Specifying Design Requirements for Pressure Vessels

Engineering Encyclopedia Saudi Aramco DeskTop Standards

SPECIFYING DESIGN REQUIREMENTS FOR PRESSURE VESSELS

Note: The source of the technical material in this volume is the Professional Engineering Development Program (PEDP) of Engineering Services. Warning: The material contained in this document was developed for Saudi Aramco and is intended for the exclusive use of Saudi Aramco’s employees. Any material contained in this document which is not already in the public domain may not be copied, reproduced, sold, given, or disclosed to third parties, or otherwise used in whole, or in part, without the written permission of the Vice President, Engineering Services, Saudi Aramco.

Chapter : Mechanical File Reference: MEX-202.03

For additional information on this subject, contact PEDD Coordinator on 874-6556

Engineering Encyclopedia

Design of Pressure Vessels Specifying Design Requirements for Pressure Vessels

CONTENT

PAGE

INTRODUCTION............................................................................................................ 6 USE OF SAUDI ARAMCO DOCUMENTS AND THE ASME CODE IN PRESSURE VESSEL DESIGN .......................................... 7 Structure and Scope of the ASME Code, Section VIII, Division 1, Design By Established Rules......................................... 9 ASME Code Structure............................................................................... 9 ASME Code Scope ................................................................................. 10 ASME Code, Section VIII, Division 2, Design By Analysis................................. 11 Saudi Aramco Pressure Vessel Design Data Sheets ........................................ 12 Content of Form 2682 for Division 1 Pressure Vessels........................... 14 Content of Form 2683 for Division 2 Pressure Vessels........................... 18 EVALUATING THE ACCEPTABILITY OF CONTRACTOR-SPECIFIED DESIGN CONDITIONS AND LOADINGS...................... 19 Pressure ............................................................................................................ 20 Operating Pressure................................................................................. 20 Design Pressure ..................................................................................... 21 Temperature ...................................................................................................... 24 Operating Temperature........................................................................... 24 Design Temperature ............................................................................... 26 Minimum Design Metal Temperature ...................................................... 26 Other Loadings .................................................................................................. 27 Weight..................................................................................................... 30 Wind........................................................................................................ 32 Hydrotest ................................................................................................ 36 External Piping........................................................................................ 37 Internal Components............................................................................... 38

Saudi Aramco DeskTop Standards

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Earthquake Loadings .............................................................................. 39 Service .............................................................................................................. 40 Wet, Sour................................................................................................ 40 Lethal ...................................................................................................... 41 EVALUATING THE ACCEPTABILITY OF CONTRACTOR-SPECIFIED PRESSURE VESSEL COMPONENT THICKNESS DESIGN CRITERIA..................... 42 Weld Joint Efficiency ......................................................................................... 42 Corrosion Allowance.......................................................................................... 44 EVALUATING CONTRACTOR-SPECIFIED DESIGN CALCULATIONS FOR PRESSURE VESSEL COMPONENTS..................... 45 Design for Internal Pressure .............................................................................. 45 Shells ...................................................................................................... 46 Sample Problem 1 - Cylindrical Shell Thickness Calculation .................. 48 Heads ..................................................................................................... 51 Sample Problem 2 - Head Thickness Calculation ................................... 55 Conical Sections ..................................................................................... 56 Sample Problem 3 - Conical Section Thickness Calculation................... 56 Design for External Pressure and Compressive Stresses ................................. 58 Shells ...................................................................................................... 60 Heads ..................................................................................................... 60 Conical Sections ..................................................................................... 60 Sample Problem 4 - External Pressure Calculation ................................ 61 Flat Covers ........................................................................................................ 67 Quick-Opening Closures ................................................................................... 67 Nozzle Reinforcement for Design Pressure....................................................... 70 Reinforcement of Pressure Vessel Openings ......................................... 70 Additional Nozzle Reinforcement ............................................................ 74


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Small Connections .................................................................................. 75 Sample Problem 5 - Nozzle Reinforcement ............................................ 76 Nozzle Flange Rating ........................................................................................ 79 Maximum Allowable Working Pressure (MAWP)............................................... 83 Stresses From Local Loads Applied to Nozzle and Attachments ...................... 85 Allowable Stress Bases .......................................................................... 86 Calculation Procedures........................................................................... 87 Shell and Nozzle Attachment Parameters .............................................. 90 Sample Problem 6 Evaluation of Stresses from Local Loads Applied to Nozzles and Attachments ................................................................... 93 EVALUATING THE CONTRACTOR-SPECIFIED DESIGN OF PRESSURE VESSEL SUPPORTS ......................................................... 95 Vertical Vessel................................................................................................... 97 Column Supports .................................................................................... 98 Sample Problem 7 - Design of Column Supports ................................. 100 Skirt Supports ....................................................................................... 108 Horizontal Vessel Saddle Supports ................................................................. 111 Design of Horizontal Cylindrical Vessels on Saddle Supports .............. 114 COMPLETING SAFETY INSTRUCTION SHEETS FOR PRESSURE VESSELS...... 117 Purpose and Use of the Safety Instruction Sheet in Saudi Aramco................. 119 Information Covered ........................................................................................ 119 Where to Find Other Information ..................................................................... 121 SUMMARY................................................................................................................. 123 Cylindrical or Spherical Shells Under External Pressure ...................... 136 Heads and Conical Sections Under External Pressure......................... 141 Allowable Compressive Stress of Cylindrical Shells ............................. 147 GLOSSARY ............................................................................................................... 171


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List of Figures Figure 1: Design Pressure.......................................................................................... 23 Figure 2: Temperature Zones in Tall Vessels............................................................. 25 Figure 3: Vertical Reactor with Additional Loads ........................................................ 29 Figure 4: Tower Layout for Determining Effective Diameter for Wind Calculations ... 33 Figure 5: Vortex Shedding.......................................................................................... 35 Figure 6: Sample Problem 1....................................................................................... 49 Figure 7: Typical Formed Closure Heads................................................................... 52 Figure 8: Thickness Transition Between Hemispherical Head and Shell ................... 54 Figure 9: Stiffener Rings on Pressure Vessel Cylinders............................................. 59 Figure 10: Sample Problem 4..................................................................................... 61 Figure 11: Factor A..................................................................................................... 63 Figure 12: Figure CS-1............................................................................................... 63 Figure 13: Cross-Sectional View of Nozzle Opening.................................................. 71 Figure 14: Typical Nozzle Design Configurations....................................................... 73 Figure 15: Sample Problem 5..................................................................................... 77 Figure 16: ASME B16.5, Table 1a, Material Specification List (Excerpt).................... 80 Figure 17: ASME/ANSI B16.5, Class 150, Pressure-Temperature Ratings (Excerpt) 81 Figure 18: Sample Problem 6..................................................................................... 93 Figure 19: Types of Pressure Vessel Supports .......................................................... 95 Figure 20: Column Support Loads.............................................................................. 99 Figure 21: Sample Problem 7................................................................................... 101 Figure 22: Types of Support Skirts ........................................................................... 109


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Figure 23: Horizontal Vessel on Saddle Supports .................................................... 111 Figure 24: Stiffener Rings at Saddle Supports ......................................................... 113 Figure 25: Pressure Vessel Safety Instruction Sheet, Form 2694............................ 118 Figure 26: Weld Joint Categories ............................................................................. 128 Figure 27: Types of Welded Joints ........................................................................... 130 Figure 28: Maximum Weld Joint Efficiency............................................................... 131 Figure 29: Nozzle Loads Applied to a Spherical Shell.............................................. 154 Figure 30: Vessel on Column Supports.................................................................... 158 Figure 31: Vessel Column Configurations and Moments of Inertia .......................... 160 Figure 32: Allowable Column Compressive Stress................................................... 161 Figure 33: Types of Support Skirts and Skirt-to-Head Welds ................................... 168 Figure 34: Vessel Safety Instruction Sheet Form 2694 With Number Key ............... 170


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INTRODUCTION MEX 202.02 discussed materials selection requirements for pressure vessels. Now that the materials have been selected, the mechanical design of individual pressure vessel components can begin. MEX 202.03, Specifying Design Requirements for Pressure Vessels, describes the use of Saudi Aramco design requirements and the ASME Code in the mechanical design of pressure vessels. This description includes determination of design conditions and loadings, and the design criteria required for calculating pressure vessel component thicknesses. The pressure vessel component design calculations themselves, as well as the design calculations required for pressure vessel supports, are also discussed. Saudi Aramco pressure vessel engineers do not perform the detailed mechanical design of pressure vessels. Pressure vessel mechanical design is done by contractors and pressure vessel manufacturers who are employed by Saudi Aramco. Therefore, the role of the Saudi Aramco pressure vessel engineers will typically be to evaluate the designs that are proposed by others. However, in order to perform this design evaluation role effectively, the Saudi Aramco pressure vessel engineer must know the design requirements that the contractors and vessel manufacturers must meet.


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USE OF SAUDI ARAMCO DOCUMENTS AND THE ASME CODE IN PRESSURE VESSEL DESIGN The scope and general use of SAES-D-001, Design Criteria for Pressure Vessels, and 32-SAMSS-004, Pressure Vessels, were discussed in MEX 202.02. These Saudi Aramco documents refer to the ASME Boiler and Pressure Vessel Code Section VIII as the basic industry standard that provides the design requirements for pressure vessels that are used by Saudi Aramco. The Saudi Aramco documents supplement the ASME Code as necessary, based on specific Saudi Aramco requirements. Specific design requirements that are contained within these Saudi Aramco documents will be discussed as appropriate in this module. The ASME Code, Section VIII, is divided into two main sections: Division 1 and Division 2. Division 1 is used most often by industry. This course concentrates on Division 1. However, it is necessary to understand Division 2 in general, how it differs from Division 1, and where its use might be appropriate for Saudi Aramco applications. The objective of ASME Code rules, aside from assigning dimensional values, 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 failure is reduced to an acceptable level by the use of the requirements and safety factors that are contained in the ASME Code.


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Pressure vessel codes are also written to be broadly applicable to more than just the refinery and petrochemical industries. Accordingly, these codes cannot anticipate and address every possible design requirement or service application. For example, the ASME Code, Section VIII, Division 1, states that "the user or his designated agent shall establish the design requirements for pressure vessels, taking into consideration factors associated with normal operation, and such other conditions as startup and shutdown." Therefore, the design of pressure vessels for refinery and petrochemical services usually involves factors that are beyond the minimum code requirements. The owner defines what additional factors, beyond normal ASME Code requirements, must be considered in each case. Additional design factors, not specifically covered in the ASME Code, are vibration, thermal or pressure cycles, corrosion, and erosion. Owners apply supplementary requirements in the design, fabrication, inspection, and testing of pressure vessels that are suitable for their applications. Saudi Aramco's Engineering Standards (SAESs) and Materials System Specifications (SAMSSs) are examples of supplementary owner requirements. However, it is often necessary to supplement even the Saudi Aramco standards to cover design requirements for a particular pressure vessel. A purchase requisition for a pressure vessel will specify specific design requirements and the appropriate Saudi Aramco Engineering Standards. It will require that the item meet ASME Code rules. The vessel will be inspected before it leaves the vendor's shop by an inspector who is authorized by the ASME Code authorities. The inspector is responsible for ensuring that all Code and other requirements have been met.


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Structure and Scope of the ASME Code, Section VIII, Division 1, Design By Established Rules The ASME Boiler and Pressure Vessel Code, Section VIII, Division 1, Pressure Vessels, is the primary standard that is used by Saudi Aramco for the design, fabrication, inspection, and testing of unfired pressure vessels. When the words "Code" or "ASME Code" are used in this module, they refer to Division 1, unless stated otherwise. A copy of Division 1 is included in Course Handout 1. ASME Code Structure

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 of construction.

•

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

•

Subsection C covers specific requirements that apply to several classes of materials that are used in pressure vessel construction. 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).

In addition to these subsections, the ASME Code also contains the following appendices: •

Mandatory Appendices address specific 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 in the design and construction of the pressure vessel under consideration.


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•

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

ASME Code Scope

The ASME Code scope defines the circumstances where its rules apply. The Code is applicable for pressures that exceed 103 kPa (ga) (15 psig) and through 20 682 kPa (ga) (3 000 psig). At pressures below 103 kPa (ga) (15 psig), the ASME Code is not applicable. At pressures above 20 682 kPa (ga) (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 are applied only through the first junction that connects to the pipe. This junction may be at the following locations: •

Welded end connection for the first circumferential joint, for cases where welded connections are used.

•

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 is also not applicable to non pressure-containing parts that are welded, or not welded, to pressure-containing parts. However, the actual 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. The exclusions of most interest are: •

Fired process tubular heaters (for example, furnaces).


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•

Pressure containers that are integral parts of rotating or reciprocating mechanical devices (for example, pump, turbine, or compressor casings).

•

Piping systems and their components.

ASME Code, Section VIII, Division 2, Design By Analysis The overwhelming majority of Saudi Aramco pressure vessels are designed in accordance with Division 1. However, the use of Division 2 is preferred in some applications. The ASME Code, Section VIII, Division 2, Pressure Vessels, Alternative Rules, contains requirements that differ from the requirements that are contained in Division 1. Several of the areas where the requirements between the two divisions differ are highlighted below. •

Stress: the maximum allowable stress for a Division 2 pressure vessel is higher than that of a Division 1 pressure vessel. From the standpoint of general primary membrane stress, a Division 2 vessel is less conservative than a Division 1 vessel for the same design parameters and materials. The Division 2 vessel is thinner, uses less material, and costs less. A Division 2 vessel compensates for the higher allowable primary membrane stress by being a much more stringent design standard 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 specific design details are not permitted in Division 2 that are allowed in Division 1.

•

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

•

Quality Control: material quality control is more stringent in Division 2 than in Division 1.


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The areas where Division 2 are more conservative than Division 1 add to the cost of a vessel. The choice between using Division 1 and Division 2 is based on economics. The lower costs that are associated with the use of less material 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. A Division 2 design is more likely to be attractive for vessels that require thicker walls. A wall thickness of approximately 50 mm (2 in.) is a good starting point at which to consider the use of a Division 2 design. The thickness break point is lower for more expensive alloy material than for plain carbon steel, and this break point 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. The Division 2 design criteria provide formulas and rules for the more common configurations of shells and formed heads. Requirements include detailed evaluations of actual stresses in complex geometries and with unusual loadings, especially cyclic loads or those that result in localized stresses. The calculated stresses are assigned to various categories and subcategories. These categories and subcategories have different allowable stress values. These allowable stress values are based on multiples of the basic allowable stress intensity value that is specified in Division 2 for the particular material specification. Participants are referred to Division 2 for additional information on the stress categories and subcategories, and their associated allowable stresses.

Saudi Aramco Pressure Vessel Design Data Sheets Saudi Aramco standard Pressure Vessel Design Data Sheets, also known as Pressure Vessel Data Sheets or Pressure Vessel Design Sheets, are used to: •

Prepare the pressure vessel purchase requisition.

•

Obtain vendor design information.


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•

Ensure a uniform bidding basis among the various vendors-

Ensure format consistency in the information that is provided;

-

Simplify bid comparisons during pressure vessel purchase.

•

Document the as-built details of the vessel.

•

Facilitate retrieval and use of design information after the vessel has been installed.

In addition to meeting the purposes that are listed above, the forms are frequently used by operations, inspection, and maintenance personnel after the vessel has been placed in service. SAES-D-001, Pressure Vessels, refers to other Saudi Aramco documents that must be complied with. Forms 2682 and 2683 are among these documents. Form 2682 is used for Division 1 pressure vessels, and Form 2683 is used for Division 2 pressure vessels. The portion of Form 2682 that contains material selection information was discussed in MEX 202.02. Copies of these forms are contained in Course Handout 3 for reference in subsequent discussions. The pressure vessel engineer is responsible for completion of the appropriate Design Data Sheet to the extent possible for use as part of the vessel purchase requisition. The portions of the form that must be completed at this time are those that are necessary to define the Saudi Aramco and ASME requirements that ensure a uniform bidding basis. Other portions of the form that do not impact the cost quotation basis, or that are determined later during the detailed vessel design phase, are initially left blank. These blank portions are subsequently completed by the vessel vendor as part of his bid. The completed form from the successful vendor will then be included in the pressure vessel's documentation file. When a vessel is rerated, it is not the Design Data Sheet that is revised but, rather, the Safety Instruction Sheet. The Safety Instruction Sheet will be discussed later in this module.


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In most cases, initial data for Forms 2682 and 2683 are provided by engineers who work for a contractor who is employed by Saudi Aramco for a project. Saudi Aramco engineers typically will review the contractor's work to ensure that it is correct. Saudi Aramco engineers fill in the initial data on the forms in cases where no contractor is involved, or when documentation is being prepared for the rerating of an existing pressure vessel to new design conditions. The sections that follow discuss the overall content of Forms 2682 and 2683. Later sections will focus in more detail on filling out specific portions of Form 2682.

Content of Form 2682 for Division 1 Pressure Vessels

Form 2682, provided in Course Handout 3, is the initial form that is used by Saudi Aramco to specify Division 1 pressure vessels. References to appropriate paragraphs within Division 1 are indicated on the form. The following paragraphs highlight particular sections of this form. Some sections may not need to be initially completed, either because they would not impact the initial bidding, or would clearly be covered by 32-SAMSS-004 or ASME Code requirements. Both 32-SAMSS-004 and the ASME Code are intrinsic parts of the purchase requisition. •

The large open space in the upper middle portion of the form is used primarily for the vessel outline drawing for relatively simple vessels, such as drums. For more complex vessels, the outline drawing with appropriate details is produced on separate vessel design sheets (Form 2526, 2527, or 2528). The drawing shows the overall dimensions and orientation of the vessel, nozzle locations, and additional features needed to define the vessel. This open space may also be used for additional notes or design information.


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•

The operating conditions of pressure and temperature must be specified. A minimum temperature of less than 0°C (32°F) must be indicated. Temperatures that are below 0°C (32°F) require additional materials and inspection steps to meet Saudi Aramco requirements. A vessel that is in wet, sour or lethal service must be specified. Wet, sour or lethal services (discussed later in this module) require additional items to meet Saudi Aramco and ASME requirements.

•

The maximum and minimum design temperatures must be specified in accordance with SAES-D-001 requirements.

•

The design pressure must be specified in accordance with SAES-D-001. When the form is first completed, the design pressure is assumed to be equal to the maximum allowable working pressure (MAWP). When the design is complete, the actual MAWP of the vessel is determined on the basis of the nominal vessel thicknesses that are actually supplied. The maximum operating static liquid head and fluid specific gravity must also be specified to permit adequate design. The static liquid head adds to the pressure that is used to determine component thickness, as will be discussed later in this module.

•

The material specifications for major vessel components must be specified. These specifications are based on 32SAMSS-004, as discussed in MEX 202.02. Material selection for vessels in hydrogen service must also consider the potential for hydrogen attack and the Nelson Curve limitations. The hydrogen partial pressure should also be specified on this form, if applicable for the service.

•

The maximum allowable stress values for the specified materials should be entered. These values are taken from the Division 1 allowable stress tables, as discussed in MEX 202.02.

•

The need for impact testing should be specified. This requirement is based on ASME Section VIII, Division 2, as specified by 32-SAMSS-004, even though the overall vessel design is in accordance with Division 1.


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•

The need for ultrasonic testing of plate that is over 50 mm (2 in.) thick and of clad plates should be specified in accordance with 32-SAMSS-004. NOTE: 32-SAMSS-004, which forms part of the purchase

requisition, specifies the requirements for impact testing and ultrasonic testing. The information for these items is entered in the upper left corner of the large space provided for the vessel outline drawing. These items may be determined and reported by the vendor. However, it is preferable to specify them here to ensure that they are properly considered in the vendor quotations. •

The required radiography of the shell seams must be specified as either "spot" or "full." The weld joint efficiency that corresponds with the specified radiography is then determined. SAES-D-001 requires full radiography for all butt welds on vessels with a minimum design temperature below 0°C (32°F) or on vessels to be used in wet, sour service. Paragraph UW-11 of Division 1 also specifies other cases that require full radiography.

•

The required corrosion allowance must be specified. Corrosion allowance was discussed in MEX 202.02.

•

The required wall thickness for internal pressure for the shell and heads must be calculated using the ASME equations that are provided on the left side of the form. If external pressure is a specified design condition, separate calculations must be made and the results entered on the form. The results of the external pressure calculations could require greater wall thickness or the addition of stiffening rings. Internal and external pressure design calculations are discussed later in this module. Any additional wall thickness that is required by other loadings must be determined by the vendor during the detailed engineering phase of vessel design. However, if the magnitudes of external loads are known, these magnitudes should be specified on this form. Examples of external loads include nozzle loads that are imposed by connected piping.


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•

The test pressure may be calculated by the vendor.

•

The vessel engineer must specify if the vessel must be Postweld Heat Treated (PWHT) for reasons other than those that are specified in the ASME Code. PWHT requirements may be determined by the vendor if only ASME requirements apply for the particular vessel.

•

The block that covers the maximum size of shell openings is left blank. ASME rules clearly provide limitations on the size of shell openings and requirements to be met if these limitations are exceeded.

•

The required Class, allowable working pressure, and test pressure of flange connections must be specified. Since the Class is specified on the basis of the design conditions, it should not be the factor that limits in the vessel design. Flange Class selection is discussed later in this module.

•

The block that covers the design loads is typically left blank, unless specific other loads are known in advance. The vendor or contractor will determine any unspecified loads during the detailed engineering phase, and the vendor will then evaluate the impact of these loads on vessel design.

•

The estimated vessel weights and capacity are specified by the vendor as part of detailed engineering.

•

The vessel Maximum Allowable Working Pressure (MAWP) will be specified by the vendor after the nominal thicknesses have been determined during detailed engineering. MAWP will be discussed later in this module.

