.0625)

= 0.0243 in.2

Therefore, A1= 0.1 in.2 available reinforcement in shell

Calculate the reinforcement area that is available in the nozzle wall, A2, as the smaller of A21 or A22.

A21 = (tn-trn)5t

A22 = 2(tn-trn)(2.5 tn + te)

Where:

trn = Required thickness of the nozzle wall, in.

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r = radius of the nozzle, in.

te = 0 if there is no reinforcing pad.

te = Reinforcing pad thickness if one is installed, in.

te = As defined in Figure UG-40 of the ASME Code for self-reinforced nozzles, in.

Calculate the required thickness of the nozzle, trn (See Figure 4.6)

P6.0SEPr

t1

rn −=

( ).in0784.0

3006.01000,150625.08125.3300

tm =×−×

+=

Calculate the available reinforcement in the nozzle neck, A2, as the smaller of A21 or A22.

A21 = (tn - trn)5t = (0.5 - 0.0625 - 0.0784) x 5(0.5625 - 0.0625)

A21 = 0.898 in.2

A22 = 2(tn - trn) (2.5 tn + te)

= 2(0.5 - 0.0625 - 0.0784) [2.5 x (0.5 - 0.0625) + 0]

= 0.786 in.2

Therefore, A2 = 0.786 in.2 available reinforcement in nozzle

Determine the total available reinforcement area, AT, and compare it to the required area.

AT = A1 + A2 = 0.1 + 0.786 = 0.886 in.2

Since AT < A, the nozzle is not adequately reinforced, and a reinforcement pad is required.

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Determine the required reinforcement pad area, A5, and pad diameter, Dp.

Since the required reinforcement area is 3.775 in.2 and the available reinforcement area is 0.886 in.2, we need to calculate the required area for the reinforcement pad.

A5 = A - AT

A5 = (3.775 - 0.886) = 2.889 in.2 required area in reinforcement pad.

Now, calculate Dp

te = 0.5625 in. (reinforcement pad thickness)

A5 = [Dp - (d + 2 tn)] te

2.889 = [Dp - (7.75 + 2(0.5 - 0.0625)] 0.5625

5.136 = [Dp - 8.625]

Dp = 13.761 in.

Therefore, the minimum required reinforcement pad diameter is 13.761 in.

Confirm that this diameter does not extend beyond the outer limit of the permitted reinforcement zone in the shell, 2d.

2d = 2 x 7.75 = 15.5 in.

Therefore, Dp = 13.761 in. is acceptable.

F. Flange Rating

ASME B16.5, Pipe Flanges and Flanged Fittings, provides steel flange dimensional details for standard pipe sizes through NPS 24. ASME B16.5 flanges are acceptable for most pressure vessel nozzles and for shell flanges when the vessel diameter corresponds to a standard pipe size. Specification of an ASME B16.5 flange involves selection of the correct material and flange "Class." The paragraphs that follow discuss the flange specification process in general terms.

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Flange material specifications are listed in Table 1A in ASME B16.5, a portion of which is excerpted as Figure 4.17. The material specifications are grouped within specific Material Group Numbers. For example, if the pressure vessel is fabricated from carbon steel, ASTM A105 is an appropriate flange material specification in most applications. ASTM A105 material is in Material Group No. 1.1. Refer to ASME B16.5 for additional acceptable material specifications and corresponding Material Group Numbers.

Material Groups Product Forms

Material Group

Number

Nominal Designation

Steel

Forgings

Castings

Plates

Spec. No. Grade Spec. No. Grade Spec. No. Grade

1.1 Carbon A105 -- A216 WCB A515 70 A350 LF2 -- -- A516 70 C-Mn-Si -- -- -- -- A537 Cl.1

1.2 Carbon -- -- A216 WCC -- -- -- -- A352 LCC -- -- 2 ½ Ni -- -- A352 LC2 A203 B 3 ½ Ni A350 LF3 A352 LC3 A203 E

ASME B16.5, Table 1a, Material Specification List (Excerpt)

Figure 4.17

Table 2 of ASME B16.5 is used to select the appropriate flange Class for the specified design conditions and Material Group Number. ASME B16.5 has seven Classes: 150, 300, 400, 600, 900, 1,500, and 2,500. Each Class specifies the design pressure and temperature combinations that are acceptable for a flange that has that designation. As the number of the Class increases, the strength of the flange increases for a given Material Group. Figure 4.18 is an excerpt from Table 2 and shows the temperature and pressure ratings for three carbon steel Material Groups.

