GUIDEBOOK
FOR THE DESIGN OF
ASME SECTION VIII
PRESSURE VESSELS
Third Edition
by
James R. Farr
Wadsworth, Ohio
Maan H. Jawad
Camas, Washington
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Library of Congress Cataloging-in-Publication Data
Farr, James R.
Guidebook for the design of ASME Section VIII pressure vessels/by James R. Farr, Maan H. Jawad.—2nd ed.
p. cm.
Includes bibliographical references and index. ISBN 0-7918-0172-1
1. Pressure vessels—Design and construction. 2. Structural engineering. I. Jawad, Maan H.
II. Title.
TA660. T34 F36 2001
681’.76041—dc21 2001046096
British Library Cataloguing in Publication Data
A catalogue record for this book is available from the British Library. Woodhead Publishing ISBN-13: 978-1-84569-119-6
Barbara R. Farr
and dedicated to
her friend and pal,
P
REFACE TO
T
HIRD
E
DITION
The ASME Boiler and Pressure Vessel Code, Section VIII, is a continuously changing document. It con-tains the latest, safe, and economical rules for the design and construction of pressure vessels and pressure vessel components.
The contents of many chapters have been updated to incorporate improvements made in design methods since the book was originally issued. Some of the changes are extensive while others are minor. The brittle fracture rules have been extensively updated and expanded to reflect the latest technology. The heat exchanger chapter has been completely re-written to reflect extensive changes made recently in VIII-1. A new Quick Reference Guide is included in Appendix A to help the reader locate various sections of the lat-est edition of the VIII-1 code. Metric units were added to some of the figures to reflect recent metrication of the ASME code.
James R. Farr Wadsworth, Ohio Maan H. Jawad Camas, Washington December 2005
P
REFACE TO
S
ECOND
E
DITION
The ASME Boiler and Pressure Vessel Code, Section VIII, is a live and progressive document. It strives to contain the latest, safe and economical rules for the design and construction of pressure vessels, pressure ves-sel components, and heat exchangers. A major improvement was made within the last year by changing the design margin on tensile strength from 4.0 to 3.5. This reduction in the margin permits an increase in the allowable stress for many materials with a resulting decrease in minimum required thickness. This was the first reduction in this design margin in 50 years and was based upon the many improvements in material properties, design methods, and inspection procedures during that time.
Chapters and parts of chapters have been updated to incorporate the new allowable stresses and improve-ments which have been made in design methods since this book was originally issued. Some of these changes are extensive and some are minor. Some of the examples in this book have changed completely and some remain unchanged. This book continues to be an easy reference for the latest methods of problem solving in Section VIII. James R. Farr Wadsworth, Ohio Maan H. Jawad St. Louis, Missouri July 2001
P
REFACE TO
F
IRST
E
DITION
The ASME Boiler and Pressure Vessel Code, Section VIII, gives rules that pertain to the design, materials selection, fabrication, inspection, and testing of pressure vessels and their components. With few exceptions, the rules that cover the design of components tend to be complicated to implement. This book was written to demonstrate the application of the design rules to various components. Other rules, such as those pertain-ing to fabrication, inspection, testpertain-ing, and materials are not covered here.
This book is intended as a reference for designers of pressure vessels and heat exchangers. The theoreti-cal background of the equations used here was kept to a minimum, since such background can be obtained from other references. Note also that while the design requirements of such components as shells and heads are interspersed throughout the ASME Code, design requirements pertaining to some specific components are given here in one chapter for easy reference. The emphasis in this book is on solved examples, which illustrate the application of the various equations given in the ASME Section VIII Code.
Chapter 1 of this book covers background information and general topics—such as allowable stresses and joint efficiencies—applicable to all components. Chapter 2 is for the design of cylindrical shells under internal and external loads. Chapter 3 covers the design of dished heads and transition sections that are under internal and external loads. Chapter 4 considers flat plates, covers, and flanges. Openings are reviewed in Chapter 5, and Chapter 6 covers special components of VIII-1, such as stayed construction, jacketed components, half-pipe jackets, and noncircular vessels. Chapter 7 covers heat exchangers, and Chapter 8 covers stress categories, fatigue, and other special analysis of components.