•

The nozzle schedule, reference mark, size, ASME/ANSI B16.5 rating, and service must be specified.

•

The required material for internal components, flanges, gaskets, and bolting must be specified.


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Content of Form 2683 for Division 2 Pressure Vessels

Form 2683, provided in Course Handout 3, is the form that is used by Saudi Aramco to specify Division 2 pressure vessels. Form 2683 is similar to Form 2682 and has comparable sections to be completed. References to appropriate paragraphs within Division 2 are indicated on the form. The previous discussion of Form 2682 applies to similar sections on Form 2683. An additional requirement for a Division 2 pressure vessel is the potential need for a fatigue analysis. Paragraph AD-160 of Division 2 contains rules that determine the need for a fatigue analysis. These rules are based on the specific parameters that follow: •

The number of pressure and temperature cycles expected.

•

The magnitude of the expected pressure and temperature cycles.

•

The expected metal temperature differences between adjacent points on the vessel.

If the specified criteria indicate that a fatigue analysis is required, the data that are needed to define the pressure and temperature cycles must be specified. The Design Data Sheet must specify where the cycle data may be found. The vendor then determines the impact that the cycle data have on his design and bid. A fatigue analysis will normally not be required for typical Saudi Aramco pressure vessel applications.


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EVALUATING THE ACCEPTABILITY OF CONTRACTOR-SPECIFIED DESIGN CONDITIONS AND LOADINGS Mechanical engineers are not responsible for specifying either the operating or design conditions for a pressure vessel. Specification of these conditions is the responsibility of the process design engineer. However, mechanical engineers must be aware of the differences between operating and design conditions. Process engineers, on the other hand, must understand the impact that operating and design conditions have on the mechanical design of pressure vessels. The operation and design conditions that are specified should not be overly conservative or liberal. The mechanical design of a pressure vessel begins with specification of the design pressure and design temperature. These two parameters must be specified together to obtain the correct mechanical design details. Pressure imposes loads on a pressure vessel that must be withstood by the individual vessel components. Temperature affects material strength and, thus, 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 operating state may govern the mechanical design of the reactor components. All pressure vessels must be designed for the most severe conditions of coincident pressure and temperature that are expected during normal service. This requirement is stated in the ASME Code, Section VIII, Division 1. Normal service must include conditions that are associated with: •

Startup.

•

Normal operation.

•

Deviations from normal operation that can be anticipated (for example, catalyst regeneration or process upsets).

•

Shutdown.


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Pressure vessels must also be designed for other loading conditions and service factors that may apply. These conditions and factors, as well as pressure and temperature, are discussed in the following sections. Work Aid 1 may be used to assist in evaluating the acceptability of design conditions and loadings. Pressure Both operating and design pressure must be considered in pressure vessel design. Operating Pressure

The operating pressure must be set on the basis of the maximum internal or external pressure that the pressure vessel may encounter. The following factors must be taken into account: •

Ambient temperature effects.

•

Normal operational variations.

•

Pressure variations due to changes in vapor pressure.

•

Pump or compressor shut-off pressure.

•

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

•

System pressure drop.

•

Normal cleaning and pre-startup activities if other conditions may occur, such as vacuum, that should be considered in the design.


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Design Pressure

Generally, design pressure is the maximum internal pressure, in kPa (gauge) or psig, that is used in the mechanical design of a pressure vessel. For full or partial vacuum conditions, the design pressure is the maximum pressure difference that can occur between the atmosphere and the inside of the pressure vessel. The design pressure is applied externally for vacuum conditions. Specific 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 internal or external pressure, depending upon which of these is the more severe pressure condition. More specifically, the design pressure of a pressure vessel is the pressure that is expected at the top of the vessel. The design pressure is normally 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 equal to the pressure vessel design pressure. SAES-D-001 specifies Saudi Aramco requirements for setting the design pressure and considers the possibility of either external or internal pressure conditions. There may be cases where a vessel is not in vacuum service during normal operation or in an upset, but may be subject to steam-out conditions which can cause an external pressure condition. If steam-out is possible, the pressure vessel must be designed for an external pressure of 52 kPa (ga) (7.5 psig) at 149°C (300°F). Work Aid 1 summarizes the procedure for setting design pressure based on Saudi Aramco requirements.


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Pressure vessels, especially tall towers, typically will 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 weight of the liquid that is contained in the vessel must be considered in the design, as will be highlighted below. The hydrostatic pressure of the liquid must also be considered in the design of the vessel components. 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.


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Figure 1 illustrates this concept. Work Aid 1A contains equations that are used to calculate the pressure below the liquid level.

PT = Design pressure at top of vessel.

γ

= Weight density of liquid in vessel.

H = Height of liquid.

;; ;

PBH = Design pressure of bottom head.

20203.F01

Figure 1: Design Pressure


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The pressure vessel designer must determine the minimum thickness that is required for each vessel section. The thickness calculation must be made for both the design and hydrostatic test cases. The minimum specified wall thickness is then based on the more severe condition. For example, determination of the minimum specified wall thickness of each section of a tall vertical tower must include consideration of the design pressure at the top of the vessel, plus the hydrostatic head that is applicable at the level being designed. Several wall thickness plates are commonly used in a tall, liquid-filled tower. Thicker plates are used in the lower sections because of higher hydrostatic pressure and higher bending moments caused by wind.

Temperature Both operating and design temperatures must be considered in pressure vessel design.

Operating Temperature

The operating temperature must be set on the basis of the maximum and minimum metal temperatures that the pressure vessel may encounter. The operation and vertical length of some pressure vessels result in large temperature reductions between the bottom and top of the vessel. For example, atmospheric and vacuum pipestill towers are typically very tall and have liquid in the bottom portion and vapor in most of the other sections. The temperature of the liquid in the bottom will be much higher than the temperature of the vapor in the top. It is permissible to specify different operating 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 2 shows a tall vessel, and illustrates the range of sections that have different design temperatures.


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; Engineering Encyclopedia

Design of Pressure Vessels

Specifying Design Requirements for Pressure Vessels

Section 4 (T-Z)

Section 3 (T-Y)

Section 2 (T-X)

Section 1 (T), °C (°F)

Support skirt Grade

20203.F02

MEX 20203.F04

Figure 2: Temperature Zones in Tall Vessels


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Sudden cyclic changes in temperature also occur during normal operation and may be associated with only minor pressure fluctuations. In this case, the design is governed by the highest and lowest probable operating temperatures of the material that is in contact with the metal, and by the corresponding pressure at those temperatures.

Design Temperature

The design temperature of a pressure vessel is the fluid temperature that occurs under normal operating conditions, plus an allowance for variations that occur during operation. SAESD-001 specifies the basis for setting the design temperature, and this basis is summarized in Work Aid 1.

Minimum Design Metal Temperature

A Minimum Design Metal Temperature (MDMT) must also be specified for pressure vessel design. A MDMT is specified to ensure that materials that have adequate fracture toughness are selected for construction. Fracture toughness was discussed in MEX 202.02. SAES-D-001 specifies the factors influencing the minimum design metal temperature. For most pressure vessels, the minimum design metal temperature equals the minimum design ambient temperature for the construction site. Some services subject pressure vessels to extremely low temperatures during normal operations or process upsets. In these cases, the lower operating temperature must be considered to determine the minimum design metal temperature. For example, if autorefrigeration is possible, it must be considered in the determination of the minimum design metal temperature. Autorefrigeration may occur during startup, shutdown, upset, failure of a piping component, or malfunction of a pressure relief device.


26



Other Loadings Paragraph UG-22 of the ASME Code, Section VIII, Division 1, specifies the loadings that must be considered to determine the minimum required shell thicknesses for the various vessel sections. These design loadings are as follows: •

Internal or external design pressure.

•

Weight of the vessel and its normal contents under operating or test conditions. This weight includes any additional pressure that is due to the static head of liquid.

•

Superimposed static reactions from the weight of attached equipment, such as motors, machinery, other vessels, piping, linings, and insulation.

•

The attachment 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.

In its simplest form, the design thickness for a pressure vessel shell component may be determined based on design pressure conditions alone. However, the detailed design of a pressure vessel must consider all combinations of pressure, weight, and external and internal loads that may occur in actual operation. The stresses that occur during erection of the vessel at the site must also be considered. It is normal practice to assume that wind and earthquake loads do not occur simultaneously. Therefore, the vessel is designed for the worst of either wind or earthquake.


27



The directions, types, and locations of stresses that are produced by the various load combinations must be evaluated using appropriate criteria. Some loads, such as wind, will affect the thickness of an entire shell section. Other loads, such as those from piping systems, affect only a local region of the vessel. A complete consideration of all loadings that act on a pressure vessel is a complicated process and is normally done using computer programs. While the Saudi Aramco pressure vessel engineer does not have to calculate the effects of all these loadings, he must be aware of what a pressure vessel designer must do to achieve a correct design. It is also not normally necessary to calculate all of these loadings before the Pressure Vessel Design Data Sheet is completed, to determine the effect, if any, of the loadings on vessel design. These loadings will generally have, at most, a localized effect on vessel design and should not be a major factor in distinguishing one vendor's bid from another. Vendors know that they will have to consider wind, hydrotest, piping, and internal loads in their final designs, and they typically will provide cost allowances for these considerations in their bids as necessary. The Pressure Vessel Design Data Sheet has an area where it may be indicated whether these loads were considered.


28



Figure 3 shows a typical vertical reactor with additional loads applied.

Pipe connection with imposed forces and moments

Catalyst bed with liquid holdup

Wind

Support grid

Grid support welded to shell

;; ;

Earthquake

20203.F03

Figure 3: Vertical Reactor with Additional Loads


29



32-SAMSS-004 requires that the following load conditions be used in the design of pressure vessels and their supports: •

Full wind or seismic load combined with all dead loads except operating liquid. This is the case for the condition when the vessel is erected but not operating and when it is exposed to the design wind or seismic load (whichever is greater).

•

Internal design pressure or external pressure loads, plus total operating weight (vessel in erected condition, trays, operating liquid) and full wind or seismic load. This case is for the condition when the vessel is under normal operation and is exposed to the design wind or seismic load (whichever is greater).

•

Test pressure combined with hydrotest weight (weight of vessel in erected condition and weight of test water), plus wind or earthquake loads reduced by 40%. This case is for the condition when the vessel is being hydrotested in the field. It is assumed that the hydrotest would proceed if the wind velocity is up to the specified value but would not be done at higher wind velocities or during an earthquake.

Weight

The weight and location of the following exterior and interior attachments must be considered in determining the dead weight load that acts on a pressure vessel: •

Attached equipment and piping

•

Catalyst bed supports

•

Fireproofing

•

Insulation

•

Platforms

•

Trays


30



The dead weight load must also include the weight of the vessel contents, such as liquid, catalyst, and inert balls, as well as the weight of the vessel shell itself. Weight load calculations must be made both for the operating fluid in the vessel, for the maximum design liquid level, and for the case with the vessel completely filled with water for hydrotest. The hydrotest calculations must be made for both the shop and field testing. Loads that act eccentrically to the vessel axis are resolved into forces and moments that act along the vessel axis. Examples of eccentrically applied loads are the weight of platforms that are attached to the side of a tower or the weight of a heat exchanger that is bolted to a nozzle on the side of a tower. These dead weight loads result in longitudinal compressive stresses in the vessel shell and supports. CSE 110 provided procedures for the calculation of weight loads. The general approach is as follows: •

Calculate the metal volume of the shell, heads, and support. This calculation takes into consideration the geometry of the individual components.

•

Calculate the weight of metal based on the volume that is found.

•

Determine the internal volume of the vessel shell and heads.

•

Calculate the weight of water for the hydrotest based on the total volume of the vessel.

•

Calculate the weight of operating liquid in the vessel based on its maximum fill height.

•

Determine the weight of appurtenances, and insulation.

•

Calculate the total operating and hydrotest weights of the vessel.

internal

components,

Refer to CSE 110.02 for the specific procedures to use and sample problems. Note also that the Pressure Vessel Design Data Sheet has an area where the vendor fills in the estimated vessel weights for three cases: empty with internal components, operating condition, and filled with the test medium.


31



Wind

Wind loads are imposed on all pressure vessels that are located outside of buildings. Wind loads induce stresses in the vessel shell components. However, wind loads are only a significant design consideration for tall, vertical pressure vessels. Wind load evaluation requires consideration of the two concepts that are discussed in the paragraphs that follow. Pressure applied across the pressure vessel surface due to wind velocity produces forces along the vessel length. These forces

produce bending moments along the length of the vessel that, in turn, produce additional longitudinal stresses in the vessel shell. These longitudinal stresses are in tension on the windward side of the vessel and in compression on the leeward side. Wind loads also cause vertical vessels to deflect. This vessel deflection must be kept within reasonable limits in order not to adversely affect process operations, for example by causing a nonuniform liquid flow across distribution trays or by creating an unsafe situation for personnel who may be on the vessel. The following must be considered in the design of pressure vessels for wind loads: •

The pressure vessel data sheet specifies the design wind velocity to be used for pressure vessel design. It must also be assumed that a field hydrotest can take place if there is a wind up to 60% of the design wind speed.

•

The effective vessel diameter, De, considers such factors as insulation, piping, platforms, and ladders, and determines the vessel area that is exposed to the wind. This determination is used to calculate the forces and bending moments that are imposed on the vessel. Figure 4 shows a generalized tower layout and illustrates how items that are attached to the vessel effectively increase the diameter that is exposed to the wind.


32



Caged ladder

Insulation thickness Insulation thickness Distance between platforms

Platform

Vessel OD

Pipe OD

20203.F04

Figure 4: Tower Layout for Determining Effective Diameter for Wind Calculations


33



CSE 110.02 provided procedures to calculate the wind loads that are imposed on a tall tower. The general approach is as follows: •

Determine the gust factor based on the height of the vessel.

•

Determine the height correction factors for various elevations along the total height of the vessel.

•

Determine the shape factor for the vessel.

•

Determine the equivalent vessel diameter.

•

Calculate the lateral shear forces that are applied to the vessel at various elevations along the total height.

•

Calculate the bending moment that is applied to the vessel due to the shear forces. The maximum moment occurs at the base.

Refer to CSE 110.02 for the specific procedures to use, and for sample problems. Note that these procedures may be used to calculate the applied loads; however, do not use these procedures to calculate the stresses in the vessel that result from these loads. This stress calculation is typically done by the vendor during detailed engineering. Note that the Pressure Vessel Design Data Sheet has a location where an indication may be made of whether the design includes wind loads. In most cases, this item will either be left blank or answered "No." Wind design calculations are typically not done to complete this sheet. Therefore, any component thicknesses specified on the Pressure Vessel Design Data Sheet will not include consideration of wind. However, this omission should not be interpreted to mean that wind does not need to be considered by the vendor in his final design.


34



The dynamic effect of vortex shedding occurs when wind flows past

a circular pressure vessel and when the air behind the vessel is no longer smooth. A region of pressure instability occurs where vortices are shed in a regular pattern. The vortex shedding alternates from one side of the vessel to the other. These vortices cause an alternating force to act perpendicular to the wind direction and causes the vessel to vibrate. The concept of vortex shedding is illustrated in Figure 5.

Shed vortex

Wind

Tower cross section

20203.F05

Figure 5: Vortex Shedding

When the vortex-shedding frequency coincides with the mechanical natural frequency of the vessel, mechanical resonance of the vessel occurs. Mechanical resonance causes an increase in vibration amplitudes, and fatigue failure of vessel sections can eventually result. The vessel must be designed so that its natural frequency is high enough to avoid resonant vibration. The parameters that affect this phenomenon are wind velocity, vessel diameter, and vessel height.


35



Hydrotest

Pressure vessels are required to pass a hydrostatic pressure test. In the hydrotest, the vessel is filled with water, either in the shop or in the field, and pressurized to the prescribed test pressure. The vessel, its associated support, and its foundation must be designed for the weight of test water, as discussed earlier. A vertical vessel is often designed to be tested horizontally in the shop. In this case, the following must be considered: •

It may be necessary to retest the vessel in the field after repairs or modifications are made.

•

During normal operation, the vessel may be accidentally filled with liquid or solids.

Given the above considerations, a vertical pressure vessel is normally designed so that it may also be tested in the installed position, even if the original shop hydrotest is in the horizontal position. The hydrotest verifies the structural adequacy of the pressure vessel. The hydrotest also provides some mechanical stress relief before the application of service conditions, as long as the test temperature is above the vessel material's ductile-to-brittle transition temperature. MEX 202.02 discussed the ductile-tobrittle transition temperature. Paragraph UG-99 of the ASME Code, Section VIII, Division 1, provides procedures to determine hydrotest pressure. Saudi Aramco supplements these requirements in 32-SAMSS-004, Pressure Vessels, and in procedures that are contained in the Pressure Vessel Design Data Sheet. These procedures are completely discussed in MEX 202.04.


36



The Pressure Vessel Design Data Sheet has a location where an indication may be made of whether the design includes hydrostatic test loading. In most cases, this item will either be left blank or answered "No." Stresses that result from hydrotest weight loads will generally not be calculated when this sheet is completed. Therefore, the specified component thicknesses will not reflect the hydrotest. The effect of hydrotest weight load on component thickness, if any, will be determined by the vendor during detailed engineering. External Piping

Most pressure vessel nozzles have piping systems attached to them. The nozzles act as anchor points for the piping systems and absorb forces and bending moments that are imposed on the vessel by the piping system. These forces and bending moments are caused by weight and differential thermal expansion of the piping. Piping systems produce additional stresses in the pressure vessel nozzle and adjacent areas of the shell. These additional stresses are localized and diminish away from the nozzle. The detailed design of the nozzle and adjacent shell must be strong enough to maintain these stresses within allowable limits. To accomplish this stress control, the following are sometimes necessary: •

Increase the nozzle thickness.

•

Increase the nozzle reinforcement pad size.

•

Use a thicker section of vessel shell in the local area.

Nozzle design modifications that are required to accommodate piping loads are not necessary in most cases. Nozzle and shell designs are generally strong enough to accommodate design pressure and to absorb piping loads. However, special attention should be paid to very low-pressure applications and largediameter nozzles (over 600 mm [24 in.]). In the low-pressure case, the vessel shell may be relatively thin and the nozzle may not need reinforcement for pressure. Thus, there may not be adequate inherent strength to absorb the additional piping loads. In the large-diameter nozzle case, piping reaction loads from differential thermal expansion may be high.


37



The evaluation of loads imposed on pressure vessel nozzles will be discussed further in a later section of this module. Here again, the Pressure Vessel Design Data Sheet may indicate if piping system reaction loads have been considered in the design. This item will typically be left blank or answered "No." Loads from attached piping systems will typically not be available when the sheet is first completed since the piping systems have not yet been designed. Therefore, the specified component thicknesses do not consider these piping loads. The vendor will typically be supplied with piping loads later, and he will determine their effect on vessel design then. Internal Components

Pressure vessel internal components, such as tower trays and reactor catalyst bed supports, are examples of components that are typically supported from the vessel shell. The weight and bending moment loadings that are associated with pressure vessel internal components induce stresses in the vessel shell and support attachment welds to the shell. These stresses must be kept within allowable limits, and the internal components and their associated supports must be designed by the pressure vessel vendor. 32-SAMSS-004 requires that all internal and external supports, support rings, pads, and structural brackets that are attached to the vessel be seal-welded all around. This seal-welding prevents any corrosion between the shell and the attachment. The evaluation of loads that are imposed by attachments to a pressure vessel shell or head will be further discussed later in this module. The Pressure Vessel Design Data Sheet may be used to designate whether loads due to internal components were considered in the specified design. This item will typically be left blank or answered "No." The vendor will typically determine the effect of loads due to internal components as part of his detailed design.


38



Earthquake Loadings

Earthquake loadings on a vessel result from a sudden, erratic, vibratory motion of the ground that supports the vessel. The vessel responds to this motion. The main factors that cause vessel damage are the intensity and the duration of the earthquake motion. The forces and stresses in the vessel shell are transient, dynamic, and complex. Accurate evaluation of earthquake forces and the vessel stresses that these forces cause requires computer analysis. To simplify vessel design procedures, the vertical component of the earthquake motion is normally disregarded. This approach is acceptable since most structures have enough excess strength in the vertical direction to be considered earthquake resistant. The horizontal, or lateral, earthquake forces are then reduced to equivalent static forces that act on the vessel. Earthquake design relies mainly on experience and observation. It is based on the performance of structures that have been previously subject to earthquakes. Structures that are built in seismic risk zones must be designed to withstand a minimum horizontal shear force applied at the base of the vessel in any direction. The shear force is translated into equivalent static forces through the height of the vessel. The static forces are used to calculate the shear forces, bending moments, and resulting stresses through the height of the vessel. The simplified equation that follows may be used to estimate the lateral seismic force at the base of the vessel: V = ZIKCSW Where:

V = Lateral seismic force at the base of the vessel, kg (lb.). Z = Seismic probability coefficient for the site. I = Importance factor; assume I = 1.0 for a pressure vessel. K = Arrangement factor; assume K = 2.0 for a pressure vessel.

(

)

C = Base shear factor; C = 1/ 15 T ≤ 0.12 .


39



T = Fundamental period of vibration of the vessel, assuming a uniformly loaded cantilever beam fixed at the base. S = Site-coefficient based on soil characteristics, assuming S = 1.5 unless an exact value is known. W = Total dead weight load of the vessel and contents above the plane being considered, kg (lb.). CSE 110.02 contains a simplified procedure to calculate the earthquake loadings on a tall, single-diameter tower. This procedure involves calculation of the shear force and overturning moment applied at the base. When earthquake is a design consideration, the calculated loadings are used to determine the resulting stresses in the shell.