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Material Group

No. 1.1 1.2 1.3

Classes 150 300 400 150 300 400 150 300 400 Temp., °F -20 to 100 285 740 990 290 750 1000 265 695 925

200 260 675 900 260 750 1000 250 655 875 300 230 655 875 230 730 970 230 640 850 400 200 635 845 200 705 940 200 620 825 500 170 600 800 170 665 885 170 585 775 600 140 550 730 140 605 805 140 534 710 650 125 535 715 125 590 785 125 525 695 700 110 535 710 110 570 755 110 520 690 750 95 505 670 95 505 670 95 475 630 800 80 410 550 80 410 550 80 390 520 850 65 270 355 65 270 355 65 270 355 900 50 170 230 50 170 230 50 170 230 950 35 105 140 35 105 140 35 105 140

1000 20 50 70 20 50 70 20 50 70

ASME B16.5, Table 2, Pressure-Temperature Ratings (Excerpt) Figure 4.18

Specification of the size, material, and Class completes most of the selection requirements for flanges. Flange type and gasket material must also be specified. Discussion of these factors is beyond the scope of this course.

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Sample Problem 4 – Determine Required Flange Rating

For the pressure vessel described below, use the following procedure to determine the required flange rating (or Class) in accordance with ASME B16.5.

Pressure Vessel Material Specifications:

Shell and Heads: SA-516 Gr.70

Flanges: SA-105

Design Temperature: 700°F

Design Pressure: 275 psig

1. Identify the material specification of the flange.

SA-105

2. Go to Figure 4.17 (Table 1A of ASME B16.5) and determine the Material Group No. for the selected material specification.

Group 1.1

3. Go to Figure 4.18 (Table 2 of ASME B16.5) with the design temperature and Material Group No. determined in Step 3.

• The intersection of design temperature with Material Group No. is the maximum allowable design pressure for the flange Class.

• Table 2 of ASME B16.5 contains design information for all seven possible flange Classes (i.e., 150, 300, 400, 600, 900, 1500, 2500).

• Select the lowest Class whose maximum allowable design pressure is equal to or greater than the required design pressure.

At 700°F, for Group 1.1 flange material, the Lowest Class that will accommodate a design pressure of 275 psig is Class 300. At 700°F a Class 300 flange of Material Group 1.1 can have a design pressure up to 535 psig.

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G. Flange Design

For some pressure vessel applications, it is advantageous to have one or more flanged joints in the vessel shell to facilitate entry, removal, and/or replacement of internal components (e.g., cartridge trays). In most applications such as these, the shell diameter is of a size that standard-sized flanges designed in accordance with either ASME B16.5 or ASME B16.47 may be used. Mechanical design calculations for these standard flanges are not necessary.

Flanges must be custom-designed in situations where standard-sized flanges are not appropriate. The most common application for custom-designed flanges is for the girth flanges of shell-and-tube heat exchangers. All custom-designed flanges must meet the requirements of Appendix 2 of Division 1. The Appendix 2 design procedure is complicated and is best done using a computer program. The following paragraphs briefly describe:

• The main steps in the ASME flange design procedure.

• The parameters that affect flange design and in-service performance.

1.0 ASME Flange Design Procedure

The ASME flange design procedure consists of determining the:

• Bolting requirements.

• Flange design loads and moments.

• Stresses in the flange ring and hub.

The first step is usually to determine the required number and size of bolts. Bolting requirements are determined by calculating the loads on the bolts for two separate cases:

• Normal operation

• Initial flange boltup

The bolt load during normal operation, Wm1, is based on the design conditions. The bolt load during initial flange boltup, Wm2, is based on the load (or stress) necessary to seat the gasket and form a tight seal.