James R. Farr Wadsworth, Ohio Maan H. Jawad St. Louis, Missouri January 1998
A
CKNOWLEDGMENTS
We are indebted to many people and organizations for their help in preparing this book and revisions and for permitting us to use material supplied by them. Thanks is given to our former co-workers, fellow ASME Committee Members, and many other persons who have given us suggestions and corrections. Special thanks is given to Mr. Mike Bytnar of the Nooter Corporation for providing us with the front cover photograph, to Mr. Thomas P. Pastor of HSB Global Standards for permitting us to use their new Quick Reference Guide for VIII-1, and to our editors Mary Grace Stefanchik and Tara Smith at ASME Press for their continued guidance.
T
ABLE OF
C
ONTENTS
Preface to Third Edition ...v
Preface to Second Edition ...vii
Preface to First Edition ...ix
Acknowledgments ...xi
List of Figures...xvii
List of Tables ...xxi
Chapter 1 Background Information...1
1.1 Introduction ...1
1.2 Allowable Stresses...2
1.3 Joint Efficiency Factors ...4
1.4 Brittle Fracture Considerations...9
1.5 Fatigue Requirements ...19
1.6 Pressure Testing of Vessels and Components...23
1.6.1 ASME Code Requirements ...23
1.6.2 What Does a Hydrostatic or Pneumatic Pressure Test Do?...23
1.6.3 Pressure Test Requirements for VIII-1...23
1.6.4 Pressure Test Requirements for VIII-2...25
Chapter 2 Cylindrical Shells ...27
2.1 Introduction ...27
2.2 Tensile Forces, VIII-1 ...27
2.2.1 Thin Cylindrical Shells ...27
2.2.2 Thick Cylindrical Shells...33
2.3 Axial Compression ...35
2.4 External Pressure ...41
2.4.1 External Pressure for Cylinders with DO/t ≥ 10 ...41
2.4.2 External Pressure for Cylinders with DO/t < 10 ...44
2.4.4 Stiffening Rings...46
2.4.5 Attachment of Stiffening Rings ...49
2.5 Cylindrical Shell Equations, VIII-2...51
2.6 Miscellaneous Shells ...52
2.6.1 Mitered Cylinders...52
2.6.2 Elliptical Shells ...53
Chapter 3 Spherical Shells, Heads, and Transition Sections ...55
3.1 Introduction ...55
3.2 Spherical Shells and Hemispherical Heads, VIII-1...55
3.2.1 Internal Pressure in Spherical Shells and Pressure on Concave Side of Hemispherical Heads ...55
3.2.2 External Pressure in Spherical Shells and Pressure on Convex Side of Hemispherical Heads ...59
3.3 Spherical Shells and Hemispherical Heads, VIII-2...61
3.4 Ellipsoidal Heads, VIII-1...62
3.4.1 Pressure on the Concave Side ...62
3.4.2 Pressure on the Convex Side...63
3.5 Torispherical Heads, VIII-1 ...65
3.5.1 Pressure on the Concave Side ...65
3.5.2 Pressure on the Convex Side...67
3.6 Ellipsoidal and Torispherical Heads, VIII-2...68
3.7 Conical Sections, VIII-1 ...70
3.7.1 Internal Pressure...71
3.7.2 External Pressure...80
3.8 Conical Sections, VIII-2 ...89
3.9 Miscellaneous Transition Sections ...93
Chapter 4 Flat Plates, Covers, and Flanges...97
4.1 Introduction ...97
4.2 Integral Flat Plates and Covers...97
4.2.1 Circular Flat Plates and Covers...97
4.2.2 Noncircular Flat Plates and Covers...99
4.3 Bolted Flat Plates, Covers, and Flanges ...101
4.3.1 Gasket Requirements, Bolt Sizing, and Bolt Loadings ...101
4.4 Flat Plates and Covers With Bolting ...102
4.4.1 Blind Flanges & Circular Flat Plates and Covers ...102
4.4.2 Noncircular Flat Plates and Covers...102
4.5 Openings in Flat Plates and Covers...102
4.5.1 Opening Diameter Does Not Exceed Half the Plate Diameter...103