Service The pressure vessel service can affect material selection, fabrication, and inspection requirements. Therefore, the service must be specified on the Pressure Vessel Data Sheet. In addition, it also must be specified whether the vessel is in wet, sour service or in lethal service. Material selection requirements based on service considerations were discussed in MEX 202.02. Fabrication and inspection requirements based on service considerations will be discussed in MEX 202.04.

Wet, Sour

In-service cracking at welds is possible in a wet, sour process environment. This is called stress corrosion cracking and is caused by the combined action of tensile stress and corrosion in the presence of water and H2S. Wet, sour service must be specified, when applicable, for the vessel vendor to know that additional Saudi Aramco material specification and inspection requirements are applicable, and must be reflected in his bid.


40



Lethal

The owner must identify to the vendor whether the pressure vessel will contain a lethal substance. The ASME Code Section VIII, Division 1, defines a lethal substance as: "Poisonous gases or liquids of such a nature that a very small amount of the gas or of the vapor of the liquid mixed or unmixed with air is dangerous to life when inhaled. For purposes of this Division, this class includes substances of this nature which are stored under pressure or may generate a pressure if stored in a closed vessel." It is very rare for a refinery service to be in this category. When the lethal service category does apply, the ASME Code requires additional measures to increase vessel quality. These measures include 100% radiography of all butt welds, PWHT of carbon and low alloy steel materials, and restrictions on the use of certain carbon steel material specifications. The vendor must consider these requirements in his cost quotation.


41



EVALUATING THE ACCEPTABILITY OF CONTRACTOR-SPECIFIED PRESSURE VESSEL COMPONENT THICKNESS DESIGN CRITERIA

Two additional pressure vessel design criteria are required before vessel component thicknesses can be calculated (by the vendor) and in turn evaluated for acceptability (by Saudi Aramco engineers). These additional design criteria are: •

Weld joint efficiency

•

Corrosion allowance

Weld Joint Efficiency Weld joint efficiency (E) is used to account for the quality of a welded joint and for the concentration of local stress. Determination of a stress concentration factor takes into consideration the fact that the stress in a localized region of a component or structure may be higher than would be calculated if normal static analysis were used. This higher local stress is due to local material or structural discontinuities. Stress concentration in welded joints arises from the following factors: •

The geometry of the weld itself.

•

The metallurgical structure of the weld with respect to the base metal.

•

Weld defects, such as slag inclusions, shrinkage cracks, or porosity.

The last two factors are functions of the procedure that is used to make the weld. The first factor is the main source of stress concentration and can be controlled by the vessel design engineer. The net effect of the three stress concentration factors is to reduce the fatigue strength or efficiency of the weld. Weld joint efficiency must be considered in the vessel design.


42



Paragraph UW-12 of the ASME Code, Section VIII, Division 1, specifies the weld joint efficiencies to be used in the formulas for pressure vessel component thicknesses. These efficiencies depend on the type of weld joint design that is used and on the degree of weld radiographic examination that is made. Types of welded joints and weld inspection methods will be discussed in MEX 202.04. For our purposes here, it is only necessary to know that the efficiency of a weld joint is determined by the type of weld joint that is used and by the extent of its radiographic inspection. The ASME Code also defines weld joint categories based on the location of a joint in a vessel. The Code then specifies, by example, the weld joint designs that may be used in each category. MEX 202.04 discusses weld joint types and inspection further. 32-AMSS-004 requires 100% joint efficiency for all hydrogen, wet sour, lethal, cyclic and unfined steam drum services. 100° weld efficiency is also required for hydrocarbon, steam, amine and caustic services above 120°C (250°F). Figure 26 in Work Aid 2A is excerpted from the ASME Code, Section VIII, Division 1, and identifies pressure vessel weld joint categories. Figure 28 in Work Aid 2A is also excerpted from the ASME Code and defines weld joint efficiencies based on the type of weld (shown in Figure 27 of Work Aid 2A) and degree of radiographic examination. Figure 28 shows that the weld joint efficiency decreases as the degree of radiography decreases for a given type of weld joint. Note from Figure 28 that the direction of weaker weld joint designs is vertically downward and that lower weld joint efficiencies correspond with weaker weld joint designs. The majority of pressure vessels use a Type 1 joint design. A Type 1 joint design has a weld joint efficiency of either 0.85 or 1.00; these values correspond with either spot or full radiographic examination. Later discussion of the ASME Code calculation formulas will show that the required shell and head thicknesses increase with decreasing weld joint efficiency. Work Aid 2A summarizes how to evaluate the acceptability of the specified weld joint efficiency. The degree of radiography and corresponding joint efficiency are specified on the Pressure Vessel Design Data Sheet.


43



Corrosion Allowance Corrosion was discussed in COE 103 and COE 105. Corrosion, erosion, or abrasion causes the components of a pressure vessel to thin during the operating life of the vessel. In order to compensate for this thinning, components must have their thicknesses increased over those that are calculated based on the ASME Code design formulas. Internal corrosion/erosionresistant linings are sometimes used as an alternative to the use of greater component thicknesses. Process design and materials engineers typically specify the corrosion allowance for pressure vessel components. These allowances are based on determinations of the expected corrosion rate for the vessel material in the anticipated process environment. The expected corrosion rate is multiplied by the design life of the vessel (normally 20 years) to determine the corrosion allowance that is to be used in the vessel design. The expected corrosion rate is a major factor that influences material selection. A high corrosion rate for a material in a particular service requires a large corrosion allowance. This larger corrosion allowance requires thicker components and increases the cost of the pressure vessel. It is often possible to use a higher-alloy material that has a lower corrosion rate and corrosion allowance in the same service. In many cases, the greater cost per pound of the higher-alloy material is offset by the ability to use thinner components, and therefore, less material. Saudi Aramco minimum corrosion allowance requirements for carbon steel were discussed in MEX 202.02. The required corrosion allowance must be specified on the Pressure Vessel Design Data Sheet to permit the vendor to determine the required component thicknesses. Work Aid 2B summarizes how to evaluate the acceptability of the specified corrosion allowance.


44



EVALUATING CONTRACTOR-SPECIFIED DESIGN CALCULATIONS FOR PRESSURE VESSEL COMPONENTS

This section discusses the ASME Code calculations for the following pressure vessel components: •

Shells

•

Heads

•

Conical sections

•

Flat covers

•

Nozzles

•

Nozzle flanges

The calculations required to determine the stresses that result from local loads that are applied to nozzles and attachments to the vessel will also be discussed in general terms. The ASME Code, Section VIII, Division 1, requires the minimum thickness of shells and heads to be 1.6 mm (0.0625 in.) for most applications, regardless of calculation results. This is the thickness after the vessel components are formed and before corrosion allowance is added. This minimum thickness requirement provides a basic level of mechanical strength for the pressure vessel, even if the calculations indicate that the vessel may be thinner for the design loads that are actually imposed.

Design for Internal Pressure This section discusses calculation of the wall thickness of shells, heads, and conical sections under internal pressure. Refer to Work Aid 3A for the ASME equations that are required to perform the calculations.


45



Shells

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

Pr Pr and σ1 = t 2t

These equations assume a uniform stress distribution throughout the thickness of the shell. Since this is an idealized state, the ASME Code formulas have been modified to account for non-ideal behavior. For hoop stress, the formula is: σθ =

Where:

Pr 0.6P + tE 1 E1

P = Internal design pressure, kPa (psig) r

= Inside radius of the vessel, mm (in.)

t

= Vessel thickness, mm (in.)

E1 = Weld joint efficiency for a longitudinal joint Rearranging this equation and substituting S (allowable stress, kPa [psi]) for σθ yields: t=

Pr SE1 − 0. 6P

The formula that follows applies when the thickness required to resist the longitudinal stress due to internal pressure must be calculated: t=

Pr 2SEc + 0.4P

Where: Ec = Weld joint efficiency for circumferential joints


46



Longitudinal stress can govern the design of particular sections in a pressure vessel. Longitudinal stress can govern when loadings other than internal pressure induce longitudinal stresses that are greater than one half of the hoop stress that is due to internal pressure. 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 of the vessel material at the design temperature. The most common example of where longitudinal stress can govern the design of a vessel component is when wind load on a tall tower causes a bending moment. This bending moment creates a longitudinal bending stress in the cylindrical sections and increases at lower tower elevations because of the greater tower length on which the wind pressure acts. The wind load sometimes requires that the thickness of the lower tower sections be increased beyond the thickness that is required for internal pressure alone, in order for the longitudinal stresses due to wind plus internal pressure to be acceptable. The thickness of a spherical shell will be approximately half the thickness of a cylindrical shell for the same design conditions, material, and diameter. Refer to Work Aid 3A for the ASME Code shell thickness equations. When the required thickness for internal pressure must be determined, the specified corrosion allowance must first be added to the new vessel inside radius so that the corroded vessel inside radius is used in the equations. The thickness calculated using these equations must then be increased by the specified corrosion allowance in order to arrive at the minimum required new vessel thickness. Note that the Pressure Vessel Design Data Sheet has an area where the thickness calculation equations are summarized.


47



Sample Problem 1 Cylindrical Shell Thickness Calculation

The pressure vessel described in Figure 6 will be used in this and subsequent Sample Problems. Hand calculations are used in the solution of all Sample Problems, Exercises, and Evaluations in this course to assist in understanding the design concepts and parameters that are involved. However, computer programs are typically used for these calculations on the job. Saudi Aramco engineers often use the CODECALC computer program for these calculations. The geometry and design data of a vertical cylindrical pressure vessel are specified in Figure 6. Cost estimates are being prepared for this vessel. It is your job to estimate the required component thicknesses. What are the minimum required thicknesses for the two cylindrical sections?


48



Hemispherical DESIGN INFORMATION Design Pressure = 250 PSIG Design Temperature = 700°F Shell and Head Material is SA-515 Gr. 60 Corrosion Allowance = 0.125" Both Heads are Seamless Shell and Cone Welds are Double Welded and will be Spot Radiographed The Vessel is in All Vapor Service Cylinder Dimensions Shown are Inside Diameters

4'-0"

60'-0"

10'-0"

6'-0"

30'-0"

2:1 Semi-Elliptical

MEX 20203.F06

Figure 6: Sample Problem 1


49



Solution Since the welds are spot radiographed, E = 0.85. S = 14 400 psi for SA-515/Gr. 60 at 700°F. Use Work Aid 3A for this solution. 6 ft. - 0 in. Shell r = 0.5D + C = 0.5 × 72 + 0.125 r = 36.125 in. tp =

Pr 250 × 36.125 = SE1 − 0.6P 14 400 × 0.85 − 0.6 × 250

tp = 0.747 in. t = tp + c t = 0.747 + 0.125 t = 0.872 in. required including corrosion allowance 4 ft. - 0 in. Shell r = 0.5 × 48 + 0.125 r = 24.125 in.

tp =

250 × 24.125 14 400 × 0.85 − 0. 6 × 250

tp = 0.499 in. t = 0.499 + 0.125 t = 0.624 in. required including corrosion allowance


50



Heads

Figure 7 shows typical types of formed closure heads that are used on pressure vessels. Elliptical, hemispherical, and torispherical are the most commonly used head types. Note in Figure 7 that all head types but the conical head have a straight flange (sf) section, which simplifies welding the head to the adjacent cylindrical shell section. The elliptical and torispherical head types have an indicated head depth (h), which is measured from the straight flange to the maximum point of curvature on the inside. As discussed previously for shells, the internal head dimensions that are used to calculate the required thicknesses must first be increased to account for the specified corrosion allowance. The following procedure is used to adjust the internal head dimensions: •

In equations where the head inside radius is a parameter, the specified corrosion allowance must first be added to the new head inside radius so that the corroded inside radius is used in the equations.

•

In equations where the head inside diameter is a parameter, double the specified corrosion allowance and then add this number to the new head inside diameter so that the corroded inside diameter is used in the equations.

The pressure vessel corrosion allowance must then be added to the thicknesses that are calculated by the ASME equations for these heads. A different equation is used to calculate the thickness of each head type. Work Aid 3A contains the ASME Code equations that are used to calculate the wall thicknesses of heads.


51



t t R sf sf ID

ID Hemispherical

Flanged t

t

h

h

sf

sf Flanged and dished (torispherical)

Elliptical

α

α

t

t

r ID

ID Conical

sf

Toriconical

MEX 20203.F07

Figure 7: Typical Formed Closure Heads


52



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

To simplify calculations and fabrication, the ASME Code permits an approximation for the actual head geometry. The elliptical head geometry may be assumed to consist of the following: •

A spherically dished head with a radius that is equal to 90% of the inside diameter of the shell to which the head is attached, and

•

A knuckle radius of 17% of the inside diameter of the shell to which the head is attached. The knuckle is the transition region between the straight flange and the spherically dished area.

The Pressure Vessel Design Data Sheet has a location where elliptical head calculations are done. 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 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. Hemispherical heads are normally fabricated from segmented sections that are welded together, spun, or pressed. Segmented hemispherical heads may be economical in carbon steel construction for thin, very large-diameter vessels, or in thick, small-diameter vessels. Hemispherical Heads -

A hemispherical head is typically half the wall thickness of the cylindrical shell to which it is attached. Therefore, the thickness transition zone between the head and shell must be contoured to minimize the effect of local stress. Figure 8 shows the thickness transition requirements that are contained in the ASME Code.


53



l > 3y

Thinner part

th

Thinner part

th

l > 3y Tangent line

y

Length of required taper, l, may include the width of the weld ts

y

ts

20203.F08

Figure 8: Thickness Transition Between Hemispherical Head and Shell

Note that the equation shown for a torispherical head on the Pressure Vessel Design Data Sheet reduces to the same equation as for a spherical head, since M=1 for a spherical head. 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.

Torispherical Heads -


54



Sample Problem 2 Head Thickness Calculation

For the same vessel described in Sample Problem 1 (See Figure 6), what are the minimum required thicknesses for the top and bottom heads? Solution Since both heads are seamless, E = 1.0. Use Work Aid 3A for the solution. Top Head Hemispherical head r = 24 + 0.125 = 24.125 in. tp =

Pr 250 × 24.125 = 2SE1 − 0. 2P 2 × 14 400 × 1− 0.2 × 250

tp = 0.21 in. t = tp + c t = 0.21 + 0.125 t = 0.335 in. required including corrosion allowance Bottom Head 2:1 Semi-Elliptical Head D = 72 + 2 × 0.125 D = 72.25 in. tp =

PD 250 × 72.25 = 2SE − 0. 2P 2 × 14 400 × 1− 0.2 × 250

tp = 0.628 in. t = 0.628 + 0.125 t = 0.753 in. required including corrosion allowance


55



Conical Sections

Tall towers commonly have sections with different diameters along their length. The transitions between the different diameters are made in conical sections. 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 Work Aid 3A. 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 the localized stresses. When knuckles are used, the transition is called toriconical. Knuckles are 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, but these design procedures will not be discussed in this course. The Pressure Vessel Design Data Sheet does not have a location for the calculation of the thickness of a conical shell section. Therefore, conical shell section calculations must be added by hand when applicable. Sample Problem 3 Conical Section Thickness Calculation

For the same vessel described in Sample Problem 1 (See Figure 6), what is the minimum required thickness of the conical section? Assume that the entire cone will be the same thickness. Solution E = 0.85 since the welds are spot radiographed. Use Work Aid 3A for the solution. tp =


PD 2 cos α(SE − 0.6P)

56



Determine the cone half-apex angle, α. α = tan−1

0. 5(DL − DS ) Cone Length

−1 0.5(72 − 48)

α = tan

α = 5.7°

120

less than 30°, so OK.

D = 72 + 2 × 0.125 D = 72.25 in. tp =

250 × 72.25 2 cos 5.7° (14 400 × 0.85 − 0.6 × 250)

tp = 0.751 in. t = tp + c t = 0.751 + 0.125 t = 0.876 in.


57



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, such as shells and heads, behave differently under these compressive forces than when they are exposed to internal pressure. This difference in behavior is due to buckling or elastic instability, which make 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. The collapse of a pressure vessel due to external pressure normally starts with small irregularities in either the physical properties or the shape of the shell. A small irregularity in the shell produces localized bending moments. These bending moments tend to emphasize the irregularity or to increase the out-of-roundness of the shell. These effects produce an unstable situation where any surface irregularity is increased by the bending moments that are produced. The critical pressure that causes collapse is not a simple function of the stress that is produced in the shell, as is true with tensile loads. The critical pressure is directly proportional to the material's modulus of elasticity (E) and the shell moment of inertia and is inversely proportional to the cube of the radius of curvature. An ASME Code 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 103 kPa (ga) (15 psig) for full vacuum conditions. The maximum allowable external pressure can be increased by welding circumferential stiffener rings (stiffeners) around the vessel shell. The addition of stiffeners 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 9 illustrates the use of stiffeners on a pressure vessel cylinder.


58



Moment axis of ring h/3

h/3

L

L

L

L

L

L

L

L

L

L h/3

h/3 h = Depth of head

MEX 20203.F9

Figure 9: Stiffener Rings on Pressure Vessel Cylinders

Other factors also affect the design of a pressure vessel for external pressure. The relationship between the modulus of elasticity and unit strain is simple for shells that are at room temperature with an applied stress below the yield point. However, temperature effects must be considered in pressure vessel design. When temperature is considered, the material stress-strain curves are nonlinear with no definite yield point and with a variable modulus of elasticity. The temperature relationship between modulus of elasticity and the stress-strain curve must be expressed as a series of curves based on experimental measurements for particular material types. Therefore, the calculation of allowable external pressure considers the unstiffened length of the vessel component, diameter and thickness, and the stress-strain diagram of the material. Paragraphs UG-28 and UG-33 of the ASME Code contain procedures to calculate the allowable external pressure on cylindrical shells and heads, respectively. These ASME Code external pressure calculation procedures use an iterative approach and are contained in Work Aid 3B. The results of the external pressure calculations must be shown on the Pressure Vessel Design Data Sheet, when applicable.


59



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 determined using the external pressure calculation procedure that is contained in Work Aid 3B.

Shells

The allowable external pressure of a cylindrical shell is a function of material, design temperature, outside diameter, corroded thickness, and unstiffened length. See Work Aid 3B for the ASME Code calculation procedure. Heads

The allowable external pressure of a head is a function of material, design temperature, outside radius, head depth, and corroded thickness. Stiffening rings are not used to increase the allowable external pressure of heads. The head thickness is increased as required to achieve the required external pressure. Since a head may sometimes be installed inside a pressure vessel to separate two chambers, it may be necessary to design the head for an external pressure that is higher than 103 kPa (ga) (15 psig). Work Aid 3B contains the ASME Code procedures to calculate the allowable external pressure for heads. 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 of a conical section may be increased by the addition of stiffener rings to reduce the unstiffened cone length. The allowable external pressure may also be increased by adding to the cone thickness. Work Aid 3B contains the ASME Code procedures to calculate the allowable external pressure for conical sections.


60



Sample Problem 4 External Pressure Calculation

This Sample Problem will provide practice in using the external pressure design procedure for a cylindrical pressure vessel. Use Work Aid 3B to assist in solving this problem. A tall cylindrical tower is being supplied to Saudi Aramco. The geometry and design conditions are specified in Figure 10. The vendor has proposed that the wall thickness of this tower be 7/16 in., and no stiffener rings have been specified. DESIGN INFORMATION Design Pressure = Full Vacuum Design Temperature = 500°F Shell and Head Material is SA-285 Gr. B, Yield Stress = 27 ksi Corrosion Allowance = 0.0625" Cylinder Dimension Shown is Inside Diameter

4'-0"

150'-0"

2:1 Semi-Elliptical (Typical)



61



A.

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. Calculate design length, L, in. L

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

Head Depth = 48/4 = 12 L

= 150 × 12 + 12 × 2/3

L

= 1 808 in.

Calculate outside diameter Do, in. Do

= 48 + 2 × (7/16) = 48.875

48. 875 Do = = 130. 33 (0. 4375 − 0.0625) t L 1 808 = = 37 D o 48.875

Determine the value of A using Work Aid 3B and the calculated Do/t and L/Do. A = 0.000065 (See Figure 11) For the specified material, which figure in Section II of the Code should be used? Figure CS-1. From the value of A and the appropriate temperature curve for the material, what is the maximum permissible external design pressure Pa? Note that A falls to the left of the temperature line (See Figure 12).


62



3


3 4 5 6 789

Do/t = 100 Do/t = 125 Do/t = 150 Do/t = 200

1.2

1.8 1.6 1.4

2.0

2.5

2

00 0 80 1,0 /t = Do

3.5 3.0

/t = Do

5.0

6.0

7.0

8.0

10.0 9.0

20.0 18.0 16.0 14.0

25.0

30.0

35.0

40.0

50.0

12.0

0 0 0 50 60 40 /t = /t = o/t = D Do Do

Do/t = 300

4.0

Do/t = 250

.00001

Do/t = 130

.0001

2

A = 0.000065

Length + Outside Diameter = L/Do

L/Do = 37

20203.f11

GENERAL NOTE:

500˚F

20,000 18,000 16,000 14,000

700˚F

12,000

800˚F

10,000 9,000 8,000

See Table CS-1 for tabular values up to 300˚F

900˚F

7,000

E-29.0 = 106

FACTOR B

Figure 11: Factor A

6,000

E-27.0 = 106

5,000

E-24.5 = 106 E-22.8 = 106

4,000

E-20.8 = 106

3,500 3,000 2,500 2,000

2 .00001

3

4

5

6

7 8 9

2

.0001

A=0.000065

3

4

5

6

7 8 9

2

3

4

5

6

7 8 9

.001

2 .01

3

4

5

6

7 8 9 .1

FACTOR A

20203.V61

Figure 12: Figure CS-1


63



Pa =

2 AE 3(Do / t)

E = 27 × 106 psi from Figure CS-1 (Figure 12) at T = 500°F

Pa =

2 × 0.000065 × 27 × 106 3 × 130. 33

Pa = 9 psi Since the calculated Pa is less than 15 psi, the proposed 7/16 in. shell thickness is not sufficient. 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 1/2 in. D o 48.875 = = 97.75 t 0. 5

L = 37 (as before) Do

A = 0.000114

Pa =

2 × 0.000114 × 27 × 106 3 × 130. 33

Pa = 15.7 psi


64



B.