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The bolt area that is required for each of these loads is then calculated by dividing each bolt load by the bolt allowable stress at design temperature and room temperature, respectively. Either the operating case or the gasket seating case may result in the minimum required bolt area, Am; therefore, both cases must be checked. Since bolts come in standard sizes, and there are limitations on the spacing between bolts, the actual bolt area, Ab, is usually greater than the required bolt area.

The next step is to determine the design loads and moments on the flange. These loads include the:

• Design bolt load on the flange (W).

• Hydrostatic pressure loads that act on the flange (HD and HT).

• Gasket sealing force (HG).

These loads do not all act at the same location on the flange, therefore, effective moment arms (hD, hT, and hG) are calculated based on the locations of the bolts and gasket, and on the flange geometry (See Figure 4.19). The appropriate loads are then multiplied by the effective moment arms to determine flange design moments for the operating and gasket seating cases.

t h

W

C

A

G BHD

HT

hT

hG

HG

hD

g0

g1

Gasket

FlangeRing

Flange Hub Flange Loads and Moment Arms

Figure 4.19

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The stresses in the flange ring and hub are then calculated using stress factors specified in the ASME Code (based on flange geometry), the applied moments, and the flange geometry. The stresses are calculated for both the operating case and gasket seating cases and are then compared to the appropriate Code allowable stresses.

All flange stresses will be lower than the appropriate allowable stresses if the flange is designed properly. It may be necessary to increase the flange thickness, change the hub dimensions, or make other changes to the flange design parameters to keep flange stresses within their allowable limits. The computer programs that suppliers use for flange design use iterative calculation procedures to optimize flange design. In this sense, the goal of optimization, from the supplier’s viewpoint, is to design a “least weight” (i.e., lowest cost) flange that will satisfy the design requirements.

2.0 Parameters That Affect Flange Design and In-Service Performance

The following parameters affect flange design and in-service performance:

• ASME Code m and y parameters.

• Specified gasket widths.

• Flange facing and nubbin width, w.

• Bolt size, number, and spacing.

The gasket factor, m, determines the amount of force required to keep the gasketed joint tight. The minimum design seating stress, y, determines how much gasket stress is required to initially seat or deform the gasket. Both parameters are used in the flange design calculations.

The ASME Code specifies m and y based on gasket type in its Table 2-5.1 (excerpted in Figure 4.20). Higher values of m and y typically indicate that a gasket is harder to seal or seat. While this is a consideration in gasket selection, gasket type and material are usually selected based on historical service experience and the corrosion resistance of the gasket material in the process environment.

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Heat exchanger flanges sometimes have leakage problems during operation. When this occurs, there is often the tendency to change the gasket to a different type that has provided leak-free performance in other applications. This problem-solving method should always be approached with caution because the flanges were designed for a specific gasket type with its associated m and y values. Therefore, the existing bolting may either impose too high a load on the gasket (and possibly crush it) or the new gasket may require a higher load to seat it (which might not be possible with the existing bolting).

Gasket Type and Material

Gasket

Factor, m

Min. Design Seating Stress y,

psi

Facing Sketch and Column in ASME

Table 2-5.2 (Figure 4.21)

Flat metal, jacketed asbestos filled:

Soft aluminum Soft copper or brass Iron or soft steel Monel 4-6% chrome Stainless steels and nickel-base alloys

3.25 3.50 3.75 3.50 3.75 3.75

5,500 6,500 7,600 8,000 9,000 9,000

(1a), (1b), (1c), (1d); (2); Column II

Solid flat metal:

Soft aluminum Soft copper or brass Iron or soft steel Monel or 4-6% chrome Stainless steels and nickel-base alloys

4.00 4.75 5.50 6.00 6.50

8,800 13,000 18,000 21,800 26,000

(1a), (1b), (1c), (1d); (2), (3), (4), (5); Column I

ASME Code m and y Factors Figure 4.20

The TEMA standard for shell-and-tube heat exchangers specifies a minimum required width for the peripheral ring gaskets at external joints (3/8 in. or ½ in. depending on shell size) and for pass partition gaskets (¼ in. or 3/8 in. depending on shell size). These minimum gasket widths are typically used over a wide range of service conditions.