4.5.2 Opening Diameter Exceeds Half the Plate Diameter ...104
4.6 Bolted Flange Connections With Ring Type Gaskets ...104
4.6.1 Standard Flanges ...105
4.7 Spherically Dished Covers ...118
4.7.1 Definitions and Terminology...118
4.7.2 Types of Dished Covers ...119
Chapter 5 Openings ...125
5.1 Introduction ...125
5.2 Code Bases for Acceptability of Opening ...125
5.3 Terms and Definitions...126
5.4 Reinforced Openings—General Requirements ...126
5.4.1 Replacement Area ...126
5.4.2 Reinforcement Limits...126
5.5 Reinforced Opening Rules, VIII-1 ...128
5.5.1 Openings With Inherent Compensation ...128
5.5.2 Shape and Size of Openings ...128
5.5.3 Area of Reinforcement Required ...129
5.5.4 Limits of Reinforcement ...131
5.5.5 Area of Reinforcement Available...132
5.5.6 Openings Exceeding Size Limits of Section 5.5.2.2 ...142
5.6 Reinforced Opening Rules, VIII-2 ...144
5.6.1 Definitions ...144
5.6.2 Openings Not Requiring Reinforcement Calculations ...144
5.6.3 Shape and Size of Openings ...146
5.6.4 Area of Reinforcement Required ...146
5.6.5 Limits of Reinforcement ...146
5.6.6 Available Reinforcement ...147
5.6.7 Strength of Reinforcement Metal...148
5.6.8 Alternative Rules for Nozzle Design ...148
5.7 Ligament Efficiency Rules, VIII-1 ...154
Chapter 6 Special Components, VIII-1...159
6.1 Introduction ...159
6.2 Braced and Stayed Construction ...159
6.2.1 Braced and Stayed Surfaces ...159
6.2.2 Stays and Staybolts ...162
6.3 Jacketed Vessels...163
6.3.1 Types of Jacketed Vessels ...163
6.3.2 Design of Closure Member for Jacket to Vessel...164
6.3.3 Design of Openings in Jacketed Vessels ...165
6.4 Half-Pipe Jackets ...168
6.4.1 Maximum Allowable Internal Pressure in Half-Pipe Jacket...169
6.4.2 Minimum Thickness of Half-Pipe Jacket...171
6.5 Vessels of Noncircular Cross Section...175
6.5.1 Types of Vessels ...175
6.5.3 Openings in Vessels of Noncircular Cross Section...180
6.5.4 Vessels of Rectangular Cross Section ...185
Chapter 7 Design of Heat Exchangers ...189
7.1 Introduction ...189
7.2 Design of Tubesheets in U-Tube Exchangers ...189
7.2.1 Nomenclature ...189 7.2.2 Preliminary Calculations ...194 7.2.3 Design Equations ...195 7.3 Fixed Tubesheets ...208 7.3.1 Nomenclature ...208 7.3.2 Preliminary Calculations ...209 7.3.3 Design Equations ...212 7.4 Additional Rules ...223
7.5 Methods of Attaching Tubes-to-Tubesheet...223
7.5.1 Push Out Tests...227
7.5.2 Tube-to-Tubesheet Locking Mechanism...228
7.6 Expansion Joints ...229
Chapter 8 Analysis of Components in VIII-2...233
8.1 Introduction ...233 8.2 Stress Categories...233 8.3 Stress Concentration ...239 8.4 Combinations of Stresses ...239 8.5 Fatigue Evaluation ...244 References ...249 Appendices...251
Appendix A—Guide to VIII-1 Requirements...251
Appendix B—Material Designation ...255
Appendix C—Joint Efficiency Factors ...257
Appendix D—Flange Calculation Sheets...279
Appendix E—Conversion Factors ...285
L
IST OF
F
IGURES
Figure Number
1.1 Welded Joint Categories (ASME VII-1) ...5
1.2 Category C Weld ...9
E1.1 ...9
1.3 Some Governing Thickness Details Used for Toughness (ASME VIII-1) ...12
1.4 Impact-Test Exemption Curves (ASME VIII-1) ...15
1.5 Charpy Impact-Test Requirements for Full Size Specimens for Carbon and Low Alloy Steels With Tensile Strength of Less Than 95 ksi (ASME VIII-1) ...16