An alternative to increasing the wall thickness is to add stiffener rings in order to reduce the unstiffened length of shell. This alternative will probably be less expensive than to increase the thickness of the whole shell to 9/16 in. To determine the number of stiffener rings that are required, you must calculate the maximum allowable design length, L. This calculation is done by working the procedure backwards. Do = 130.33 from before. This is based on the originally t specified shell thickness of 7/16 in.

Calculate the required value of B using Pa = 15 psi.

Pa =

B=

4B 3(Do / t)

3Pa (D o / t ) 3 × 15 × 130. 33 = 4 4

B = 1 466 Locate the calculated B in Figure CS-1 (Figure 12). Since B is below the bottom of the chart (i.e., to the left of the temperature lines), the alternative calculation procedure must be used. Calculate A using the following equation.

Pa =

A=


2 AE 3(Do / t)

3Pa (D o / t) 3 × 15 × 130. 33 = 2E 2 × 27 × 106

65



A = 0.0001086 Using the value of A calculated above, and the value of Do/t, go to Figure G in Section II of the Code and determine L/Do. L/Do = 8 L = 8 × 48.875 = 391 in. This value is the maximum acceptable length between stiffeners. It is also the maximum permitted length to the first stiffener from the top or bottom, considering 1/3 of the head depth as part of the length. Total length to stiffen = tangent/tangent length + 2 × 1/3 × (head depth) = 1 808 in. Maximum number of Total length 1 808 = 391 L

spaces

between

stiffeners

=

= 4.62 Rounding up to the nearest whole number = 5 spaces Number of stiffeners, N = (Spaces - 1) =4


66



Flat Covers

The unstayed, circular flat head or cover is another type of head. Unstayed means that the cover is merely a flat plate that does not have any reinforcing bars to strengthen it. The ASME Code contains design rules for flat covers that include items such as minimum required thickness, bolting, and welding requirements. Flat covers are not often used for pressure vessels in refineries and petrochemical plants. However, the flat cover is commonly used as a bolted cover on the channel end of shell and tube heat exchangers. Flat covers will not be discussed further in this course due to their limited applicability to pressure vessels. Flat covers are discussed in MEX 210.

Quick-Opening Closures

Some pressure vessel applications require that the vessel be opened frequently for maintenance or operational reasons and not just during normal T & I’s. The following are examples of such applications: •

Large strainers or filters that are installed in piping systems that must be frequently opened for cleaning.

•

Pressure vessels in batch process operations that must be frequently entered for cleaning.

•

Mixing vessels where additional material is manually added during the process after the vessel has been depressurized.

In these situations where frequent pressure vessel entry is required, it is preferable to have a faster means to open and close the vessel than is provided by a standard bolted flanged connection. A quick-opening closure provides this faster internal access.


67



A quick-opening closure consists of the following main parts (Refer to several standard figures that are contained in Course Handout 3): •

Hub. The hub has an integrally-forged ring that extends beyond its outside diameter at one end. There is a gasketed joint in the face of this ring. The other end of the hub is welded to the nozzle neck where the quick-opening closure will be installed.

•

Dished head. The dished head closes the opening of the pressure vessel nozzle.

•

Gasketed joint. A gasketed joint between the hub ring and dished head provides a pressure-tight seal at the closure.

•

Clamping ring or yoke. This component is split into two pieces across its diameter, clamps the head against the hub ring, and forces the head against the gasketed joint to seal the opening. Loosening the yoke allows the closure to be opened.

•

Bolts. The bolts locate the yoke with respect to the head and hub ring. When the bolts are tightened, the yoke pieces are brought together and tighten the head against the hub ring. When the bolts are loosened, the yoke pieces are moved apart and the head can be moved away from the hub ring. Two bolts are used whatever the closure diameter for most pressure vessel applications.

•

Hinge. The hinge is attached to both the dished head and the hub so that the head can remain supported and be easily swung open and shut.

The closure can be opened and closed quickly because there are usually only two bolts. Opening and closing may be done either manually or by hydraulic or electric operators depending on the application and the size of the closure. It is much simpler to open a quick-opening closure than a standard flanged joint. Consequently, it is also much easier to make a mistake and open the closure when the pressure vessel or piping system is still under pressure. Opening the closure while the system is still under pressure is dangerous.


68



The detailed design of quick-opening closures must meet the normal ASME Code requirements with respect to material selection and dimensions (e.g., head thickness). Paragraph UG-35(b) of the ASME Code contains additional design requirements that focus specifically on quick-opening closures. The ASME Code has these additional requirements because of the potential danger if the closure is not properly operated. Several of these requirements are highlighted below. Refer to Paragraph UG-35(b) for details. •

The closure must have a locking mechanism. The locking mechanism must be designed such that a failure of any one component of the mechanism cannot result in failure of all other locking elements and subsequent release of the closure. One design approach that meets this requirement is to have separate plates bolted between the head and the two halves of the yoke. These plates must be unbolted first before the yoke bolts are used to separate the yoke halves and unbolt the closure.

•

It must be possible to see from external visual observation that the holding elements are in good condition and that their locking elements are in full engagement when the closure is in the closed position.

•

The closure and its holding elements must be fully engaged in their intended operating position before pressure can be built up inside the vessel. This feature ensures that all components of the closure are in the position that they were designed to be in before they are exposed to the pressure loads.

•

If internal pressure would force the closure away from the vessel, the closure must be designed such that the pressure must be fully released before the closure can be fully opened for access. This feature ensures that the closure head will not be rapidly blown back due to a high internal pressure and cause damage or injury.


69



•

Warning devices may be required that will warn the operator if pressure is applied to the vessel before the closure and its holding elements are fully engaged in their intended position. These warning devices also signal if an attempt is made to operate the locking mechanism before releasing the pressure inside the vessel.

Several companies manufacture standard quick-opening closures that meet all ASME Code requirements. Nozzle Reinforcement for Design Pressure

The wall thickness of a nozzle under pressure loading is determined by the same procedures used for cylindrical shell sections. External loadings that are transmitted through a nozzle connection to a pressure vessel, such as by connected piping, may also affect the required nozzle thickness. These external loadings are discussed in a later section of this module. However, pressure loading is normally the main factor and is discussed in detail in this section. Material thickness tolerance is an additional factor that must be considered in nozzle thickness calculations and that is not considered in vessel shell calculations. Pressure vessel nozzles are frequently fabricated using pipe material specifications when the nozzle is a standard pipe size. Pipe material specifications permit the wall thickness that is supplied to be less than the nominal thickness that is ordered by an undertolerance. The permissible undertolerance is stated in the pipe material specification and can be as much as 12.5% of the nominal thickness. Therefore, the undertolerance must be considered when the required wall thickness for a nozzle is calculated. Reinforcement of Pressure Vessel Openings

Calculation of the required wall thickness for a nozzle is one step in the design of openings in pressure vessels. However, 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 specifies design rules that must be followed.


70



The design of pressure vessel openings must address two types of stress conditions. First, the membrane stresses must be kept within allowable limits. Second, localized peak stresses that are caused by abrupt geometric changes at the nozzle-to-shell corner must be kept within allowable limits. Evaluation of these peak stresses is important for the evaluation of cyclic loads that may cause a fatigue failure but is beyond the scope of this course. 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 Rn

tn

trn

2.5t or 2.5 tn + te Use smaller value

t

2.5t or 2.5 tn Use smaller value

;;;

te

tr

c

h

d

d or Rn + tn + t

d or Rn + tn + t

Use larger value

Use larger value

For nozzle wall inserted through the vessel wall

For nozzle wall abutting the vessel wall

Figure 13: Cross-Sectional View of Nozzle Opening


71



When the opening is made, a specific volume of material is removed from the pressure vessel shell or head. Therefore, 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, as shown in Figure 13. The use of this simplification permits the nozzle reinforcement calculations to be made in terms of metal crosssectional area rather than metal volume. The ASME Code design rules state 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 to the opening and be must contained within defined geometric limits in order to provide adequate reinforcement. The replacement metal area may take two forms: •

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

•

Reinforcement that is added to the shell or nozzle.

Figure 14 shows several typical nozzle design configurations. Please note 32-AMSS-004 requires that all nozzles in hydrogen, hydrocarbon, caustic amine, wet sour and steam services shall be attached by welding through the total thickness of the vessel shell or head, including reinforcement.


72



(a-1)

(a) Full Penetration Weld With Integral Reinforcement

(b)

(c)

Separate Reinforcement Plate Added

(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

Figure 14: Typical Nozzle Design Configurations


73



Figure 14 provides examples of inserted versus abutted nozzles, pad reinforcement versus no reinforcement, and selfreinforced nozzles. Note that self-reinforced nozzles are forged fittings that are designed with extra thickness in the nozzle-tovessel junction area in order to provide adequate reinforcement. Work Aid 3C summarizes the ASME Code procedure used to calculate the required nozzle reinforcement. Additional Nozzle Reinforcement

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

A reinforcement pad.

•

Additional thickness in the vessel shell or head around the opening.

•

Additional thickness in the lower part of the nozzle near its attachment to the vessel.

In all cases, the reinforcement must be located within the reinforcement zone boundaries in order for the reinforcement to be considered effective. 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 that the reinforcement pad is attached to. No credit can be taken for the additional strength of any reinforcement that has a higher allowable stress than that of the vessel shell or head material to which it is attached. If a material with a lower allowable stress than the vessel material is used for reinforcement, the reinforcement area must be increased in inverse proportion to the ratio of allowable stress values for the two materials. This increase compensates for the lower allowable stress of the reinforcement material. The reinforcement pad material is normally selected to be the same as the vessel material in order to avoid the need to make this compensation.


74



When a reinforcing pad is used, its thickness is normally set equal to the vessel shell or head nominal thickness, and its required diameter is calculated based on the amount of additional reinforcement area that it must provide. However, any portion of the reinforcing pad that extends outside the boundaries of the reinforcement zone cannot be considered effective. Additional reinforcement must be provided in another manner should it be necessary to extend the reinforcing pad diameter outside the boundaries of the reinforcement zone. It should also be noted that the Pressure Vessel Design Data Sheet requires that nozzle reinforcement not be the factor that limits the maximum allowable working pressure of a pressure vessel. Therefore, a nozzle cannot be the weakest component of a pressure vessel. The ASME Code specifies circumstances under which no nozzle reinforcement evaluations 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, and Participants are referred to the ASME Code for details. The nozzle reinforcement that is determined using the ASME Code procedure considers pressure design only. Vessel nozzle-to-shell intersections must also be adequate for the loads that are imposed by any attached piping or equipment. The ASME Code does not contain specific procedures for the evaluation of these external nozzle loads. The external nozzle loads must be checked by the pressure vessel designer by the use of generally accepted procedures.

Small Connections

SAES-D-001 specifies that connections less than 2 in. size shall be used in utility services only.


75



Sample Problem 5 Nozzle Reinforcement

You are reviewing the nozzle design details that are proposed by a vendor for a new drum and have selected an 8 in. nominal pipe size 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. Using Work Aid 3C, 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. The information that is needed to perform your evaluation is in Figure 15.


76



DESIGN INFORMATION Design Pressure = 300 PSIG Design Temperature = 200°F Shell Material - SA-516 Gr. 60 Nozzle Material - SA-53 Gr. B, Seamless Corrosion Allowance = 0.0625" Vessel is 100% Radiographed Nozzle does not pass through Vessel Weld Seam

8" Nozzle (8.625" OD) 0.5" Thick

0.5625" Thick Shell, 48" Inside Diameter

20203.FIG15



77



tr =

Pr 300 × (24 + 0. 0625) = = 0.487in. SE1 − 0. 6P 15 000 × 1− 0.6 × 300

t rn =

300(3. 8125 + 0.0625) = 0.0784 15 000 × 1− 0.6 × 300

A = dtrF A = (8.625 - 1.0 + 0.125) × 0.487 × 1 A = 3.775 in.2 required area A11 = (Elt - Ftr)d = (0.5625 - 0.0625 - 0.487) × 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 in shell A21

(tn - trn)5t = (0.5 - 0.0625 - 0.0784) × 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 × (0.5 - 0.0625) + 0] = 0.786 in.2

Therefore, A2 = 0.786 in.2 available in nozzle A1 + A2 = 0.1 + 0.786 = 0.886 in.2 Since this value is less than A, the nozzle is not adequately reinforced, and a reinforcement pad is required. A5 = (3.775 - 0.886) = 2.889 in.2 required area in reinforcement pad.


78



te = 0.5625 in. 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 permitted reinforcement limit. 2d = 2 × 7.75 = 15.5 in. Therefore, Dp = 13.761 in. is acceptable

Nozzle Flange Rating

ASME B16.5, Pipe Flanges and Flanged Fittings, provides steel flange dimensional details for standard pipe sizes through 600 mm (24 in.). Standard ASME B16.5 flanges are acceptable for most pressure vessel nozzle flanges 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 this process in general terms. Work Aid 3D provides the specific procedure to follow. Flange material specifications are listed in Table 1A in ASME B16.5, a portion of which is excerpted as Figure 16. A copy of ASME B16.5 is contained in Course Handout 1. The material specifications are grouped within specific Material Group Numbers. The process for determining the Material Group Number is contained in Work Aid 3D. 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.


79



TABLE 1A LIST OF MATERIAL SPECIFICATIONS Material Groups Product Forms Material Nominal Group Designation No. Steel Forgings Castings Plates Spec. Spec. Spec. No. Grade Notes No. Grade Notes No. Grade Notes 1.1 Carbon A105 -(1)(2) A216 WCB (1) A515 70 (1) A350 LF2 ----A516 70 (1) C-Mn-Si ------A537 Cl.1 -1.2 Carbon ---A216 WCC (1) ------A352 LCC ----2 1/2 Ni ---A352 LC2 -A203 B -3 1/2 Ni A350 LF3 -A352 LC3 -A203 E --

Figure 16: ASME B16.5, Table 1a, Material Specification List (Excerpt)

Table 2 of ASME B16.5 provides the information that is necessary to select the flange Class that is appropriate for the specified design conditions. The Class accounts for the required flange design temperature and pressure. ASME B16.5 has 7 classes, designated as Class 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. A higher flange Class can withstand higher pressure and temperature combinations. Figure 17 is an excerpt from Table 2 and shows the temperature and pressure ratings for carbon steel Material Groups in ASME Class 150. Refer to the copy of ASME B16.5 in Course Handout 1 for information on other Material Groups in Class 150 and on the other flange Classes.


80



Material Group No.

1.1

Temp., °F

1.2

1.3

1.4

Carbon Steel

- 20 to 100 200 300 400

285 260 230 200

290 260 230 200

265 250 230 200

235 215 210 200

500 600 650 700

170 140 125 110

170 140 125 110

170 140 125 110

170 140 125 110

750 800 850 900

95 80 65 50

95 80 65 50

95 80 65 50

95 80 65 50

950 1000

35 20

35 20

35 20

35 20

Figure 17: ASME/ANSI B16.5, Class 150, Pressure-Temperature Ratings (Excerpt)


81



The Material Group Number is read across the top of Table 2. Flange design temperature is read down the left side of the table. The numbers in the table are maximum allowable flange design pressures for a particular combination of flange material (as defined by Material Group Number) and design temperature. Allowable design pressure decreases as the design temperature increases. The dimensions of a standard ASME B16.5 flange are fixed for a given Class and pipe size. The decrease in allowable pressure as the temperature increases ensures that the standard flange design with fixed dimensions will not fail due to the reduction in material strength at higher temperature. Material and design temperature combinations that do not have a pressure indicated are not acceptable. As previously noted, the strength of a flange increases as the Class number increases for a given Material Group. This strength increase is accomplished by increasing flange dimensions, such as thickness; this increase results in a more substantial structure. The increase in flange dimensions requires that more steel be used, and individual flange cost then increases. Therefore, in selecting flange Class, the lowest acceptable Class that is suitable for the design conditions should be used in order to minimize cost. Specification of the size, material, and Class completes most of the selection requirements for flanges that are covered by ASME B16.5. Flange type and gasket material must also be specified. However, discussion of these factors, as well as the design of flanges that are outside the size limits of ASME B16.5, is beyond the scope of this course. Several specific flange design requirements are stated in 32SAMSS-004. For example, slip-on type flanges cannot be used in the following cases: •

Severe cyclic conditions

•

Hydrocarbon or sour services

•

Design temperatures over 232°C (450°F)


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In addition, flange faces must have a 125-250 microinch machine finish if spiral wound gaskets are used. This finish is a relatively smooth finish for a raised face and increases the probability that a tight seal is maintained with a spiral wound gasket.

Maximum Allowable Working Pressure (MAWP)

The maximum allowable working pressure (MAWP) of a pressure vessel is the maximum permissible gauge pressure in the vessel and is expressed as the lowest MAWP of all the vessel's components. 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" that is coincident with the MAWP. The MAWP is based on calculations that are made for every component of a pressure vessel and that are based on the actual supplied thicknesses of the material. The material thicknesses that are used in these calculations do not include any excess thickness that was added for corrosion allowance or to absorb loadings other than pressure. Each component must be checked to determine which component limits the pressure vessel MAWP. The MAWP is shown on the Pressure Vessel Design Data Sheet (and the vessel nameplate) after the vendor has completed his detailed design and knows the component thicknesses that he will use. The MAWP is always greater than or equal to the design pressure, also because the "designated temperature" equals the design temperature. The component thicknesses that are actually supplied will frequently be slightly greater than the minimum thicknesses that are required for the design conditions because standard available plate thicknesses are typically used. Pressure vessels are normally designed so that the main shell or heads are the components that limit the MAWP, rather than other components such as nozzle intersections. As previously stated, Saudi Aramco requires that the vessel shell or heads govern the MAWP.


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The MAWP may be used later if a change in operation is being considered that requires a more severe design pressure and/or temperature. The MAWP shows whether the same pressure vessel may be used at the new design conditions or whether a new vessel must be purchased. The MAWP is normally calculated for two conditions: •

New and cold

•

Design

The new and cold calculation is based on ambient temperature and new component thicknesses. The design calculation is made for the design temperature and corroded component thicknesses. Calculation of the MAWP for both conditions provides an exact range that the vessel can withstand. From a practical standpoint, the design MAWP is the more meaningful calculation, and this is the value that is shown on the Pressure Vessel Design Data Sheet. If a vessel is re-rated in accordance with the MAWP, its safety valve set pressure must also be adjusted. A new hydrotest based on the MAWP and revised ASME Code documentation is also typically required if the MAWP was not known during the original vessel design. However, since Saudi Aramco requires that the MAWP be determined during the initial design, the initial vessel hydrotest pressure is based on the MAWP. Calculation of the required hydrotest pressure is discussed in MEX 202.04. The ASME Code requirements and equations for calculation of the minimum required wall thicknesses for various pressure vessel components were discussed in earlier sections of this module. In all these cases, the design pressure, P, is assumed as given information and the minimum required wall thickness, t, is calculated. The MAWP is calculated with the same equations by reordering them so that "P" is calculated as a function of "t". In this case, P = MAWP, and "t" is the actual supplied component thickness (less corrosion allowance and any excess thickness provided for reasons other than design pressure). The equations that are used to calculate MAWP are shown in Work Aid 3A.


84



Stresses From Local Loads Applied to Nozzle and Attachments

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 in the pressure vessel shell or head that are in addition to the stresses that are caused by pressure, weight, and wind loads. The sources of the external loads may be the following: •

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

•

Loads from platforms, internal or external piping, or equipment that are supported from a pressure vessel shell by lugs or clips that are attached to the shell.

•

Loads at vessel supports, such as columns or lugs.

The following discussion focuses on loads that are applied at a pressure vessel nozzle, although the general approach also applies to loads that are transmitted to a vessel shell or head through a lug-type attachment. In the discussion that follows it is assumed that: •

The shell has been properly designed for such factors as internal pressure, weight, and wind.

•

Any nozzle opening in the shell has been adequately reinforced by an area replacement method such as that discussed earlier in this module.

•

The attachment is not located near another geometric discontinuity in the vessel.

•

The nozzle or attachment is not subjected to high thermal gradients, and the applied loadings are not cyclic.

•

The nozzle neck itself has also been adequately designed for the external loads, in addition to the local stresses at the point of attachment.


85



Allowable Stress Bases

The combination of the additional local stresses that are caused by external loads with the general stresses in the vessel shell, such as those due to internal pressure, must be kept within allowable limits. A typical and conservative approach for vessels that are designed to Division 1 is to use Division 1 allowable stresses with a Division 2 stress categorization. In Division 2, the stresses in the shell near an opening are considered as local membrane (PL) stresses, and the bending stresses are considered as secondary (Q) stresses. Division 2 also classifies the stresses that are associated with the corner or fillets as peak (F) stresses. Division 2 limits the local membrane stress to 1.5 Sm and limits the sum of local membrane and bending stresses to 3 Sm. Peak stresses are stresses that can cause fatigue failure if the stresses are highly cyclic. Higher allowable stresses are permitted since the stresses that are produced by the external loads are very local to the junction and die out rapidly away from the junction. For a Division 1 vessel, the Division 1 allowable stress, S, is typically substituted for Sm, the Division 2 allowable membrane stress. If the internal pressure or external loads are highly cyclic (over 400 full range cycles per Division 2), a fatigue analysis should be performed. For a fatigue analysis, the local membrane and bending stresses in the shell/nozzle due to the cyclic loads should be multiplied by stress concentration factors. The fatigue analysis is typically done in accordance with Appendix 5 of Division 2. It should be noted that while discussion of fatigue analysis is outside the scope of this course the first step in such an analysis is to ensure that the local stresses without stress concentration factors do not exceed the allowable limits for local membrane and secondary stresses.