The gasket widths referred to in TEMA are actual minimum widths, N. In addition to N, two other gasket widths are referred to in the ASME Code: the basic seating width, bo, and the effective seating width, b. The effective seating width is a function of the basic seating width, and the basic seating width is a function of the actual width and the type of flange face. See Table 2-5.2 in the ASME

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Code (excerpted in Figure 4.21). In general, wider gaskets provide better sealing, but a wider gasket also requires a larger bolt load (i.e., more bolt area) to seat and seal the gasket. The required flange thickness increases as the bolting area increases.

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Basic Gasket Seating Width bo Facing Sketch (Exaggerated)

Column I Column II

(1a) N N

(1b) N

N

2N

2N

(1c) N

wT

w ≤ N

(1d) N

wT

w ≤ N

++

max4

Nw;

2Tw

++

max4

Nw;

2Tw

HG

G hG

bO.D. Contact Face

For bo > ¼ in. For bo < ¼ in.

HG

G hG

GasketFace

CL

ASME Code Gasket Widths (Table 2 -5.2 excerpt) Figure 4.21

The effective seating width, b, is also a function of the flange facing type and the nubbin width, w, for flat metal gaskets. Table 2 -5.1 in the Code (excerpted in Figure 4.22) indicates which facing sketch is applicable for a given gasket type and material.

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Gasket Materials and Contact Facings

Gasket Factors m for Operating Conditions and Minimum Design Seating Stress y

Gasket Material Gasket Factor

m

Min. Design Seating Stress y,

psi

Sketches Facing Sketch and Column in Table 2-5.2

Flat metal, jacketed asbestos filled: Soft aluminum Soft copper or brass Iron or soft steel Monel 4% - 6% chrome Stainless steels and nickel-base alloys

3.25 3.50 3.75 3.50 3.75 3.75

5500 6500 7600 8000 9000 9000

(1a), (1b), (1c),2, (1d)2, (2)2, Column

II

Gasket Materials and Contact Facings (Table 2-5.2 Excerpt)

Figure 4.22

The equations for determining b are based on w, N, and the type of flange facing. Note that b is used in the Code equations to determine the bolt load required for sealing the gasket during operation, Wm1, and the bolt load required for seating the gasket initially, Wm2. Once a gasket type, material, width, and facing are selected, the required bolting area can be determined.

• The bolt size, number, and spacing that are used to clamp the flanges together are interrelated parameters that affect their overall design.

• The number of bolts multiplied by the bolt root area of a single bolt must be greater than the minimum required bolt area, Am.

• The bolts must be far enough away from the shell or hub of the flange, and be far enough apart circumferentially, so that there is adequate clearance to permit access for a wrench.

• There must be adequate distance to other flange or vessel surfaces to ensure adequate clearance for standard wrenches.

It may appear that maintaining these minimum bolt dimensions can be easily achieved if a few large bolts are used. However, the bolts should also be spaced as close together as practical for several reasons.

• Having fewer bolts increases the bolt load moment arms. Larger moment arms increase the bending moments for which the flange must be designed and thus increase the required flange thickness.

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• TEMA requires that the flange design moment be increased if the bolts are widely spaced. This results in a thicker flange.

• Excessive bolt spacing could make the flange more prone to leakage. The portions of the gasket located between the bolts might not be compressed sufficiently to maintain a tight seal.

H. Maximum Allowable Working Pressure (MAWP)

The MAWP of a pressure vessel is the maximum permissible gauge pressure in the vessel. It is determined as the lowest MAWP of all the vessel's components based on their actual supplied thicknesses. The MAWP is specified at the top of the vessel when the vessel is in its operating position. The MAWP is also specified at a "designated temperature" (i.e., the design temperature) that is coincident with the MAWP. The material thicknesses used in these calculations do not include any excess thickness that was added for corrosion allowance or to absorb loadings other than pressure.

The MAWP may be used later if a change in operation is being considered that requires a more severe design pressure and/or temperature than what were originally specified. The MAWP shows whether the same pressure vessel may be used at the new design conditions.

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V. Other Design Considerations

A. Vessel Support

The type of support that is used for a pressure vessel depends primarily on the vessel’s size and orientation.

Shown in Figure 2.1, a saddle support spreads the weight load of a horizontal drum over a large area of the shell. This prevents excessive local stress in the shell at the support points. The size and design details used for the saddle depend on the diameter and thickness of the drum and the imposed load.