1.6 Reduction of MDMT Without Impact Testing (ASME VIII-1)...17
1.7 Fatigue Curves for Carbon, Low Alloy, Series 4XX, High Alloy Steels, and High Tensile Steels for Temperatures Not Exceeding 700∞F (ASME VIII-2)...18
2.1 ...28
2.2 Comparison of Equations for Hoop Stress in Cylindrical Shells ...29
2.3 ...30
2.4 Chart for Carbon and Low Alloy Steels With Yield Stress of 30 ksi and Over, and Types 405 & 410 Stainless Steels ...37
2.5 C Factor as a Function of R/T (Jawad, 2004)...38
E2.8 ...39
2.6 Geomatic Chart for Cylindrical Vessels Under Extrenal Pressure (Jawad and Farr, 1989) ...42
2.7 Some Lines of Support of Cylindrical Shells Under External Pressure (ASME VIII-1)...43
E2.13 ...48
2.8 Some Details for Attaching Stiffener Rings (ASME VIII-1)...50
2.9 MiteredBend...53 2.10 Elliptical Cylinder ...54 3.1 ...57 E3.4 ...59 3.2 ...63 3.3 ...66 3.4 ...67 3.5 ...69 3.6 ...71 3.7 ...74 E3.11 ...75 E3.12 ...78 E3.13 ...86 3.8 ...89
3.10 Values of Q for Large End of Cone-to-Cylinder Junction (ASME VIII-2)...91
3.11 Inherent Reinforcement for Small End of Cone-to-Cylinder Junction (ASME VIII-2)...92
3.12 Values of Q for Small End of Cone-to-Cylinder Junction (ASME VIII-2)...93
3.13 ...94
4.1 Some Acceptable Types of Unstayed Flat Heads and Covers (ASME VIII-1) ...99
4.2 Multiple Openings in the Rim of a Flat Head or Cover With a Large Central Opening (ASME VIII-1) ...104
E4.5 Ring Flange Sample Calculation Sheet...106
E4.6 Welding Neck Flange Sample Calculation Sheet ...109
E4.7 Reverse Welding Neck Flange Sample Calculation Sheet...114
4.3 Spherically Dished Covers With Bolting Flanges (ASME VIII-1) ...119
E4.8 Example Problem of Spherically Dished Cover, Div. 1 ...122
5.1 Reinforcement Limits Parallel to Shell Surface...127
5.2 Chart for Determining Value of F for Angle q (ASME VIII-1 and VIII-2) ...130
5.3 Determination of Special Limits for Setting t,rfor Use in Reinforcement Calculations ...131
E5.1 Example Problem of Nozzle Reinforcement in Ellipsoidal Head, Div. 1 ...132
E5.2 Example Problem of Nozzle Reinforcement of 12 in. ¥ 16 in. Manway Opening, Div. 1 ...135
E5.3.1 Example Problem of Nozzle Reinforcement of Hillside Nozzle, Div. 1...137
E5.3.2 Example Problem of Nozzle Reinforcement of Hillside Nozzle, Div. 1...138
E5.4 Example Problem of Nozzle Reinforcement With Corrosion Allowance, Div. 1 ...140
5.4.1 (ASME VIII-1) ...143
5.4.2 (ASME VIII-1) ...143
5.5 Nozzle Nomenclature and Dimensions (Depicts General Configurations Only) (ASME VIII-2) ...145
5.6 Limits of Reinforcing Zone for Alternative Nozzle Design (ASME VIII-2) ...149
E5.5 Example Problem of Nozzle Reinforcement in Ellipsoidal Head, Div. 2 ...149
E5.6 Example Problem of Nozzle Reinforcement of 12 in. ¥ 16 in. Manway Opening, Div. 2 ...152
E5.7 Example Problem of Nozzle Reiforcerment of Series of Openings Div. 1 ...156
6.1 Typical Forms of Welded Staybolts (ASME VIII-1) ...161
6.2 Typical Welded Stay for Jacketed Vessel (ASME VIII-1) ...161
6.3 Some Acceptable Types of Jacketed Vessels (ASME VIII-1)...164
6.4 Some Acceptable Types of Closure Details (ASME VIII-1) ...166
6.5 Some Acceptable Types of Penetration Details (ASME VIII-1) ...170
6.6 Spiral Jackets, Half-Pipe and Other Shapes...171
6.7 Factor K for NFS 2 Pipe Jacket ...172
6.8 Factor K for NFS 3 Pipe Jacket ...173
6.9 Factor K for NFS 4 Pipe Jacket ...174