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Calculation Procedures

A generally accepted approach for the calculation of the stresses from external loads is contained in the Welding Research Council Bulletins No. 107 and No. 297. WRC 107, Local Stresses in Spherical and Cylindrical Shells Due to External Loadings, was first published in 1965 and has been revised/corrected and reprinted several times. Therefore, it is important that the latest revision (March 1979) and printing (1989) be used. WRC 107 covers solid and hollow and square and round attachments to spherical shells and solid rectangular and round attachments to cylindrical shells. WRC 297, Local Stresses in Cylindrical Shells Due to External Loadings on Nozzles - Supplement to WRC Bulletin No. 107 was first published in 1985. Revision 1 of WRC 297 was published in September 1987. WRC 297 gives a method for the calculation of local stresses due to external loads on hollow cylindrical attachments to cylindrical shells. Therefore, WRC 297 is more appropriate to use for nozzles into cylindrical shells than WRC 107. It should be noted that, although the two bulletins appear to be similar, there are subtle differences in nomenclature and sign convention. WRC 297 also covers a wider range of nozzle and shell geometries than WRC 107. Both bulletins provide a method to calculate stresses on the upper (outside) and lower (inside) surfaces of the vessel shell at four points around the intersection of the attachment to the shell. The methods are complicated since the external loads can be either positive or negative and may contribute either a positive (tensile) or negative (compressive) membrane stress component and a positive or negative bending stress component. The signs of the stress components depend on the location of the point on the shell. Another difference between WRC 297 and WRC 107 is that WRC 297 includes a method to calculate the local stresses in the nozzle neck itself in addition to those in the shell. In general, if the nozzle neck is thinner than the shell, and if the local stresses in the shell as calculated by the WRC 107 procedure are at or near their allowable limits, the stresses in the nozzle neck should also be checked using the WRC 297 procedure.


87



WRC 107 and 297 can be used for a wide range of geometries, but they do not cover all possible geometries. The accuracy of the methods is also a function of the geometry. Therefore, extrapolation of the methods beyond their geometric limits is not recommended. For nonstandard situations, a finite element analysis should be considered. The following paragraphs briefly summarize the WRC 107 calculation approach that is applicable to a nozzle on a hemispherical head. The actual calculations involve the determination of the following: •

Several geometric parameters.

•

Stress coefficients found from multiple graphs that are functions of these geometric parameters.

•

Multiple stress components that are found using the applied loads and the stress coefficients.

•

The stresses resulting from combination of the appropriate stress components.

From a practical standpoint, these calculations are now done using readily available computer programs rather than by tedious hand calculations. These computer programs determine all the required parameters, coefficients, and resultant stresses based on geometry and load information which is entered as input. Saudi Aramco will typically use the CODECALC computer program for these vessel shell stress calculations. The piping loads that are applied to the nozzles will typically be found using the Simflex II piping analysis computer program. Work Aid 3E summarizes the approach to use in the evaluation of vessel stresses that result from applied piping loads.


88



Since the calculation procedure is very similar but the nomenclature and sign convention are different, the calculation procedure will be discussed using the terminology and sign conventions of WRC 107. In the case of a nozzle on a hemispherical head, which is covered by WRC 107, the following loads can be considered:

External Loads on Nozzle Attached to Hemispherical Head -

•

Radial force, P

•

Bending moments, M1 & M2

•

Transverse shear forces, V1 & V2

•

Torsional Moment, Mt

Note that, in the case of a hemispherical head, the maximum stress will occur in the plane of maximum moment. Therefore, moments M1 and M2 should be combined vectorially to give the maximum moment, M, and V1 and V2 resolved into two components in the plane of maximum moment and perpendicular to the plane. In the case of ellipsoidal or torispherical heads, the WRC 107 procedures for a hemispherical head can be used if the mean radius is assumed to be equal to the crown radius of the head. Figure 29, contained in Work Aid 3E, is based on WRC 107 sign conventions and nomenclature and shows the positive direction for external loads applied to a nozzle in a hemispherical head. Also shown are the directions of positive stresses on the top surface of the shell. The figure also shows the locations of four points around the nozzle-to-shell intersection at which stresses will be calculated on the upper and lower surfaces. This figure should be referred to in conjunction with the following discussion.


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Shell and Nozzle Attachment Parameters

In both WRC bulletins, various shell and attachment parameters must be calculated. In the case of a round hollow nozzle into a hemispherical head that is covered by WRC 107, the parameters are as follows: •

Shell parameter, U:

U = ro/(RmT)0.5

•

Attachment parameter, γ:

γ = rm/t

•

Attachment parameter, ρ:

ρ =T/t

Where: Rm

= Mean radius of spherical shell, mm (in.)

ro

= Outside radius of the nozzle, mm (in.)

T

= Thickness of the spherical shell, mm (in.)

t

= Thickness of the nozzle, mm (in.)

The applicable WRC bulletin should be consulted for other types of attachment configurations. The next step is to determine the stress coefficients, which are the values of the stress resultant functions for a given set of attachment parameters. These stress coefficients are obtained from graphs that are in WRC 107. In the case of a round hollow nozzle into a hemispherical head, eight stress coefficients are required.

Stress Coefficients and Local Stress Components -

Local stress components are then determined by multiplying the stress coefficients by the external loads and appropriate shell parameters. It is important that the algebraic sign of these stress components be changed as needed, based on the direction of the applied load. For comparison with the allowable stress basis, the membrane stress components due to pressure should be added algebraically, and then the combined membrane plus bending stress should be calculated.


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An average shear stress is then calculated due to the torsional moment, Mt. The transverse shear forces, V1 and V2, may also be calculated. However, the shear stress will normally not be a significant design consideration. The preceding discussion summarized the procedures that are used to determine the local stresses around a cylindrical nozzle attachment into a hemispherical head. Similar procedures are used for other types of attachments to hemispherical or cylindrical shells, such as rectangular or square attachments for support lugs. Reference should be made directly to WRC 107 or 297 for details covering these items. The actual stress calculations are done using computer programs such as CODECALC. The following paragraphs discuss additional considerations.

Additional Design Considerations -

If the external loads are cyclic, they may cause a fatigue failure of the nozzle or attachment. If a fatigue analysis is performed, stress concentration factors should be used to account for the contour of the nozzles and welds. Appendix B in WRC 107 gives some guidance on stress concentration factors. The ASME Code Section VIII, Division 2 Appendix 4 Article 4-6 gives guidance on stress concentration factors to be used for pressure. Note that stress concentration factors are sometimes applied to the entire load to simplify the calculations, although they need only be applied to the cyclic portion of the load. In many instances, the nozzle reinforcement that is used for internal pressure is also adequate for the externally applied loads. In other cases, the shell must be further reinforced for the external loading. Since the local stresses in the junction reduce and distribute themselves away from the nozzle or attachment, local reinforcement by means of a thickened insert plate or a pad is used. This reinforcement complicates the analysis in that the stress functions depend on geometry, and an iterative solution is required. Typically, the pad plus shell thickness, or insert plate thickness, is determined assuming that the stress functions are unchanged. Then new stress coefficients are determined for the assumed thickness, and the procedure is repeated until the change in thickness is negligible.


91



If a pad is used to reinforce the shell, the pad thickness should generally be between 0.75 and 1.5 times the thickness of the shell to which it is attached. The pads should be continuously welded to the nozzle neck and shell with full penetration and full fillet welds, respectively. If local reinforcement is used, the extent of the reinforcement along the shell must also be determined. One method is to assume that the pad is a rigid plug and to determine the diameter of the plug such that the stress in the shell at the edge of the plug is acceptable. The maximum width of the reinforcement pad should not exceed 16 times the lesser of the pad or shell thickness; otherwise, the pad or shell may fail by buckling due to large bending stresses. If the nozzle neck is thinner than the shell, the stresses in the nozzle neck should be determined by a procedure such as that given in WRC 297. Typically, the thickness of the nozzle neck would be increased as required; however, in some cases pad reinforcement can be used. Since the applied bending moments dampen out away from the nozzle, the extent of the reinforcement along the nozzle neck should be at least 1.56(rt)0.5. Finally, it should be noted that not all nozzles or attachments are analyzed to the same level of detail. If the nozzle or attachment is lightly loaded, and if the thickness of the shell is greater than the nozzle neck thickness, the shell stresses are probably acceptable, and no analysis need be done. On the other hand, if the nozzle or attachment is of primary importance, or if the geometry is not typical, a computerized finite element analysis may be necessary.


92



Sample Problem 6 Evaluation of Stresses from Local Loads Applied to Nozzles and Attachments

Using the procedure that is provided in Work Aid 3E, determine if the stresses in a pressure vessel spherical head under applied piping loads are acceptable. The spherical head geometric information and applied loads are shown in Figure 18.

P = -12253 lb. MT = 24000 in. -lb.

V2 = 1000 lb. M2 = 24000 in. - lb. rm = 7.75 in.

V1 = 1000 lb. M1 = 24000 in. - lb.

0.5 in.

Nozzle B

C

A

Spherical Head

D 0.5 in.

Rm = 227.75 in.



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The head material is SA-516 Gr. 70, the vessel design temperature is 400°F, and the design pressure is 75 psig. The Division 1 allowable stress in tension, S, is 17 500 psi. Therefore, 1.5 S 3S

= 26 250 psi, allowable membrane stress intensity; and = 52 500 psi, allowable membrane plus bending stress intensity.

The contractor used a computer program, such as CODECALC, to calculate the stresses in the head in accordance with WRC 107. The results from the computer calculations are as follows: Maximum membrane stress intensity = 24 059 psi. Maximum membrane plus bending stress intensity = 45 332 psi. Both of these calculated stresses are lower than their respective allowable values. Therefore, the head design is acceptable for the applied loads.


94



EVALUATING THE CONTRACTOR-SPECIFIED DESIGN OF PRESSURE VESSEL SUPPORTS

The most common types of pressure vessel supports are the following: •

Skirts

•

Columns

•

Saddles

;;;

Figure 19 shows these types, all of which are used at Saudi Aramco facilities.

MEX 20203.F19

Column Support

;

Skirt Support

Saddle Support

Figure 19: Types of Pressure Vessel Supports


95



Most vertical pressure vessels are supported by skirts. Skirts are an economical design because they generally transfer the loads from the vessel by shear action. They also transfer the loads to the foundation through anchor bolts and bearing plates. Column-supported vessels are normally relatively small and light in weight. The columns permit easy access under the vessel. Figure 19 shows a column support design where the columns attach directly to the vessel and the loads are transferred by shear action. Column supports are also often used for pressurized storage vessels. Cross-bracing of the columns may be necessary to minimize lateral and torsional movements under wind or earthquake loadings. An alternate design has the support columns attached to lugs and the lugs welded to the vessel shell. The bending stiffness of the shell and its ability to adequately resist the moments must be considered in the lug support design. Horizontal vessels are normally supported by saddles. Circumferential stiffener rings may be required at saddle supports if the vessel shell is too thin to transfer the loads to the saddles. Axial thermal expansion of a horizontal vessel is normally handled by anchoring one saddle support to the foundation and letting the other saddle support move freely. The weight load of the pressure vessel must be calculated to permit correct design of the vessel support system. Weight load calculation is necessary regardless of the type of support that is used or regardless of the vessel geometry. Wind and earthquake loads must also be determined in order to design the support system. Wind and earthquake loads are especially important in the design of vertical pressure vessel supports since the bending moments that are applied at the support increase as the vessel height increases. The calculation of wind and earthquake loads that are applied to pressure vessels was discussed earlier and in CSE 110.


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Vertical Vessel

Small and medium-sized vertical vessels (1.8 - 3.6 m [6 - 12 ft.]) that are located on the ground are usually supported on uniformly spaced column supports. If a short vessel is located above ground on structural steel construction, or if a short vessel is connected by piping to a reciprocating pump or compressor, it is sometimes supported by a skirt to avoid vibration problems. Larger vertical vessels are typically supported by skirts. Once the required weight, wind, and earthquake loads have been determined, the design of vertical pressure vessel supports includes the following: •

Determine the required support size. For columns, this determination involves selection of the appropriate standard structural steel member or pipe size which is necessary for the applied load. For skirts, this determination involves determination of the appropriate skirt diameter and thickness. The height of the support column or skirt is normally set by process design conditions and considers fluid head and flow requirements.

•

Determine the need for and sizing of cross-bracing for column supports.

•

Calculate the stresses in the vessel shell at the support attachment points and determine if these stresses are acceptable. Determine if additional reinforcement at the shell is needed to keep the stresses within acceptable limits.

•

Determine appropriate details to use at the support base to the foundation. These details include baseplate dimensions, anchor chair details, and anchor bolt number, diameter, and location.

•

Determine appropriate weld details to use at the pressure vessel and baseplate attachment points.


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•

Determine if other specific vessel design considerations would affect the support design. These considerations might include the calculation and evaluation of localized stresses that are caused by high temperature applications or cyclic loads.

The sections that follow concentrate on the requirements for the basic design of support columns and skirts. Discussion of the other items is beyond the scope of this course.

Column Supports

Column support design considers axial weight loads, bending moment, and shear forces in the vessel. Vessel design pressure is not a consideration since the column supports are not exposed to operating pressures. Figure 20 shows the forces and bending moment expressed as W, V, and M. The axial weight force, W, is carried uniformly by all columns. The shearing forces, V, are carried by the columns that are closest to the neutral axis (Columns B). The bending moment, M, is carried by the columns that are located away from the neutral axis (Columns A).


98



B

A

A

W V M

B

A

A Section A-A

;;;

W

M

V

20203.F20

Figure 20: Column Support Loads


99



The shearing force, VB, at the top of column B causes a bending moment in the column if no cross-bracing is used. With crossbracing, VB is resolved into components. One component of VB acts along the axis of the cross-bracing, another acts along the column axis, and a third acts radially on the shell. The design of cross-bracing will not be discussed in this course. Selection of the appropriate column size, number, and position for a particular situation is a trial and error process. The following Sample Problem, in conjunction with Work Aid 4A, will be used to demonstrate the general approach to the design of column supports.

Sample Problem 7 Design of Column Supports

Column supports must be designed for the vertical pressure vessel that is shown in Figure 21. The following additional information is available: Wo = 92 000 lb., vessel operating weight WT = 208 000 lb., vessel hydrotest weight Wc = 6 000 lb., vessel empty weight Po

= 2 000 lb., lateral wind force during operation (based on 85 MPH wind velocity)

PT

= 720 lb., lateral wind force during hydrotest (based on 51 MPH wind velocity, 60% of design)

Assume that four support columns will be sufficient and that they are SA-36 carbon steel (36 000 psi yield stress). The discussion above covers Steps 1 and 2 in Work Aid 4A, plus some information about the columns. The following continues with Step 3.


100



D = 48 in.

A

t = 0.5 in.

x

P e

192 in.

Direction of wind or earthquake

P1

P2

H = 96 in.

Db y

y P

T.L.

F x L = 72 in. 3 L 4 a

A a

Section a-a

C L column

MEX 20203.F21



101



3.

Calculate the bending moments at the column base and the vessel tangent line, Mb and Ma respectively. This calculation is done for both the operating and hydrotest cases. H

= 96 in.

L

= 72 in.

Mb

= P(H + L)

Ma

= PH

Operating Case Mbo Mao

= 2 000 (96 + 72) = 336,000 in.-lb. = 2 000 × 96 = 192 000 in.-lb.

Hydrotest Case MbT MaT 4.

= 720 (96 + 72) = 120 960 in.-lb. = 720 × 96 = 69 120 in.-lb.

At this point, assumptions must be made for the column design to be used. It was already assumed that four columns could be used. This assumption is applicable to most small, vertical pressure vessels. Next, a standard column section must be selected. An iterative process is then used to ensure that the column is not overloaded and to optimize the design. For the purpose of this discussion, we will only go through the process once to illustrate the overall approach.


102



Assume that the columns are fabricated using a W8 × 31 standard "wide-flange" beam. This beam has the following properties (obtained from a standard civil engineering design manual, or from an equivalent source). A

= 9.12 in.2

Lx

= 110 in.4

Ly

= 37 in.4

Zx

= 27.4 in.3

Zy

= 9.24 in.3

rx

= 3.47 in.

ry

= 2.01 in.

Depth = 8 in. Therefore, r = 2.01 in. (least radius of gyration) e = 4 in. (half of the depth for column orientation used) I = Ix = 110 in.4 (x-direction is perpendicular to wind direction for this column orientation) ∑ I1 = 2(Ix + Iy) = 2(110 + 37) = 294 in.4 5.

Determine allowable compressive stress. K1L 1.5 × 72 = = 53.73 2.01 r

Fa = 18 014 psi, from Figure 32 6.

Determine allowable bending stress. Fb = 0.6 × 36 000 = 21 600 psi


103



7.

Calculate the maximum axial compressive load on the leeward side of each column. See Figure 30 in Work Aid 4A. Db = Vessel outside diameter + 2 × (Distance between shell and column centroid) Db = (48 + 2 × 0.5) + 2 × 4 Db = 57 in. Operating Case Co =

W o 4Mbo 92 000 4 × 336 000 + = + 4 4 × 57 N NDb

Co = 28 895 lb. Hydrotest Case CT =

WT 4MbT 208 000 4 × 120 960 + = + N NDb 4 × 57 4

CT = 54 122 lb. 8.

Calculate the maximum total axial uplift load on the windward side of each column. This calculation is done for the operating conditions and with the vessel empty. See Figure 30 in Work Aid 4A. Operating Case To = −

W o 4Mbo 92 000 4 × 336 000 + =− + 4 4 × 57 N NDb

To = -17 105 lb. Since this value is negative, it indicates that the operating vessel weight will overcome the wind load that tends to overturn the vessel. Empty Vessel Case Tc = −


W e 4Mbo 6 000 4 × 336 000 + =− + 4 4 × 57 N ND b

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Tc = 4 395 lb. This result indicates that the wind would tend to overturn the vessel if it is empty. The columns must be bolted down to prevent this occurrence, which is always the case. 9.

The eccentric loads at the top of the columns are then calculated. See Figure 30 in Work Aid 4A. Operating Case P10 =

Wo 4Mao 92 000 4 × 192 000 + = + N ND 4 × 48 4

P10 = 27 000 lb. P 20 =

4Mao W o 4 × 192 000 92 000 − = − ND N 4 × 48 4

P20 = -19 000 lb. Hydrotest Case WT 4M aT 208 000 4 × 69 120 + = + 4 N ND 4 × 48

PIT =

PIT = 53 440 lb. Empty Vessel Case P 2E =

4Mao W e 4 × 192 000 6 000 − = − 4 × 48 ND N 4

P2E = 2 500 lb. 10. Calculate the lateral force per column for both the design and test cases. See Figure 30 in Work Aid 4A. Operating Case Fo =

PoI 2 000 × 110 = = 748 lb. 294 ΣI

Hydrotest Case P I 720 × 110 FT = T = = 269 lb. SI 294 Saudi Aramco DeskTop Standards

105



11. The axial and bending stresses in the column are now compared to allowable values. This comparison is done for both the operating and test conditions. Note that since the tensile loads in the column that were calculated in Step 8 will typically be lower than the compressive loads from Step 7, they will not need to be checked. Also, the allowable tensile stress for the columns (typically 0.67 Fy) will typically be higher than the allowable compressive stress. a.

Operating Case Axial Compression f ao =

C o 28 895 = = 3 168 psi A 9.12

Bending P10e 0.75F oL + Zx Zx 27 000 × 4 0.75 × 748 × 72 = + = 5 416 psi 27.4 27.4

f bo =

b.

Hydrotest Case Axial Compression faT =

C1 54 122 = = 5 934 psi 9.12 A

Bending f bT =

W T e 0.75F TL P1T e + + NZ x Zx Zx

fbT =

208 000 × 4 0.75 × 269 × 72 53 440 × 4 + + 4 × 27.4 27.4 27.4

fbT = 15 923 psi


106



c.

Combine the axial and bending stresses and compare to the allowable values. Operating Case fo = fao + fbo = 3 168 + 5 416 = 8 584 psi Hydrotest Case fT = faT + fbT = 5 934 + 15 923 = 21 857 psi Therefore, f = fT = 21 857 psi The hydrotest case governs the column design. Fa = 18 014 psi and Fb = 21 600 psi Since f exceeds the lower of Fa or Fb, we must proceed to Step 12 to further evaluate the proposed design.

5 934 faT = = 0.33, which is greater than 0.15 Fa 18 014

12.

Fe ' =

12π 2E 12 π 2 × 29 × 106 = = 51 724 2 23 × 53.732  K1L  23  r 

0.85fbT 5 934 0.85 × 15 923 faT + = + = 1.04 18 014  5 934  Fa  faT  1− Fb 21 600 1−  Fe '   51724 

This value exceeds 1.0, and therefore, the selected beam section does not meet the design requirements. Using the next heavier 8-inch “W” section should satisfy the strength requirements for this vessel. The student is left to verify this. 13. The local stresses in the shell due to the column load would then be evaluated, and the shell will be locally reinforced as needed.