As shown in Figure 2.2, small vertical drums are typically supported on legs that are welded to the lower portion of the shell. Support legs are also typically used for spherical pressurized storage vessels (See Figure 2.5). The support legs for small vertical drums and spherical pressurized storage vessels may be made from structural steel columns or pipe sections, whichever provides a more efficient design. Cross bracing between the legs (See Figure 2.5) is typically used to help absorb wind or earthquake loads.

Lugs may also be used to support vertical pressure vessels. As shown in Figure 5.1, the lugs are typically bolted to horizontal structural members. It is common for a reinforcement pad to be first welded to the vessel shell, and then the lugs welded to it.

A support skirt (See Figures 2.3 and 2.4) is a cylindrical shell section that is welded either to the lower portion of the vessel shell or to the bottom head. Support skirts are commonly used for tall towers.

B. Local Loads

It is common for external loads to be applied to nozzles or lugs that are attached to pressure vessel shells or heads. External loads cause local stresses that are in addition to those caused by pressure, weight, and wind loads. External loads may be caused by the following:

• Piping system weight, wind, and thermal expansion loads that are applied at vessel nozzles.

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• Loads from platforms, internal or external piping, internal components, or equipment items supported from a vessel shell by lugs or clips attached to the shell.

• Loads at vessel supports, such as columns or lugs.

Vertical Vessel on Lug Supports Figure 5.1

The total stress in the vessel shell, including that caused by locally applied loads, must be kept to within allowable limits. Division 1 does not contain detailed procedures for evaluating these local loads. Other industry practices (e.g., Welding Research Council Bulletins 107 and 297) and Division 2 are commonly used to evaluate local loads.

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C. Vessel Internals

1.0 Types of Internals

There are many different types of vessel internals used to perform various process functions. The following highlights several (but not all) of these types:

• Trays. Located at various elevations along the length of a tower. Provide liquid/vapor flow distribution and separation along the length. Various tray types are available to suit specific process needs.

• Inlet distributor. Installed as an internal extension to the inlet nozzle. Used to direct the inlet flow stream and properly distribute it within the vessel.

• Anti-vortex baffle. Installed at vessel outlet to prevent the formation of flow vortices at the exit from the vessel.

• Catalyst bed grid and support beams. An open steel gridwork may be used to support one or more intermediate catalyst beds installed inside fixed bed reactors. The gridwork is typically covered by wire mesh screen to prevent the solid catalyst from passing through the grid, while the gridwork permits process flow. Supplementary beams are typically used to support the grid from the vessel shell.

• Outlet collector. Typically placed at the outlet of fixed bed reactors. Designed to allow process flow while preventing catalyst from passing into the downstream system.

• Flow distribution grid. In fluidized solids processes (e.g., FCCU, Fluid Coker, etc.), a flow distribution grid is used to direct and distribute the fluidization media that is needed to keep the solids (i.e., catalyst or coke) in a fluidized state inside a vessel.

• Cyclone and plenum chamber system. In fluidized solids processes (e.g., FCCU, Fluid Coker, etc.), a cyclone and plenum chamber system separates entrained catalyst from process vapor before the vapor exits the vessel through the overhead line.

ASME Code design requirements only apply to the external, pressure-containing “envelope” of the vessel (i.e., shell, heads, nozzles, etc.) and not to items contained inside it. The only exceptions to this are:

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• Loads that are applied from the internals to pressure-containing parts must be considered in the vessel design.

• All welding to pressure-containing parts must meet ASME Code requirements.

The end-user, vessel vendor, internals supplier, prime contractor, and/or a combination of these entities must develop the detailed design requirements for all vessel internals.

2.0 Treatment of Corrosion Allowance.

Removable pressure vessel internals that are subject to corrosion should typically have a corrosion allowance equal to that of the shell. In this way, the design of removable internals considers only half of the expected total corrosion. The rationale for this approach is that removable internals that are designed for only the expected total corrosion will cost less initially and can easily be replaced later, based on the actual corrosion that occurs.