107



Skirt Supports

The design pressure of the vessel need not be considered in the design of a skirt support because the skirt is not exposed to the operating pressure. The allowable stress that is specified by the ASME Code for the skirt material need not be used because the skirt is an external attachment and is not part of the pressure vessel itself. The local building codes or civil engineering standards usually specify the maximum allowable tensile and compressive stresses for a steel support structure such as a skirt. These may be used in skirt design. However, from a practical standpoint, ASME Code allowable tensile and compressive stresses are normally used for skirt design. Since the skirt is not designed for pressure, it would appear that the skirt thickness should be less than the vessel shell thickness at the attachment point if the same material is used for both components. However, the skirt tends to absorb greater bending moments that are caused by wind or earthquake loads. These increased bending moments may require a greater skirt thickness. Therefore, the skirt thickness is often the same as the thickness of the bottom portion of the vessel shell. Failure of a cylindrical skirt under axial compressive stresses may occur due to axial column buckling or local wrinkling if the skirt is not properly designed. This type of failure is the same as overloading a structural column in compression. As with the vessel shell itself, failure is more likely to occur due to columnwrinkling that is produced by excessive combined axial compressive loads. These axial compressive loads are caused by weight, plus either wind or earthquake. Paragraph UG-23 of the ASME Code limits the maximum compressive stress to prevent failure (as was discussed earlier in this module). This ASME code allowable compressive stress basis should be followed for skirt design. Figure 22 shows support skirts that are welded directly to the vessel bottom head or shell. A Type 1 skirt may be either straight or flared and is butt-welded to the knuckle portion of the head. A Type 2 skirt may be either straight or flared and is lapwelded to the cylindrical portion of the shell. The type of weld attachment that is used between the skirt and vessel determines the weld joint efficiency which must be used in the skirt design calculations.


108



Type 1: Butt weld blends smoothly into head contour

;;; ;;; ;;; ;;; Straight Type 1a

15°MAX

Flared Type 1b

Type 2: Lap weld blends smoothly into shell contour

;;; ;;; ;;;;;; Straight Type 2a

Flared Type 2b

MEX 20203.F18

Figure 22: Types of Support Skirts


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Skirt Type 1a is most often used for tall vessels. The centerlines of the cylindrical skirt plate and the corroded shell plate are approximately coincident. 32-SAMSS-004 requires that this detail be used for all but hemispherical heads. If the skirt plate is thicker than the bottom shell plate, the outside diameter of the skirt is made equal to the outside diameter of the bottom shell. If the uplift caused by the imposed external moment is high, and if the anchor bolt spacing becomes too small for the required bolt size, the skirt is designed as a Type 1b. The skirt flare increases the skirt diameter at the base plate and permits the use of larger diameter and/or more anchor bolts, as required. The skirt flare also increases the skirt section modulus in going from the attachment point to the base, which makes it more resistant to the applied bending moment. The Type 2a skirt is attached to the flanged portion of the bottom head in such a way that it does not obstruct any required inspection of the head-to-shell junction weld seam. The Type 2a skirt is more difficult to fabricate and is used mainly in situations that involve high external loads, high design temperatures, or cyclic operating temperatures. A good fit between the outside diameter of the shell and the inside diameter of the skirt is essential. A flared Type 2b skirt is used for the same reasons as a Type 1b skirt. Work Aid 4B contains the procedures to use for skirt design.


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Horizontal Vessel Saddle Supports

Figure 23 shows a typical horizontal vessel on two saddle supports.

Less effective portion of unstiffened shell a

R b θ

A b= R= θ= A=

Width of saddle Radius of shell Saddle contact angle Distance between vessel tangent line and centerline of saddle

a

Section a - a

MEX 20203.F23

Figure 23: Horizontal Vessel on Saddle Supports


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The required cylindrical shell and head thicknesses are generally governed by the membrane stress that is due to pressure and are calculated using the ASME Code design equations that were previously discussed. However, the design of a horizontal vessel that is supported on saddles must proceed further through the use of procedures that are not contained in the Code. The paragraphs that follow highlight the details that must be considered. The actual design calculations for a horizontal vessel on saddle supports will typically be done with a computer program such as CODECALC, which is used by Saudi Aramco. The most common horizontal vessel support design uses two saddle supports that are located an equal distance from the vessel midpoint. With this support configuration, the load that results from the weight of the vessel and its contents will be equally divided between the two supports, even if one support should eventually settle more than the other. If more than two supports are used, the load may not be equally divided among the supports after settlement occurs. A horizontal vessel on two saddle supports is analyzed as a uniformly loaded beam that is simply supported. The uniform weight load produces longitudinal bending stresses in the shell at mid-span and above the saddle supports. These longitudinal bending stresses are combined with the longitudinal pressure stress and are kept below the Code allowable stress. A complication occurs at the saddle location because high bending moments occur at the location where the saddle attachment stops along the shell circumference. This location is called the "horn" of the saddle. These high localized bending moments cause localized shell deformation and reduce the ability of the shell to effectively absorb bending. This localized shell deformation must be accounted for in the calculations that are made at the horn of the saddle. Figure 23 shows the zone above each saddle support where the shell is not completely effective. If the stresses in the saddle area are excessive, a modified saddle design is required. Saddle design modifications may include the following actions: •

Increasing the width of the saddles.


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•

Increasing the contact angle between the saddles and the shell.

•

Welding circumferential stiffener rings to the vessel shell.

If stiffener rings are used, they must be located either in the plane of the saddle and welded both to the saddle and vessel shell, or they must be located on both sides of the saddle and welded only to the shell. The addition of stiffeners prevents local deformation of the shell and makes the entire shell section effective in resisting the local bending moment. The ring stiffeners must be strong enough to prevent shell deformation, without being overstressed themselves or allowing the shell to become overstressed. The stiffeners may be fabricated from plate or standard structural sections, whichever is most appropriate for the specific design loads. Figure 24 shows stiffener rings located at saddle supports.

Ring stiffener Vessel shell

Saddle support

Single stiffener

Two stiffeners

MEX 20203.F24

Figure 24: Stiffener Rings at Saddle Supports


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The design of a saddle support system for a horizontal vessel is a complex process that requires the calculation of several different stresses at various locations in the vessel. The paragraphs that follow provide an overview of the overall procedure and of the stresses which must be calculated. Design of Horizontal Cylindrical Vessels on Saddle Supports

As with any other vessel, horizontal vessels on saddle supports are designed for specified internal and/or external pressure. With regard to weight loading, the vessel behaves like a beam resting on supports. L. P. Zick developed an analysis procedure for calculating the stresses that are induced in the cylindrical vessel shell due to the weight loads ("Stresses in Large Horizontal Cylindrical Pressure Vessels on Two Saddle Supports," The Welding Journal Research Supplement, 1971). This procedure calculates and evaluates the following stresses: •

Maximum longitudinal bending stress.

•

Tangential shear stress.

•

Circumferential stress at the horn of the saddle (that is, where the saddle is welded to the vessel shell).

•

Ring compression in the shell over the saddle.

•

Additional stress in the head used as a stiffener.

The procedure also considers the strengthening effect of stiffening rings and the design requirements for the rings themselves. The maximum unstiffened length of the vessel between heads and the total horizontal force that acts against the horns of the saddle may also be determined.


114



The paragraphs that follow discuss several of these stresses (except in stiffening rings) and design details that must be determined and evaluated for the design of horizontal cylindrical vessels on saddle supports. Discussion of the specific equations that are used to calculate these stresses and their allowable limits is beyond the scope of this course. Participants are referred to Zick's paper, or other readily available pressure vessel references for additional details. As previously noted, computer programs such as CODECALC are used for these calculations. Longitudinal Bending Stress - As with an overhanging beam with

two supports, two maximum bending moments caused by the weight load exist in the longitudinal direction of the vessel. One maximum bending moment occurs over the saddle supports, and the other maximum bending moment occurs in the center of the vessel span. The shell acts as a beam over the two supports under the uniform weight load of the vessel and its contents. At midspan, the weight loading causes tensile stresses at the bottom of the cylinder and compressive stresses at the top. At the saddles, the tensile stresses are on the top and compressive stresses are at the bottom. These stresses are combined with the longitudinal stress in the cylinder that is caused by the design pressure, which is a tensile stress for internal pressure. The combined stresses that result are compared to allowable tensile and compressive stresses to determine their acceptability.

Tangential Shear Stress - The distribution and magnitude of the

shear stresses in the shell that are produced by the vessel weight in the plane of the saddle depend on how the shell is reinforced, if at all. •

If the saddle is located away from the head (A > R/2 in Figure 23), but if the saddle is stiffened by a circumferential stiffening ring that is welded to the saddle and shell in the plane of the saddle, the entire cylindrical cross-section resists the load-induced shear stresses. The maximum shear stress occurs at the vessel horizontal centerline.


115



•

If the saddle is located away from the head (A > R/2 in Figure 23), or if the shell is reinforced by two stiffening rings that are located adjacent to the saddle, the effective cross-section of the shell that is able to resist the shear stresses is reduced. The maximum shear stress occurs just beyond the tip of the saddle.

•

If the saddle is located close to the head (A ≤ R/2 in Figure 23), the shell is stiffened on the side of the head. A large part of the weight load that causes the tangential shear stress is carried across the saddle to the head and is then back to the head side of the saddle. The maximum shear stress occurs just beyond the tip of the saddle.

Circumferential Stress at the Horn of the Saddle - The tangential

shear forces in the shell cross-section in the plane of the saddle also cause tangential bending moments and bending stresses in the shell. The maximum bending stress occurs at the horn of the saddle and is compressive. The magnitude of the circumferential bending stress depends on the distance between the head and saddle and thus on whether the head provides any stiffening.

Ring Compression in the Shell Over the Saddle - Forces that act on

the bottom shell band that is located directly over the saddle cause compression in the shell band. These forces are resisted by a portion of the shell on each side of the saddle. It is common for an additional circumferential plate, called a wear plate, to be attached to the vessel shell directly over the saddle, and the saddle is then welded to the wear plate. The wear plate is somewhat larger than the saddle and reduces the compressive stress in the shell since the applied load is spread over a larger area of the shell. Participants are referred to the previously referenced paper by L. P. Zick for additional information on the design of horizontal pressure vessels on two saddle supports.


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COMPLETING SAFETY INSTRUCTION SHEETS FOR PRESSURE VESSELS

The Pressure Vessel Safety Instruction Sheet, Form 2694, must be completed for every new pressure vessel that is within the scope of SAES-D-001. Form 2684 also must be prepared or revised for every pressure vessel which is re-rated or modified. Form 2694 is completed based on the final, certified, as-built, manufacturer's data for the pressure vessel, not the data that is on the vessel data sheet. A copy of a Safety Instruction Sheet is shown in Figure 25, and additional copies are in Course Handout 3.


117



Figure 25: Pressure Vessel Safety Instruction Sheet, Form 2694


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Purpose and Use of the Safety Instruction Sheet in Saudi Aramco

The purpose of the Safety Instruction Sheet is to ensure that operations, maintenance, and inspection personnel will have adequate information in a consistent format. This information concerns safe operating limits, protective devices, and any special safety precautions that are required for pressure vessels. As a result, these personnel have a convenient reference to determine the primary design information and limiting factors of every pressure vessel. The project manager is responsible for the preparation and issue of the Safety Instruction Sheet for new pressure vessels. The actual preparation would probably be done by a contractor's pressure vessel engineer or project engineer. Once completed, the sheet is then reviewed by Saudi Aramco engineers. Saudi Aramco Engineering Standard SAES-A-005, Preparation of Safety Instruction Sheets, outlines the procedures for preparing safety instruction sheets. These procedures are referenced in Work Aid 5. Information Covered

The following highlights several of the primary types of information required on Form 2694. Refer to Figure 25 or Course Handout 3. •

Complete descriptive information of the pressure vessel is required. This information includes the pressure vessel's title, function, plant number, manufacturer's name and country of fabrication, vessel position, serial number, year built, accounting plant number, and Saudi Aramco purchase order number.

•

Other applicable reference information must be supplied. This information includes the applicable design Code and year of issue, reference drawings, and specifications.


119



•

•

Next comes a section that provides detailed mechanical design information for the pressure vessel. This information includes: -

Material specifications, thicknesses, and geometry of the major components.

-

Weld joint types and efficiencies of the major seams.

-

Whether the vessel was stress relieved and the extent of radiographic inspection.

-

The test pressures in the shop and field and the limiting component.

-

Packing, lining, and insulation details.

-

Nozzle and manhole flange rating and facing type.

-

Any special design considerations or unusual construction features. For example, the vessel might have a maximum operating liquid level limitation because the support and associated foundation have a load limitation.

The next section describes the operating limits of the pressure vessel. These limits include: -

Design pressure and temperature, the limiting component, and whether the vessel is adequate for full vacuum. Identification of the limiting component is especially useful when vessel re-rating is under consideration. This identification immediately focuses attention on the first area of the pressure vessel that must be addressed.

-

The location of the safety valve and its set pressure. The safety valve location information is especially useful when a single location is protecting an entire system, rather than just one pressure vessel.

-

The minimum required thickness for the major vessel sections based on specified design conditions. This information is especially useful for maintenance personnel. It tells them when they should become


120



concerned with corrosion that is found in the pressure vessel during periodic inspection. •

The last section provides an area to identify anything else that is special or unusual regarding the pressure vessel. It is important to note here any special hazards, recommendations, inspections, or tests that are important for the pressure vessel.

Form 2694 provides much information about a pressure vessel. It makes it possible for operations, maintenance, and inspection personnel to get needed information from one source without reviewing many drawings. There will be situations when the detailed vessel fabrication drawings must be checked to resolve questions. However, having the information on this one form reduces the need to refer to the drawings and focuses the research on the necessary items. Where to Find Other Information

Almost all of the information that is required on the Safety Instruction Sheet is obtained from the final version of the Pressure Vessel Design Data Sheet and the vessel vendor as discussed above. The paragraphs that follow highlight other items which might have to be obtained from other sources. •

Detailed information about vessel internal trays and packing is often available from the vessel vendor if he also has responsibility for the purchase of these items. Otherwise, the needed information should be available from the contractor if he purchased these items.

•

Information as to the thickness and type of insulation that is used should be available from the contractor. The vessel vendor would have this information since he will supply the needed attachments to the vessel. However, since the vendor does not supply or install the insulation, this information might not appear on the vessel drawings.


121



•

Any special design considerations or unusual construction features would have been developed either during the initial specification of the vessel or during its detailed engineering. Pertinent information could be obtained from the process and mechanical engineers who were assigned to the work.

•

The location of the safety valve that protects the vessel is available from the contractor. The Process and Instrument Diagram (P & ID) for the system will typically show safety valve locations.

•

Information and guidance with regard to any special safety hazards, recommendations, inspections, or tests can be obtained from General Safety Instructions, Aramco GI 2608, and discussions with process, safety, maintenance, and inspection personnel who are assigned to the project and are familiar with the vessel and its application.


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SUMMARY

This module has discussed the mechanical design requirements for pressure vessel components. It has described the use of Saudi Aramco and ASME Code requirements that are applicable to pressure vessel design. Design conditions and loadings, pressure vessel design criteria, and their application to the calculation of pressure vessel component thickness were discussed. Participants have now become familiar with the required mechanical design calculations for pressure vessel components and pressure vessel supports. Participants are now able to evaluate pressure vessel design requirements, using applicable data and Saudi Aramco and industry requirements. After the pressure vessel is designed, it is then fabricated, inspected and tested. These subjects are discussed in MEX 202.04.


123



WORK AID 1: PROCEDURES FOR EVALUATING THE ACCEPTABILITY OF CONTRACTOR-SPECIFIED DESIGN CONDITIONS AND LOADINGS

The procedures in this Work Aid may be used to determine if the design conditions and loadings that are specified in a Contractor Design Package are acceptable. Work Aid 1A: Design Pressure and Temperature

All pressure vessels must be designed for the most severe combination of pressure and temperature which may be imposed. Multiple pressure/temperature combinations may be possible for a particular vessel, and the most severe combination must be used for design. Calculated design pressures and temperatures must be specified on the Pressure Vessel Design Data Sheet. The Contractor Design Package for the pressure vessel must be checked to ensure that these design conditions have been properly specified. Design Pressure •

Determine the maximum expected internal operating pressure, Po. This should be specified by process engineers. Po = _____kPa(ga), (psig)

•

Calculate the minimum required design pressure at the top of the vessel, PT, as the larger of that determined from the following equations: PT = P0 + [103 kPa(ga) or 15 psig], kPa(ga), (psig) PT = 1.1P0, kPa(ga), (psig)


124



If the vessel is partially or completely filled with liquid, calculation of the pressure that is used for the design of any components that are located below the liquid level, P, must include consideration of the hydrostatic head of the liquid. In these cases, the following equations are used to calculate the pressure below the liquid level. SI Units P = PT + 0.00981 γH or P = PT + 9.81(SG)H English Units P = PT +

γH , psig 144

or P = PT + 0.433(SG)H, psig Where:

P

= Design pressure at the point under consideration, kPa (psig)

PT

= Design pressure at the top of the vessel, kPa (psig)

γ

= Weight density of the liquid in the vessel, kg/m3 (lb./ft.3)

H

= Height of the liquid above the point under consideration, m (ft.)

SG = Specific Gravity of the liquid


125



•

Is vessel subject to external pressure conditions during operation (i.e. full or partial vacuum conditions)? Yes_____ No_____ If "Yes," what is the maximum expected external operating pressure, Poe?

Determine the required external design pressure, Pe, as the smaller of 1.25Poe or 103 kPa (15 psig). For full vacuum conditions, Pe = 103 kPa (15 psig). •

Is vessel subject to steam-out conditions? Yes_____ No_____ If "Yes," vessel must also be designed for an external pressure of 52 kPa(ga) (7.5 psig) at 149°C (300°F).

Design Temperature •

Determine the maximum expected operating temperature, T0. This should be specified by process engineers. T0 = _____ °C (°F)

•

Calculate the minimum required design temperature, T, based on maximum operating temperature as: T = T0 + (28°C or 50°F) If T0 is less than -18°C (0°F), the design temperature shall be no higher than: T = T0 - (14°C or 25°F)

•

Determine Minimum Design Metal Temperature or Critical Exposure Temperature (CET). The CET is the minimum metal temperature that is coincident with a pressure greater than 25% of the vessel design pressure. CET = _____°C (°F)


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Work Aid 1B: Other Loadings

Other loadings will typically not have to be initially specified on the Pressure Vessel Design Data sheet since they either will be already known to the vendor (such as the design wind velocity), or are unknown at the time the sheet is completed (such as piping reaction loads at nozzles). Use the following checklist for review purposes to ensure that the vendor has what he needs for bidding purposes. 1.

Additional weight loads that are imposed by equipment supported by vessel specified.

2.

Required internal components, lining, and insulation specified.

3.

No indication that design includes other loads unless backup information is provided.

Work Aid 1C: Service

Vessel service information must be specified on the Pressure Vessel Design Data Sheet to ensure that material selection, fabrication, and inspection requirements will be included in the vendor bid. This service is specified by process engineers. Confirm that the following information is specified: 1.

Vessel service specified.

2.

Yes/No response required to whether wet, sour or lethal service requirements apply. Identify which service applies if "yes" response.


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WORK AID 2: PROCEDURES FOR EVALUATING THE ACCEPTABILITY OF CONTRACTOR-SPECIFIED PRESSURE VESSEL COMPONENT THICKNESS DESIGN CRITERIA

The procedures in this Work Aid may be used to determine if the vessel component thickness design criteria that are specified in a Contractor Design Package are acceptable. Work Aid 2A: Determination of Weld Joint Efficiency

1.

Determine the Category of welded joint from Figure 26. This is adapted from Figure UW-3 of the ASME Code that is in Course Handout 1.

C

C

C A C

A A

D

B B

D

A

B D B A

D

B

A C C

D

20203.F26

Figure 26: Weld Joint Categories

2.

Determine the Type of welded joint being used. See Figure 25 for weld Types. Type 1 joints will typically be used.

3.

Confirm that the Type of welded joint determined in Step 2 is acceptable for the Category determined in Step 1. See Table UW-12 in the ASME Code that is in Course Handout 1. Table UW-12 is excerpted in Figure 28. A Type 1 joint is acceptable for all weld joint categories.


128



4.

Determine the degree of radiographic examination that is specified for the weld: Full, spot, or none. Either full or spot radiography will always be done.

5.

Determine the weld joint efficiency, E, from Table UW-12 in the ASME Code that is in Course Handout 1 (excerpted in Figure 28). For a Type 1 joint:

•

E = 1.0 for full radiography

•

E = 0.85 for spot radiography


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Joint Type

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

1

Backing strip if used shall be removed after completion of weld. Single-welded butt joint with backing strip which remains in place after welding.

2

For circumferential joint only

Single-welded butt joint without backing strip.

3

4

Double-full fillet lap joint.

5

Single-full fillet lap joint with plug welds.

6

Single-full fillet lap joint without plug welds.

MEX 20203.F27

Figure 27: Types of Welded Joints


130



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

Figure 28: Maximum Weld Joint Efficiency

6.

Confirm that the weld radiography and joint efficiency that are specified on the Pressure Vessel Design Data Sheet are consistent with these determinations.


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Work Aid 2B: Determining Corrosion Allowance

1.

Determine the corrosion allowance that is required, by discussion with process and materials engineers as necessary.

2.

For carbon steel pressure vessels: •

Minimum and maximum corrosion allowances are 1.6 mm (1/16 in.) and 6 mm (1/4 in.) respectively for pressure containing components.

•

Removable carbon steel internal components must have a corrosion allowance that is equal to that of the shell.

•

Nonremovable carbon steel internal components must have two times the specified shell corrosion allowance.

3.

Confirm that the corrosion allowance specified on the Pressure Vessel Design Data Sheet is consistent with other available information that is cited above.

4.

Determine the necessary corrosion allowance for noncarbon steel vessels or internal components by discussion with process and materials engineers.


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WORK AID 3:

PROCEDURES FOR EVALUATING CONTRACTOR-SPECIFIED DESIGN CALCULATIONS FOR PRESSURE VESSEL COMPONENTS

The procedures in this Work Aid may be used to determine if design calculations for pressure vessel components (or their end results) that are contained in a Contractor Design Package are acceptable. The following checklist may be used as a reference for pressure vessel components that are designed to the ASME Code Section VIII, Division 1. Later sections of this Work Aid provide detailed procedures for performing (or checking) portions of this work. •

Determine that vessel design conditions, corrosion allowance, and component geometries are specified.

•

Determine that component stresses are specified.

•

Determine that weld inspection and joint efficiency are specified and are consistent with each other.

•

Identify types of heads used, and whether vessel has any internal heads.

•

Calculate minimum required design thickness of shell. Specify nominal thickness (may be left to vendor).

•

Calculate minimum required design thickness of heads. Specify nominal thickness (may be left to vendor).

•

Determine that Flange Class is consistent with specified design pressure and temperature.

•

Determine that MAWP is calculated for design temperature and corroded components. Confirm that MAWP is not limited by nozzle reinforcement.


materials

and

allowable

133



•

Work Aid 3A:

Determine that stresses from local loads applied to nozzles and attachments have been evaluated as needed. This determination is typically not part of initial preparation of the Pressure Vessel Design Data Sheet, but is done during detailed engineering. If loads are known and specified on sheet, confirm that the vendor at least considered these in his bid (even if actual calculations are not submitted with his bid).

Required Wall Thickness for Internal Pressure of Pressure Vessel Components

The following procedure may be used to evaluate pressure vessel components for internal pressure.

Part Cylindrical shell Spherical shell 2:1 Semi -Elliptical head Torispherical head with 6% knuckle Conical Section (α ≤ 30°)

1.

Identify the geometry of the part under consideration: Cylinder, spherical shell or head, 2:1 semi-ellipitical head, torispherical head, conical section. Also determine appropriate material allowable stress and weld joint efficiency. Obtain this information from the Contractor Design Package.

2.

Use the appropriate equation for calculation of the required wall thickness for internal pressure, tp, from the following: Thickness, tp, in. Pr SE1 − 0.6P Pr 2SE1 − 0. 2P PD 2SE − 0.2P

Pressure, P, psi SE1t r + 0.6 t 2SEt r + 0.2 t 2SEt D + 0.2 t

Stress, S, psi P (r + 0.6 t ) tE1 P (r + 0. 2t ) 2tE P (D + 0.2 t ) 2tE

0.885PL SE − 0.1P

SEt 0.885L + 0.1t

P (0.885L + 0.1t ) tE

PD 2 cos α (SE − 0.6P)

2SEt cos α D + 1. 2t cos α

P (D + 1. 2t cos α ) 2 tE cos α


134



Where:

P = Internal deign pressure, kPa (psig). When used in the pressure calculation equations, this is the MAWP. r

= Internal radius, mm (in.). Add corrosion allowance to specified uncorroded internal radius.

S = Allowable Stress, kPa (psi). When used in the thickness calculation equations, this is the allowable stress for the material used. Allowable stress was discussed in MEX 202.02. 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, mm (in.). D = Inside diameter, mm (in.). Add twice the corrosion allowance to specified uncorroded inside diameter. DL = Cone inside diameter at large end, mm (in.). Add twice the corrosion allowance to specified uncorroded inside diameter. DS = Cone inside diameter at small end, mm (in.). Add twice the corrosion allowance to specified uncorroded inside diameter. L = Inside crown radius of torispherical head, mm, (in.). Add corrosion allowance to specified uncorroded inside crown radius. α = One half of the apex angle of the cone at the centerline, degrees. α = tan−1


0.5(DL − Ds ) (Cone Length)

135



3.

Determine the minimum required component thickness, t, by adding the specified corrosion allowance, C, to tp that was determined in Step 2. t = tp + c

4.

Work Aid 3B:

Include calculated required thickness from Step 3 in appropriate section on Pressure Vessel Design Data Sheet. When checking a Contractor Design Package for acceptability, confirm that the specified component thickness is at least equal to the calculated value.

Required Wall Thickness for External Pressure of Pressure Vessel Components, and Allowable Compressive Stress of Cylindrical Shell

The following procedure may be used to evaluate pressure vessel components for external pressure. It may also be used to determine the allowable compressive stress of a cylindrical shell. Cylindrical or Spherical Shells Under External Pressure

Nomenclature A

= Factor determined from Figure G in Subpart 3 of Section II, Part D of the ASME Code. It is used to enter the applicable material chart in Subpart 3 of Section II, Part D. See the copy of Section II that is in Course Handout 1. Note that Figure 11 is an excerpt from the appropriate figure.

B

= A factor determined from the applicable material chart in Subpart 3 of Section II, Part D of the ASME Code, for maximum design metal temperature, kPa (psi). Note that the lower of "B" or the allowable tensile stress is the allowable compressive stress of cylindrical shells. See the copy of Section II that is in Course Handout 1. Note that Figure 12 is an example of one of these material charts.

Do

= Outside diameter of a cylindrical shell, mm (in.).


136



E

= Young's modulus of elasticity at design temperature for the material, kPa (psi). If needed, obtain from same material chart that was used to determine "B". Do not confuse this parameter with the weld joint efficiency, E, that is used elsewhere.

L

= The total length, mm (in.), of a tube between tubesheets or the design length of a vessel section between lines of support. A line of support is: (1)

A circumferential line on a head at one-third the depth of the head from the head tangent line (excluding conical heads and sections).

(2)

A cone-to-cylinder junction or a knuckle-tocylinder junction of a toriconical head or section, that satisfies the Code moment of inertia requirements.

(3)

A stiffening ring that meets Code requirements.

(4)

A jacket closure of a jacketed vessel that meets Code requirements.

P

= The external design pressure, kPa (psi). This is 103 kPa (15 psi) for full vacuum design, and 52 kPa (ga)(7.5 psig) for steamout conditions.

Pa

= The calculated maximum allowable external working pressure for the assumed value of t, kPa (psi).

Ro

= The outside radius of a spherical shell, mm (in.).

t

= The minimum required thickness of a cylindrical shell or tube, or spherical shell, mm (in.). The required corrosion allowance must be added to this value.

ts

= The nominal thickness of a cylindrical shell or tube, mm (in.).


137



Two procedures exist for calculating allowable external pressure for cylindrical shells and tubes. The selection between the two procedures is based on the ratio Do/t. Cylinders with a Do/t ≥10 are calculated as follows: Step 1:

Determine L and Do.

Step 2:

Assume a value of t and determine the ratios of L/Do and Do/t.

Step 3:

Enter Figure G in Subpart 3 of Section II, Part D at the value of L/Do from Step 2. If L/Do is greater than 50, use L/Do = 50. For L/Do less than 0.05, use L/Do = 0.05.

Step 4:

Move horizontally to the line for the value of Do/t determined in Step 1. Use interpolation for intermediate values of Do/t. Move vertically downward from this intersection point to determine Factor A.

Step 5:

Using the value of A from Step 4, enter the applicable material chart in Subpart 3 of Section II, Part D. Move vertically in this chart to the intersection with the correct design temperature line. Use interpolation for intermediate temperatures. If A is to the right of the end of the material/temperature line, assume an intersection with the horizontal projection of the upper end of the material/temperature line. If A is to the left of the material/temperature line, go to Step 8.

Step 6:

From the intersection obtained in Step 5, move horizontally to the right and read the value of Factor B.

Step 7:

Using the value of B from Step 6, calculate Pa using the following: Pa =


4B 3(Do / t) 138



Step 8:

If A is to the left of the material/temperature line, calculate Pa using the following: Pa =

Step 9:

2 AE 3(Do / t)

Compare Pa with P. If Pa is greater than or equal to P, the design is acceptable. The minimum thickness of the cylinder must be at least equal to t, plus any corrosion and forming allowances. If there is an economic incentive, reduce the value of t and repeat the procedure to arrive at a Pa that is closer to P. If Pa is less than P, increase t or decrease L and repeat the procedure. This procedure must be repeated until Pa is greater than or equal to P.

Step 10: Include the required minimum thickness on the Pressure Vessel Design Data Sheet. If stiffening rings are required for this minimum shell thickness, this and their maximum permitted spacing, L, must be specified. Cylinders with a Do/t < 10 are calculated as follows: Step 1:

Use the same procedure as with Do/t > 10 to determine B. However, for Do/t less than 4, calculate A from the following: A=

1.1 (Do / t)2

For A greater than 0.1, use A = 0.1 Step 2:

Using the value of B from Step 1, calculate Pa1, using the following:  2.167  Pa1 =  − 0. 0833 B  (Do / t) 


139



Step 3:

Calculate Pa2 using the following: Pa 2 =

2S Do / t

 1  1−  (Do / t) 

S is the lesser of two times the material allowable stress in tension at design temperature, or 90% of the material yield strength at design temperature. Yield strength may be obtained from the applicable external pressure chart as follows:

Step 4:

(a)

For a given temperature curve, determine the value of B that corresponds to the right end point of the curve,

(b)

The yield strength is double the value of B that is obtained from (a).

The smaller of Pa1 or Pa2 is compared with P to determine acceptability of the value of t. The procedure from this point is the same as for Do/t > 10.

The minimum required thickness for spherical shells under external pressure is determined using the procedure that follows. Step 1:

Assume a value for t and calculate the Factor A using the following: A=

Step 2:

0.125 Ro / t

Using the value of A from Step 1, enter the applicable material chart in Subpart 3 of Section II, Part D. Move vertically in this chart to the intersection with the correct design temperature line. Interpolate for intermediate temperatures. If A is to the right of the end of the temperature line, assume an intersection with the horizontal projection of the upper end of the material/temperature line. If A is to the left of the temperature line, go to Step 5.


140



Step 3:


Step 4:

Using the value of B from Step 3, calculate Pa using the following: Pa =

Step 5:

If A is to the left of the temperature line, calculate Pa using the following: Pa =

Work Aid 3B:

B Ro / t

0.625E (Ro / t)2

Required Wall Thickness for External Pressure of Pressure Vessel Components, and Allowable Compressive Stress of Cylindrical Shell, cont'd

Step 6:

Compare Pa with P. If Pa is greater than or equal to P, the design is acceptable. The minimum thickness of the spherical shell must be at least equal to t, plus any corrosion and forming allowances. If there is an economic incentive, reduce the value of t and repeat the procedure to arrive at a Pa that is closer to P. If Pa is less than P, increase t and repeat the procedure until Pa is greater than or equal to P.

Step 7:

Include the required minimum thickness on the Pressure Vessel Design Data Sheet.

Heads and Conical Sections Under External Pressure

Nomenclature The following is additional nomenclature used in the design of heads and conical sections for external pressure. The definitions of A, B, E and P are the same as for cylindrical and spherical shells. Do =


Outside diameter of the head skirt, mm (in.).

141



Do/2ho = Ratio of the major to minor axis of elliptical heads, which equals the outside diameter of the head skirt divided by twice the outside height of the head. =

Effective thickness of a conical section, mm (in.).

=

t cos α.

=

Equivalent length of a conical section, mm (in.).

=

(L/2)(1+Ds/DL)

L

=

Axial length of a cone or conical section, mm (in.).

Ds

=

Outside diameter at the small end of the conical section under consideration, mm (in.).

DL

=

Outside diameter at the large end of the conical section under consideration, mm (in.).

ho

=

One-half of the length of the outside minor axis of the elliptical head, or the outside height of the elliptical head, measured from the tangent line (head-bend line), mm (in.).

Ko

=

A factor that depends on Do/2ho as determined from Table UG-33.1 of the ASME Code. See the copy of the ASME Code in Course Handout 1.

Ro

=

For hemispherical heads, the outside radius, mm (in.).

Ro

=

For elliptical heads, the equivalent outside spherical radius taken as KoDo, mm (in.).

Ro

=

For torispherical heads, the outside radius of the crown portion of the head, mm (in.).

t

=

The minimum required thickness of a head after forming, mm (in.). The required corrosion and forming allowances must be added to this.

α

=

One-half of the apex angle in conical heads and sections, degrees.

te

Le


142



Heads may be installed inside a pressure vessel to separate it into compartments, in addition to their more common use as a part of the pressure shell. When an internal, intermediate head is used, it must be designed for the pressure on each side of it. The pressure that acts on the convex side is an external pressure. The external pressure that acts on an internal head is based on the process design conditions, and may greatly exceed 103 kPa(g)(15 psig). The minimum required thickness for heads under external pressure must be the greater of the following thicknesses: •

The thickness calculated using the equations for internal pressure, but using a pressure equal to 1.67 times the external design pressure.

•

The thickness computed using the procedures for the specified head type, which are detailed in the following paragraphs.

Minimum Required Thickness for Elliptical Heads The required thickness of an elliptical head under external pressure, or with pressure on the convex side, is calculated using the procedure that follows. The correct value of Ro is as previously defined. Step 1:

Determine Ro

Step 2:

Assume a value for t and calculate Factor A using the following: A=

Step 3:


0.125 (Ro / t)

With the value of A from Step 2, use the same procedure that was previously discussed for spherical shells.

143



Minimum Required Thickness for Torispherical Heads The required thickness of a torispherical head under external pressure, or with pressure on the convex side, is calculated using the same procedure as for elliptical heads. The correct value of Ro is as previously defined. Minimum Required Thickness for Hemispherical Heads The required thickness of a hemispherical head under external pressure, or with pressure on the convex side, is calculated using the same procedure as for a spherical shell. Minimum Required Thickness for Conical Heads or Sections The required thickness of a conical head or section (without transition knuckles) under external pressure is calculated by the use of one of the following procedures. The selection of a procedure is based on the values of α and DL/te. When α is < 60 ° and for cones having DL/te ≥10, use the steps that follow:

Step 1:

Determine Le and DL.

Step 2:

Assume a value for te and determine the ratios Le/DL and DL/te.

Step 3:

Enter Figure G of Subpart 3 of Section II, Part D at a value of L/Do equivalent to the value of Le/DL found in Step 2. For Le/DL greater than 50, enter the chart at Le/DL = 50.

Step 4:

Move horizontally to the line for the value of Do/t equal to the value of DL/Te determined in Step 2. Interpolation may be used. From this intersection point, move vertically down to determine Factor A.

Step 5:

Using A determined chart in Subpart 3 vertically to the line temperature. Use temperatures.


in Step 4, enter the applicable of Section II, Part D. Move that corresponds to the design interpolation for intermediate

144



If A is to the right of the end of the material/temperature line, assume an intersection with the horizontal projection of the upper end of the material/temperature line. If A is to the left of the temperature line, go to Step 8. Step 6:


Step 7:

Using the value of B from Step 6, calculate the value of Pa using the following: Pa =

Step 8:

If A is to the left of the material/temperature line, calculate Pa using the following: Pa =

Step 9:

4B 3(DL / t e )

2 AE 3(DL / t e )

Compare Pa with P. If Pa is greater than or equal to P, the design is acceptable. The minimum thickness of the cone must be at least equal to t, plus any corrosion and forming allowances. If there is an economic incentive, reduce the value of t and repeat the procedure to arrive at a Pa that is closer to P. If Pa is less than P, then increase t and repeat the procedure. Repeat this procedure until Pa is greater than or equal to P.

Step 10: Include the required minimum thickness on the Pressure Vessel Design Data Sheet.


145



The cone-to-cylinder junction must also be checked to determine if there is adequate reinforcement. This is also covered by ASME Code procedures. However, cone-to-cylinder junction reinforcement does not influence design of the cone in most circumstances. Refer to the ASME Code for details. Use the following procedure for cones having DL/te < 10.

Step 1:

Use the same procedure as above to determine B. However, for DL/te less than 4, calculate A using the following: A=

1.1 (DL / t e )2

If A is greater than 0.1, use A = 0.1 Step 2:

Using the value of B from Step 1, calculate Pa using the following:  2.167  Pa1 =  − 0. 0833 B  (DL / te ) 

Step 3:

Calculate Pa2 using the following: Pa 2 =

2S DL / t e

  1 1−  (DL / t e ) 

S is defined identically as for cylinders having Do/t <10, as previously discussed. Step 4:

The smaller of Pa1 or Pa2 is compared with P to determine the acceptability of the value for t. The procedure from this point is the same as for cones with DL/te > 10.

Step 5:

Reinforcement of the cone-to-cylinder junction must be checked as for cones with DL/te > 10.


146



Use the following procedure when α > 60°:

The thickness of the cone must be the same as the required thickness of a flat head under external pressure. In this case, the diameter of the head is assumed to be equal to the largest diameter of the cone. It is unusual to see a cone with an angle this large. Refer to the ASME Code for the required calculation procedure.

Allowable Compressive Stress of Cylindrical Shells

The allowable compressive stress of a cylindrical shell is the lower of the allowable tensile stress that was discussed in MEX 202.02, or the value of B determined from the following procedure. Step 1:

Determine Ro, outside radius of the cylindrical shell, in.

Step 2:

Calculate: A=

0.125 (Ro / t)

Step 3:

Using the value for A calculated from Step 2, enter the applicable material chart in Subpart 3 of Section II, Part D. Move vertically in this chart to the intersection with the correct design temperature line. Use interpolation for intermediate temperatures. If A is to the right of the material/temperature line, assume an intersection with the horizontal projection of the upper end of the material/temperature line. If A is to the left of the material/temperature line, go to Step 5.

Step 4:

From the intersection that is obtained in Step 3, move horizontally to the right and read the value of factor B. This is the maximum allowable compressive stress for the values of t and Ro that were used in Step 1.


147



Step 5:

If A is to the left of the material/temperature line, calculate B using the following formula: B=

AE 2

B is the allowable compressive stress.

Work Aid 3C:

Step 6:

If the value of B that was determined in Steps 4 or 5 is smaller than the computed compressive stress, a greater value of t must be selected and the procedure is repeated. The calculated value of B must be greater than the calculated compressive stress.

Note:

In this procedure, the efficiency of butt-welded joints may be taken as one.

Nozzle Reinforcement for Pressure

Refer to Figure UG-37.1 in the ASME Code for nozzle geometry and the general forms of the equations that are used for nozzle reinforcement calculations. Figure 13 is an excerpt from this figure. The following procedure is valid for the most common case where the strengths of the nozzle and reinforcing pad materials are at least equal to that of the shell or head material to which they are attached. The procedure also neglects any reinforcement contribution from weld metal, since this contribution is small, and assumes that there is no internal nozzle projection. 1.

Calculate the required reinforcement area, A. A = dtrF, mm2 (in.2) 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, mm (in.).

tr =

Minimum required thickness of the shell using appropriate ASME Code formula and a weld joint


148



efficiency of 1.0, mm (in.). determine this thickness. F =

Use Work Aid 3A to

Correction factor normally equal to 1.0. See Figure UG-37 of the ASME Code for integrally reinforced openings in cylindrical shells and cones.

"A" for openings that are subject to external pressure is 50% of that calculated using this equation. tr is the value that is required for external pressure. 2.

Determine the reinforcement limits measured parallel to the vessel wall as a distance on each side of the axis of the opening equal to the greater of the following: d, or (Rn + tn + t)

Where: Rn = t

= Thickness of the vessel in the corroded condition, mm (in.).

tn = 3.

Radius of the finished opening in the corroded condition, mm (in.).

Nominal thickness of the nozzle in the corroded condition, mm (in.).

Calculate the reinforcement limits measured normal to the vessel wall as the smaller of the following: 2.5t, or (2.5 tn + te) Where:

4.

te =

0 if there is no reinforcing pad.

te =

Reinforcing pad thickness if one is installed, mm (in.).

te =

As defined in Figure UG-40 of the ASME Code for self-reinforced nozzles, mm (in.).

Calculate the reinforcement area that is available in the vessel wall, A1, as the larger of the following:


149



A1 = (Elt - Ftr)d or A1 = 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. 5.

Calculate the reinforcement area that is available in the nozzle wall, A2, as the smaller of the following: A2 = (tn-trn)5t or A2 = 2(tn-trn)(2.5 tn + te) Where: trn =


Required thickness of a seamless nozzle wall, mm (in.). Use Work Aid 3A to determine this thickness.

150



6.

Nozzle is adequately reinforced if: A1 + A2 > A If this relationship is not true, then additional reinforcement is required.

7.

If a reinforcing pad is used, its area contribution to reinforcement, A5, may be calculated from the following: A5 = [Dp - (d + 2 tn)]te Where: Dp = Reinforcing pad diameter, mm (in.). When a reinforcing pad is used, in general select te = (t + c), the nominal shell thickness. Then determine the required Dp such that: A1 + A2 + A5 > A Dp must not extend beyond the reinforcement limit that is parallel to the shell.

Work Aid 3D: Determine Required Flange Rating

Use the following procedure to determine the required flange rating (or Class) in accordance with ASME/ANSI B16.5. Reference the copy of ASME/ANSI B16.5 that is in course Handout 1. 1.

Note the design pressure and temperature, and generic material type and specification (i.e. carbon steel, 1 1/4 Cr1/2 Mo, etc.) of the pressure vessel. These values are on the Pressure Vessel Design Data Sheet. P

=

_____ kPa (psig)

T

=

_____°C (°F)

Material

=

_____


151



2.

Go to Table 1 of 32-SAMSS-004 and select the flange material specification corresponding to the "Vessel Service Classification" (Reference Course Handout 2). If the pressure vessel material does not correspond to a Table 1 selection due to special circumstances, select a flange material specification with matching chemistry to the vessel material. See MEX 202.02 for additional information on material selection.

3.

Go to Table 1A of ASME/ANSI B16.5 and determine the Material Group No. for the selected material specification. Note that Figure 16 is an excerpt from this table.

4.

Go to Table 2 of ASME/ANSI B16.5 with the design temperature and Material Group No. determined in Step 3. Note that Figure 17 is an excerpt from this table.

5.

•

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

•

Table 2 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.

Confirm that the flange Class that is specified in the Contractor Design Package is acceptable.