Most pressure vessel internals can corrode on both sides. From a strength-design viewpoint, corrosion from both sides should be considered with regard to non-removable internals. Non-removable internals, and those that are major load-bearing members (e.g., catalyst bed supports), must typically have a total corrosion allowance that is equal to twice that of the shell.

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VI. Fabrication

A. Acceptable Welding Details

All pressure vessel welds, including the welds that attach heads, nozzles, small fittings, and nonpressure components to a shell, must conform to ASME Code requirements. Details that are used for the primary circumferential and longitudinal welds were discussed earlier in conjunction with weld joint categories.

The ASME Code specifies weld detail requirements for vessel fabrication (e.g., type and size of weld, weld locations, etc.). It also specifies welder and welding procedure qualification requirements. The paragraphs that follow highlight several of the ASME Code requirements. Refer to the ASME Code for further information related to these and other weld details.

1.0 Thickness Transitions

The thickness of a pressure vessel head sometimes differs from the thickness of the shell it is attached to (e.g., when a hemispherical head is attached to a cylindrical shell). The transition between the component thicknesses must be made in a taper to avoid excessive local stress. Head-to-shell thickness transitions are illustrated in Figure 6.1.

2.0 Intermediate Heads

An intermediate head is attached to the inside of a cylindrical shell when it is needed to separate two sections of the vessel. The butt weld between shell sections also attaches to the head, and a fillet weld is also located between the head and shell. The ASME Code permits elimination of the fillet weld if there is no access and if the service is noncorrosive. However, the fillet weld should generally be used for all refinery applications to avoid the potential for accelerated corrosion due to process fluid getting between the head and shell. The attachment of an intermediate head to a cylindrical shell is illustrated in Figure 6.1.

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FilletWeld

Butt Weld

Intermediate Head Attachment

th

y

l

ts

Thin

ner

part

TangentLine

th

y

l

ts

Thin

ner

part

th

y

l

ts

Thin

ner

part

TangentLine

th

y

l

ts

Thin

ner

part

Typical Head-to-Shell Transitions Figure 6.1

3.0 Openings

Fabrication details for various types of openings are specified. These include unreinforced nozzles (e.g., a nozzle neck welded directly to the vessel shell or head), a nozzle with a reinforcing pad added, and a self-reinforced nozzle (i.e., where extra thickness is

145

provided in the nozzle neck to provide the necessary reinforcement). These were illustrated in Figure 4.15.

In some cases, a nozzle neck that has a weld-end may be attached to a pipe that is thinner. This attachment between components of different thicknesses could occur if extra thickness was included in the nozzle neck for reinforcement or if the pipe and nozzle materials and/or allowable stresses differ. In such cases, the nozzle neck must be tapered to the pipe thickness. Tapers are also used to join shell sections that are of different thicknesses. Shell thickness and nozzle thickness tapers are illustrated in Figures 6.2 and 6.3, respectively.

l

l

y

CL

CL

CLIn all cases, l shall notbe less than 3y.

Typical Shell Transitions

Figure 6.2

Nozzle Neck Attachment to Thinner Pipe Figure 6.3

4.0 Stiffener Rings

Stiffener rings may be attached to the vessel shell by continuous, intermittent, or a combination of continuous and intermittent welds.

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Intermittent welds must be placed on both sides of the stiffener and may be either staggered or in-line. The ASME Code specifies acceptable spacing, size, and length of the welds. Stiffener ring attachment weld options are illustrated in Figure 6.4.

In-LineIntermittent Weld

StaggeredIntermittent Weld

Continuous Fillet Weld OnOne Side, Intermittent Weld

On Other Side

Stiffener Ring Attachment Figure 6.4

B. Postweld Heat Treatment Requirements

Welding heat changes the crystal structure and grain size of the weld heat affected zone (HAZ). Postweld heat treatment (PWHT) may be necessary to restore the material structure to the required properties. The need for PWHT for these metallurgical reasons depends on the materials involved and the service conditions that they are exposed to. PWHT requirements for these metallurgical or process reasons are not included in the ASME Code. They must be specified by the user based on the service and materials involved.