152



Work Aid 3E:

Evaluating the Stresses Resulting From Nozzle Loads Applied to a Vessel Shell

This Work Aid may be used to evaluate the local stresses in a spherical shell that result from loads imposed at a nozzle attachment. It is based on requirements that are contained in WRC 107. Refer to WRC 107 or WRC 297 for vessel or attachment geometries that are not covered by this Work Aid. It is assumed that these stresses have been calculated by a contractor or vendor using a computer program such as CODECALC, and that they must be evaluated for acceptability. 1.

Determine vessel geometric information and applied loads. This should be readily available from the vessel drawing information and other given information. Confirm that the contractor has used this information in his stress calculations. Note the nomenclature and coordinate directions for the applied loads in Figure 29.

Applied Loads Radial Load

P

=

kg (lb.)

Shear Load

V1

=

kg (lb.)

Shear Load

V2

=

kg (lb.)

Overturning Moment

M1

=

kg-m (in.-lb.)

Overturning Moment

M2

=

kg-m (in.-lb.)

Torsional Moment

MT

=

kg-m (in.-lb.)

Design Pressure

P

=

kPa (psig)


153



P, Radial Load MT, Torsional Moment

V2, Shear C to D M2, Moment A to B rm, Nozzle Mean Radius

V1, Shear A to B M1, Moment C to D

t, Nozzle Thickness

Nozzle B

C

A

Spherical Shell or Head

D T, Sphere Thickness

Rm, Sphere Mean Radius

Figure 29: Nozzle Loads Applied to a Spherical Shell


154



Geometry Sphere Thickness

T

=

mm (in.)

Sphere Mean Radius

Rm

=

mm (in.)

Nozzle Thickness

t

=

mm (in.)

Nozzle Mean Radius

rm

=

mm (in.)

2.

3.

Determine the allowable stresses for the vessel design temperature and material. The design temperature, material and allowable stress in tension will be available from the vessel drawings and Design Data Sheet •

Design Temperature

•

Sphere Material

•

Allowable stress in tension, S

kPa (psi)

•

Allowable membrane stress intensity = 1.5S

kPa (psi)

•

Allowable combined membrane plus bending stress intensity = 3S

kPa (psi)

°C (°F)

Confirm that the computer program used by the contractor or vendor for calculating the vessel shell stresses is based on WRC 107.


155



4.

5.

From the computer program results provided by the contractor or vendor, determine the following stresses: •

Membrane stress in the shell due to internal pressure.

•

Membrane stresses in the shell due to the applied loads. Note that these stresses will typically be reported at four locations around the nozzle as shown in Figure 29, and at both the inner and outer surfaces of the shell.

•

Bending stresses in the shell due to the applied loads. These stresses will be reported at the same locations as the membrane stresses.

From the computer program results provided by the contractor, determine the following combined stresses: •

Maximum membrane stress intensity at any point.

•

Maximum membrane plus bending stress intensity at any point.

These combined stress intensities consider both internal pressure and the applied loads. 6.

Compare the maximum stress intensities determined in Step 5 with the allowable values determined in Step 2. The design is acceptable if the calculated stress intensities are no greater than their allowable values. If the calculated stress intensities are greater than their allowable values, then additional reinforcement is required.


156



WORK AID 4:

PROCEDURES FOR EVALUATING THE CONTRACTOR-SPECIFIED DESIGN OF PRESSURE VESSEL SUPPORTS

The procedures in this Work Aid may be used to evaluate the design of pressure vessel supports that are specified in a Contractor Design Package. Work Aid 4A: Vertical Vessel on Column Supports

This procedure may be used for evaluating unbraced column supports of vertical pressure vessels. Refer to Figure 30 in using this procedure. Note that this procedure is only for design of the columns and does not evaluate the stresses in the vessel shell. 1.

2.

3.

Determine the total weight of the pressure vessel, attachments, and contents to be included for the case under consideration. This information should be available from the vendor or contractor and shown on the Pressure Vessel Design Data Sheet, or may be calculated using procedures that are contained in PEDP course CSE 110. Wo =

kg (lb.), for operating conditions

WT =

kg (lb.), for hydrotest conditions

We =

kg (lb.), for empty conditions

Determine the horizontal force, P, at the vessel centroid due to either wind or earthquake for both the design wind velocity (137 km/h [85 MPH]) and hydrotest wind velocity (48 km/h [30 MPH]). This information should be available from the vendor or contractor, or may be calculated using procedures that are contained in PEDP course CSE 110. Po =

kg (lb.) during operation

PT =

kg (lb.) during test

Calculate the moments, Mb and Ma, for both the design case and test case. Mb = P(H + L)


157



Ma = PH Where: H

= Distance from vessel centroid to the bottom tangent line, mm (in.)

L

= Distance from bottom tangent line to grade, mm (in.)

Mb = Moment at the base, kg-mm (in.-lb.) Ma = Moment at the tangent line, kg (lb.)

D

A

t

x

P e

Direction of wind or earthquake

P1

P2

H

Db y

y P T.L.

F x L 3L 4 a

A a

Section a-a

CL column

Elevation Moment at base: Mb = P(H + L) Moment at T.L.: = Ma = PH ΣI = 2lx + 2ly

Figure 30: Vessel on Column Supports


158



4.

Make an initial assumption of the number and structural shape to be used for the column supports. These may be standard structural members or pipe. N = Number of columns A = Cross sectional area of one column, mm2 (in.2) Ix, Iy, Iz =

Moments of inertia of selected column section (See Figure 31), mm4 (in.4)

r = Least radius of gyration of the selected section, mm (in.) I = Moment of inertia of one leg perpendicular to the direction of the wind or earthquake load, mm4 (in.4) For example, a vessel with four column supports as shown in Figure 30 will have: Σ I1 = 2Ix + 2 Iy and I = Ix See Figure 31 for other column orientations. 5.

Determine the maximum allowable compressive stress for the column, Fa, from Figure 32. Note that for most column and base plate attachment details, K1 = 1.5 is a reasonable approximation of the column's behavior. Use "r" equal to the least radius of gyration of the column section.

6.

Determine the maximum allowable bending stress, Fb, as: Fb = 0.6 Fy Where: Fy =


Minimum specified yield stress of the column material, kPa (psi)

159



7.

Calculate the maximum total axial compressive load on the leeward side of each column, kg (lb.).

Beams, channels and rectangular tubing

Angle legs Y V

c1

e

e

Z

X

Fh

X

e

Fh

Y

Y

X

U U

Fh Y

c1

e Fh

X

W

W

c1

V

e

Z b

Y

a c1

V

e

c1

Y

Z

a

Fh

X

e

Fh

X

U U

Fh

X

X

Fh

W θ

W

e

45° only

Z

V

Y

Y Z

b ≅ e

fb = M C1 l

θ

lw = lx sin2θ + ly cos2θ lz = lx cos2θ + ly sin2θ

[

fb = M

Fh

c1 W

]

b sinθ + a cosθ ly lx

W

fb = M C1 l

Z

fb = M C 1 l

lz = r2A lw = lx + ly + lz lv = lw cos2θ + lz sin2θ lu = lw sin2θ + lz cos2θ

e

MEX 20203.F31

Figure 31: Vessel Column Configurations and Moments of Inertia


160

; ; ; ; ; ; ; ; ;




K1 End Connection Coefficient (a)

(b)

(c)

(d)

(e)

(f)

Buckled shape of column is shown by dashed line

Theoretical K1 value Recommended design value when ideal conditions are approximated

0.5

0.7

1.0

1.0

2.0

2.0

0.65

0.80

1.2

1.0

2.10

2.0

Rotation fixed and translation fixed Rotation free and translation fixed

End condition code

Rotation fixed and translation free Rotation free and translation free

Reprinted by permission of AISC.

MEX 20203.F32

Fa, Allowable Compression Stress K1l r 21 22 23 24 25

Fa (ksi) 20.54 20.48 20.41 20.35 20.28

41 42 43 44 45

19.11 19.03 18.95 18.86 18.78

61 62 63 64 65

17.33 17.24 17.14 17.04 16.94

81 82 83 84 85

15.24 15.13 15.02 14.90 14.79

101 102 103 104 105

12.85 12.72 12.59 12.47 12.33

26 27 28 29 30

20.22 20.15 20.08 20.01 19.94

46 47 48 49 50

18.70 18.61 18.53 18.44 18.35

66 67 68 69 70

16.84 16.74 16.64 16.53 16.43

86 87 88 89 90

14.67 14.56 14.44 14.32 14.20

106 107 108 109 110

12.20 12.07 11.94 11.81 11.67

31 32 33 34 35

19.87 19.80 19.73 19.65 19.60

51 52 53 54 55

18.26 18.17 18.08 17.99 17.90

71 72 73 74 75

16.33 16.22 16.12 16.01 15.90

91 92 93 94 95

14.09 13.97 13.84 13.72 13.60

111 112 113 114 115

11.54 11.40 11.26 11.13 10.90

36 37 38 39 40

19.50 19.42 19.35 19.27 19.19

56 57 58 59 60

17.81 17.71 17.62 17.53 17.43

76 77 78 79 80

15.79 15.69 15.58 15.47 15.36

96 97 98 99 100

13.48 13.35 13.23 13.10 12.98

116 117 118 119 120

10.85 10.71 10.57 10.43 10.28

Main and Secondary Members

Figure 32: Allowable Column Compressive Stress


161



Co =

W o 4Mbo + , for the operating conditions N NDb

CT =

W T 4MbT + , for the test conditions N NDb

Where: Db = Diameter of circle passing through centroids of columns, mm (in.). 8.

9.

Calculate the maximum total axial uplift load (tensile) on the windward side of each column, kg (lb.). To = −

W o 4Mbo + , for the operating conditions N NDb

Te = −

W e 4Mbo + , for the empty vessel N NDb

Calculate the eccentric loads at the top of the column, kg (lb.). P1 =

W o 4Mao + , for the operating conditions N ND

P1 =

W T 4MaT + , for the test conditions N ND

P2 =

4Mao W o − , for the operating conditions ND N

P2 =

4Mao W e − , for the empty vessel ND N

10. Calculate the lateral force per column, F, for both the design and test cases, kg (lb.). F=


PI ∑I1

162



11. Calculate the individual and combined stresses in the column, and compare them to the allowable values. a.

Operating Conditions Axial compression stress, fb: SI Units f a = 9 810

English Units Co ,kPa A

fa =

Co , psi A

Bending stress, fb: SI Units

English Units

 P e 0.75FL  f b = 9 810  1 + ,kPa Z x   Zx P e 0. 75FL fb = 1 + , psi Zx Zx Where: Zx = e

b.

Section modulus of the column that resists bending, mm3 (in.3).

= Distance between vessel shell and centroid of the column, mm (in.).

Test conditions Axial compression stress, fa: SI Units f a = 9 810

English Units CT ,kPa A

fa =

CT , psi A

Bending stress, fb: SI Units Saudi Aramco DeskTop Standards

163



 W e 0.75FL P1e  f b = 9 810  T + + ,kPa Zx Zx   NZ x English Units

 W e 0.75FL P1e  fb =  T + + ,psi Zx Z x   NZ x c.

Combined compressive stress in the column, f, for the higher of either the operating or test conditions. f = fa + fb, kPa (psi) If the calculated value of "f" is no greater than the smaller of Fa or Fb that were determined in Steps 5 and 6, then the result is acceptable (but conservative). Proceed to Step 12 if this conservative limit is exceeded.

12. Compare the calculated stresses to the allowable values to determine the acceptability of the number and size of the columns selected. •

If

•

If

fa f f ≤ 0.15, then a + b ≤ 1 Fa F a Fb

f Cmf b fa > 0.15, then a + ≤1 Fa Fa  fa  1− F  F'e  b

Where: Cm = 0.85 '

Fe =

12π 2E 2 Kl 23 1   r 

E = Modulus of elasticity, kPa (psi)


164



Note that Fa, Fb and Fe' may be increased by one-third where fa and fb are computed on the basis of wind or seismic loads in combination with dead loads. 13. After the columns have been designed, the attachment details at the vessel shell must be developed. These attachment details must ensure that the vessel shell is not overstressed due to the locally applied loads, and will often include reinforcing pads or rings. This aspect of column support design is beyond the scope of MEX 202.03. 14. Confirm that the column support design that is specified in the Contractor Design Package is in accordance with the preceding procedure. Work Aid 4B: Vertical Vessel on Skirt Support

This procedure may be used to determine the required thickness of the skirt for a vertical pressure vessel. 1.

Determine the total weight of the pressure vessel, attachments, and contents to be included for the case under consideration. This information should be available from the vendor or contractor, or may be calculated using procedures that are contained in PEDP course CSE 110. Wo =

kg (lb.), for operating conditions

WT =

kg (lb.), for hydrotest conditions

We =

kg (lb.), for empty conditions

2.

Calculate or determine the bending moment at the base of the skirt, M, kg-mm (in.-lb.), due to wind (from procedures that are contained in CSE 110).

3.

Calculate or determine the bending moment at the skirt to vessel shell intersection, Ms,as follows: Ms = M - hs (V - 0.5qzDehs), kg-mm (in.-lb.)


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Where: M = As determined in Step 2. V = qzDehs, kg (lb.) hs = Skirt height, mm (in.) De = Effective diameter of skirt, mm (in.) qz = Wind pressure, kg/mm2 (psi) De and qz may be determined using procedures that are contained in CSE 110. Note that for relatively short skirts and tall towers, the difference between Ms and M will not be large and Ms may be taken as equal to M without being overly conservative. 4.

Assume a value of skirt thickness, tsk, mm (in.). A good initial assumption would be to choose tsk equal to the vessel shell thickness at the skirt attachment.

5.

Calculate the longitudinal stress in the skirt at the base, σL. SI Units

English Units

 W 4M  σL = 9810  − ± 2  , kPa  πDsk t sk πDsk t sk 

σL = −

W 4M ± , psi πDskt sk πD2sk tsk

Where: Dsk = Skirt diameter at the base in the middle of its thickness, mm (in.).


M

= Bending moment at the base for the case under consideration, kg-mm (in.-lb.).

W

= Weight of the vessel for the case under consideration, kg (lb.).

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6.

For the skirt material, determine the allowable stress in tension and compression per the ASME Code at ambient temperature, Sal1, kPa (psi). •

Allowable tensile stress determined per Work Aid 2A.

•

Allowable compressive stress determined per Work Aid 3B.

7.

Confirm that σL ≤ Sal1. If it is not, increase tsk and/or Dsk and recalculate.

8.

Determine the weld joint efficiency, E, for the skirt-to-shell weld. See Figure 33.

9.

•

E = 0.55 for Type 1 skirts

•

E = 0.80 for Type 2 skirts

Determine the allowable stresses of the skirt and vessel materials at the vessel design temperature. Use the lower of the two as Sal2, kPa (psi).

10. Recalculate tsk as follows: SI Units

[

(

t sk = 9 810 (W / πDskv ) + 4Ms / πD2skv

)]/ ESal2 , mm

English Units

t sk =

[(W / πD

skv

) + (4Ms / πD2skv )]/ ESal2 , in.

Where: Dskv = Skirt diameter at its attachment to the vessel in the middle of its thickness, mm (in.). 11. Set a uniform skirt thickness that is equal to the larger of that determined in Steps 7 or 10.


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12. Confirm that the skirt support design that is specified in the Contractor Design Package is in accordance with the preceding procedure.

D

D

t

t T.L.

;;; ;;; ;;; ;;; T.L.

Butt weld blends smoothly into head contour tsk

tsk

Dsk = D

(a) Straight

15°max.

(b) Flared

Type 1: Skirt butted to the knuckle portion of the head

Lap weld blends smoothly into shell contour

D

t

;;; ;;; ;;; ;;; 2tsk

1.75tsk

T.L.

15°max.

tsk

Dsk

(a) Straight

(b) Flared

Type 2: Skirt lapped to the cylindrical portion of the shell.

Figure 33: Types of Support Skirts and Skirt-to-Head Welds Saudi Aramco DeskTop Standards

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WORK AID 5:

STEPS FOR COMPLETING A SAFETY INSTRUCTION SHEET FOR A PRESSURE VESSEL

Use the procedural steps that are contained in SAES-A-005, Preparation of Safety Instruction Sheets, to complete a Safety Instruction Sheet for Vessels, Form 2694. The key numbers indicated in the procedure are shown on the edited Form 2694 in Figure 34. SAES-A-005 is contained in Course Handout 2.


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Figure 34: Vessel Safety Instruction Sheet Form 2694 With Number Key Saudi Aramco DeskTop Standards

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GLOSSARY allowable stress

The minimum stress permitted by ASME Code design rules for a material at a specified design temperature.

alloy

An intentional combination of two or more substances, at least one of which is a metal, which exhibits metallic properties. It can be either a mixture of two types of crystalline structures or a solid solution.

ANSI rating

The flange class that is determined from ASME/ANSI B16.5 and that is needed to withstand the design pressure and temperature.

area replacement method

The design approach used by the ASME Code to determine the reinforcement requirements for an opening in a pressure vessel.

ASME Code

The ASME Boiler and Pressure Vessel Code, Section VIII, is the basic design code that is used for pressure vessels. It is composed of Divisions 1 and 2.

buckling

The collapse of a pressure vessel component due to elastic instability that is caused by an external pressure or a compressive stress.

contractor

Within Saudi Aramco usage, the contractor is the company engaged by Saudi Aramco to provide detailed engineering, design, and procurement services for a capital project. For capital projects, the contractor is responsible for specifying the requirements for pressure vessels on data sheets in accordance with industry codes and Saudi Aramco requirements.

corrosion

Deterioration of a material, usually metal, due to its reaction with the environment. Corrosion may be caused either by direct chemical attack or by an electrochemical action.

corrosion allowance

Additional wall thickness that is added to a pressure vessel component to compensate for deterioration during operation.


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creep

A condition that occurs at elevated temperature where continuing deformation takes place without any increase in applied load.

design pressure

The pressure that is used for the mechanical design of a pressure vessel.

design temperature

The temperature that is used for the mechanical design of a pressure vessel.

ductility

The ability of a metal to deform in the plastic range without fracturing under stress.

erosion

The destruction of a metal by the abrasive action of a liquid or vapor or of solid particles that are suspended in the operating liquid or vapor.

fatigue strength

Stress at which failure occurs at some definite number of cycles.

head

The component of a pressure vessel which closes the ends, on the outside, or which separates two sections of a pressure vessel on the inside.

hoop stress

Stress acting in a direction that is perpendicular to the axis of revolution of a pressure vessel.

joint efficiency

A factor that is used to reduce the assumed strength of a welded joint to account for material or structural discontinuities.

knuckle

A curved uniform transition between two different geometries in a pressure vessel. A knuckle may be located between two different sections of a pressure vessel (such as at a cone-to-cylinder junction) or within a single component (such as a formed head).


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lethal service

An application that involves poisonous gases or liquids of such a nature that a very small amount of the gas or of the vapor of the liquid, mixed or unmixed with air, is dangerous to life when inhaled. For purposes of the ASME Code, these gases or liquids include substances of this nature that are stored under pressure or may generate a pressure if stored in a closed vessel.

longitudinal stress

Stress acting in the direction of the axis of revolution of a pressure vessel.

manufacturer

Within Saudi Aramco usage, the manufacturer is the company that supplies equipment items, such as pressure vessels and heat exchangers, to Saudi Aramco. The manufacturer is responsible for the final design, materials, fabrication, inspection, and testing in accordance with the data sheets, the applicable industry codes, and Saudi Aramco requirements. The terms "manufacturer" and "vendor" may be used interchangeably.

maximum allowable working pressure (MAWP)

The maximum gauge pressure that is permissible at the top of a completed vessel, in its operating position, for a designated temperature.

mechanics

The study of the response of matter to various force systems.

membrane stress

The component of normal stress that is uniformly distributed and that is equal to the average value of stress across the thickness of the section under consideration.

membrane theory

For thin-walled cylindrical vessels, internal pressure will cause a uniform stress to occur in the wall.

modulus of elasticity

The ratio of the unit stress to the unit strain for a material.

operating pressure

The maximum internal or external pressure to which a pressure vessel may be exposed.


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operating temperature

The maximum and minimum metal temperature to which a pressure vessel may be exposed.

Poisson's ratio

The ratio of the strain in the lateral direction to the strain in the axial direction.

principal stresses

The maximum and minimum normal stresses acting on planes passing through a solid.

radiographic examination (RT)

A method for detecting imperfections in materials by passing X-ray or nuclear radiation through the material and presenting its image on a recording medium.

solid mechanics

The representation and determination of internal force systems and the deformation of solid bodies subjected to external forces.

strain

The ratio of the elongation of a solid under an applied load to the original length of the solid.

stress

A force per unit area acting on a material.

thermal fatigue

The development of cyclic thermal gradients producing high cyclic thermal stresses and subsequent local cracking of the material.

ultimate strength

The maximum load per unit of original cross-sectional area that a test specimen of a material can sustain before fracture, usually in single tension or compression.

ultrasonic examination

A method for detecting imperfections in materials by passing ultrasonic vibrations (normally 1-5 MHz) through the material.

vendor

Within Saudi Aramco usage, the vendor is the company that supplies equipment items, such as pressure vessels and heat exchangers, to Saudi Aramco. The vendor is responsible for the final design, materials, fabrication, inspection, and testing in accordance with the data sheets, the applicable industry codes, and Saudi Aramco requirements. The terms "vendor" and "manufacturer" may be used interchangeably.


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weld joint efficiency

The efficiency of a welded joint that is expressed as a numerical (decimal) quantity. It is used in the design of a joint as a multiplier of the appropriate allowable stress value taken from the appropriate allowable stress table of the ASME Code.

yield strength

The ordinate of a material's stress-strain curve at which the material has a permanent deformation of 0.002 in./in.


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Specifying Design Requirements for Pressure Vessels

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