As the weld metal and HAZ cool from the very high welding temperatures, the thermal contraction that occurs in the locally heated area is resisted by the cooler base metal that surrounds it. This resistance results in residual stresses that remain in the structure. For thicker plates, these residual stresses must be removed by PWHT. PWHT requirements based on stress relief considerations are contained in the ASME Code, Section VIII.

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The ASME Code contains the temperature and hold time requirements when PWHT is needed for stress relief considerations. These ASME Code PWHT requirements are based on material type and thickness, as specified in Paragraph UCS-56 for carbon and low-alloy steels. The ASME Code specifies the minimum PWHT temperature and the minimum holding time at temperature based on the material P-No. and thickness. Acceptable PWHT procedures are also specified to ensure that adequate stress relief will occur. Heatup and cooldown rates must be controlled within specified limits in order to avoid excessive local thermal stresses during PWHT.

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VII. Inspection and Testing

A. Inspection

Overall inspection of completed pressure vessels includes an examination of the following:

• Base material specification and quality

• Welds

• Dimensional requirements

• Equipment documentation

The most common defects for which welds are examined are as follows:

• Poor weld shape due to part misalignment.

• Cracks in welds or HAZ of the base metal.

• Pinholes on the weld surface.

• Slag inclusions or porosity in the form of voids.

• Incomplete fusion between weld beads or between the weld and the base metal.

• Lack of penetration or an insufficient extent of penetration of the weld metal into the joints.

• Undercut, an intermittent or continuous groove that is located adjacent to the weld and that is left unfilled by weld metal.

Several of these common weld defects are illustrated in Figure 7.1.

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Undercut

Incomplete Penetration

Lack of Fusion

Between Weld Bead and Base Metal Between Adjacent Passes

Incomplete Filling at Root on One Side Only Incomplete Filling at Root

Internal Undercut

External Undercut

Typical Weld Defects Figure 7.1

The presence of defects reduces the strength of the weld below that required by the design calculations, reduces the overall strength of the fabrication, and increases the risk of failure. Weld inspection must be performed in a manner that will detect unacceptable defects while not damaging the vessel material. This type of inspection is called nondestructive examination, or NDE.

The five primary weld NDE methods are as follows:

• Radiographic examination (RT)

• Visual Inspection (VT)

• Liquid penetrant examination (PT)

• Magnetic particle test (MT)

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• Ultrasonic examination (UT)

The choice of which weld examination method or methods to use depends on the weld quality required of the joint, the position of the weld, the material to be joined, and the particular defects that are most likely to occur. These weld NDE methods are briefly discussed in the paragraphs that follow. Figure 7.2 summarizes the types of NDE, the defects typically found by each, and the advantages and limitations of each process.

NDE TYPE DEFECTS DETECTED

ADVANTAGES LIMITATIONS

Radiographic Gas pockets, slag inclusions, incomplete penetration, cracks

Produces permanent record. Detects small flaws. Most effective for butt-welded joints.

Expensive. Not practical for complex shapes.

Visual Porosity holes, slag inclusions, weld undercuts, overlapping

Helps pinpoint areas for additional NDE.

Can only detect what is clearly visible.

Liquid Penetrant Weld surface-type defects: cracks, seams, porosity, folds, pits, inclusions, shrinkage

Used for ferrous and nonferrous materials. Simple and less expensive than RT, MT, or UT.

Can only detect surface imperfections.

Magnetic Particle Cracks, porosity, lack of fusion

Flaws up to ¼ in. beneath surface can be detected.

Cannot be used on nonferrous materials.

Ultrasonic Subsurface flaws: laminations, slag inclusions

Can be used for thick plates, welds, castings, forgings. May be used for welds where RT not practical.

Equipment must be constantly calibrated.

Summary of NDE types Figure 7.2

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1.0 Radiographic Examination (RT)

The most important NDE method is radiographic examination. In radiographic examination, a ray is emitted from a controllable source, penetrates a test specimen, and leaves an image on a strip of film that is mounted behind the test specimen. This is illustrated in Figure 7.3.

Test Specimen

Film

X-Ray

X-Ray Tube

Typical RT Setup

Figure 7.3

2.0 Visual Inspection (VT)

A thorough visual inspection is usually satisfactory for minor structural welds. All weld surfaces that will be examined by more extensive means are first subject to VT. VT provides an overall impression of weld quality and helps to locate areas where additional NDE should be performed.

3.0 Liquid Penetrant Examination (PT)

A liquid penetrant examination involves applying a penetrant which contains a fluorescent or visible dye to mark potential defect areas. The liquid penetrates into defects by capillary action. Then, by using

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a developing procedure, the liquid bleeds out through a capillary action at surface flaws and makes them visible.

4.0 Magnetic Particle Test (MT)

MT examination is based on the magnetic lines of flux (or force lines) that can be generated within a test piece. These force lines are parallel if no defects are present. If there is a defect, a small break in the force lines appears at the defect location. In MT examination, iron powder is applied to the surface and then the test piece is magnetized. If there are no defects, the iron powder is aligned in straight lines along the North-South magnetic flux lines. If there is a defect, the iron powder alignment is disturbed and flows around the defect.

5.0 Ultrasonic Examination (UT)

In UT examination, sound waves are generated by a power source and applied to the test piece through a transducer. Figure 7.4 shows a pulse echo ultrasonic examination system. The sound waves pass through the test piece and are reflected back to the transducer either from the far side of the test piece or from a flaw that is located at an intermediate position within the test piece. By careful calibration, the UT operator knows if a flaw has been detected and knows its location and its size.

B. Pressure Testing

All pressure vessels that are designed to ASME Code requirements must be pressure tested after fabrication and inspection to demonstrate their structural integrity before they are placed into operation. The pressure test is made at a pressure that is higher than the design pressure. This excess pressure provides a safety margin since the vessel component stress levels during the test will be higher than those that will occur during operation. The objective of the pressure test is to bring the vessel to a high enough internal pressure, under controlled conditions, to demonstrate its mechanical integrity. Successful completion of the pressure test signifies that the vessel is acceptable for operation.

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A

Transducer

Cable

Flaw

Couplant

B

Test Specimen

Read Out

Base Line

Cathode Ray Tube (CRT)

CB

A

Input-OutputGenerator

C

Pulse Echo UT System

Figure 7.4

Pressure tests are typically made using water as the test medium because of the relative safety of water compared to a pneumatic test. The ASME Code permits a pneumatic pressure test as an alternative to a hydrostatic test under certain circumstances. However, a pneumatic test should only be considered on an exception basis due to the increased safety risks involved.

Since the hydrostatic test will almost always be used, only the hydrostatic test will be discussed here. Refer to the ASME Code for pneumatic test requirements.

The standard hydrotest pressure at the top of the vessel is calculated as follows:

PT = 1.5P (Ratio)

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Where:

PT = Hydrotest pressure at the top of the vessel, psig

P = Vessel MAWP (use vessel design pressure if the MAWP was not determined), psig

Ratio = The lowest ratio of the allowable stress at the test temperature to that at the design temperature for the vessel materials used.

The following points must also be considered:

• Hydrotest pressures must be calculated for the shop test with the vessel in the horizontal position, for the field test with the vessel in the final position and with uncorroded component thicknesses, and for the field test with the vessel in the final position and with corroded component thicknesses.

• The calculated shop hydrotest pressure cannot exceed the test pressure of the flanged connections.

• During the pressure test, the stress at any section of the vessel cannot exceed 90% of the material minimum specified yield strength (MSYS), based on use of the design weld joint efficiency (E).

• Vessels also must typically be designed to permit a hydrotest in the field at a wind velocity that is typically 25-35% of the design wind velocity for the site.

During a field hydrotest, water at a specific gravity of 1.0 is used, and the vessel is filled to the top. The larger specific gravity and fill height of hydrotest water results in a higher weight and hydrostatic head load than occurs during normal operation. Therefore, thicker plates are sometimes required for lower sections of a tall tower than would be required for the operational loads.

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VIII. Summary

This course provided an overview of pressure vessel mechanical design requirements. It summarized the main components of pressure vessels and discussed the scope of the ASME Code Section VIII, structure of Division 1, materials of construction, design requirements and considerations, fabrication, inspection and testing. Participants now have a good overall understanding of pressure vessel mechanical design requirements, are prepared to use this knowledge in their jobs, and have sufficient prerequisite information to take more detailed pressure vessel courses.

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