6.10 Vessels of Rectangular Cross Section ...176
6.11 Vessels of Rectangular Cross Section With Stay Plates ...178
6.12 Vessels of Obround Cross Section With and Without Stay Plates and Vessels of Circular Cross Section With a Stay Plate ...179
6.13 Plate With Constant-Diameter Openings of Same or Different Diameters ...180
6.14 Plate With Multidiameter Openings...181
E6.8 Example Problem of Noncircular Vessels (ASME VIII-1)...186
7.1 Various Heat-Exchanger Configurations (TEMA, 1999)...190
7.2a Tubesheet Integral with Shell and channel (ASME. VIII-1) ...191
7.2b Tubesheet Integral with Shell and Gasketed with channel, Extended as Flange (ASME. VIII-1) ...192
7.2c Tubesheet Integral with Shell and Gasketed with Channel, not Extended as a Flange (ASME. VIII-1) ...193
7.2d Tubesheet Gasketed with Shell and Channel (ASME. VIII-1) ...194
7.2e Tubesheet Gasketed with Shell and Integral with Channel, Extended as a Flange (ASME. VII-1) ...195
7.2f Tubesheet Gasketed with Shell and Integral with channel, not Extended as a Flange (ASME. VIII-1) ...196
7.3 Tubesheet Geometry (ASME. VIII-1)...197
7.4 Typical Untubed Lane Configurations (ASME VIII-1) ...198
7.5 Curves for the Determination of E*/E and v* (ASME. VIII-1) ...199
7.6a Tubesheet Integral with Shell and Channel (ASME. VIII-1) ...210
7.6b Tubesheet integral with Shell and Gasketed with Channel, Extended as a Flange (ASME. VIII-1) ...211
7.6c Tubesheet Integral with Shell and Gasketed with Channel, not Extended as a Flange (ASME. VIII-1) ...212
7.6d Tubesheet Gasketed with Shell and Channel (ASME. VIII-1) ...213
7.7 Xavs.Zd, Zv, or Zm(ASME.VIII-1) ...214
7.8 Xavs. Fmwith negative Q3(ASME. VIII-1) ...215
7.9 Xavs.Fmwith positive Q3(ASME. VIII-1) ...222
7.10 Some Acceptable Types of Tube-to-Tubesheet Attachments (ASME VIII-1)...225
7.11 Some Acceptable Types of Tube-to-Tubesheet Strength Welds (ASME VIII-1) ...226
7.12 Schematic Detail of a Hydraulic Tube Expander ...000
7.13 Typical Detail of Grooves in a Clad Tubesheet ...227
7.14 Some Bellows-Type Expansion Joints ( ASME VIII-1) ...230
7.15 Some Flanged and Flued Expansion Joints ( ASME VIII-l) ...231
E8.1 ...238
8.1 Linearizing Stress Distribution ...240
E8.4 Model of a Finite Element Layout in a Flat Head-to-Shell Junction ...243
8.2 Fatigue Curves for Carbon, Low Alloy, 4XX High Alloy, and High Strength Steels for Temperatures Not Exceeding 700∞F (ASME VIII-2) ...245
8.3 Cyclic Curves ...246 A.1 ...000 C.1 ...257 C.2 ...258 C.3 ...259 C.4 ...260 C.5 ...261 C.6 ...262 C.7 ...263 C.8 ...264 C.9 ...265 C.10 ...266 C.11 ...267 C.12 ...268 C.13 ...269 C.14 ...270 C.15 ...271 C.16 ...272 C.17 ...273 C.18 ...274 C.19 ...275 C.20.E ...276
D.1 Fig. D.1—Ring Flange With Ring-Type Gasket...279
D.2 Fig. D.2—Slip-On or Lap-joint Flange With Ring-Type Gasket ...280
D.3 Fig. D.3—Welding Neck Flange With Ring-Type Gasket ...281
D.4 Fig. D.4—Reverse Welding Neck Flange With Ring-Type Gasket...282
D.5 Fig. D.5—Slip-On Flange With Full-Face Gasket...283
L
IST OF
T
ABLES
Table Number
1.1.1 Criteria for Establishing allowable Stress Values for VIII-1 Except for UCI, UCD,
and ULT Material and for Bolting (ASME II-D)...2
1.1.2 Criteria for Establishing allowable Stress Values for VIE-1 for UCI, UCD, and ULT Material (ASME II-D)...3
1.2 Criteria for Establishing Design Stress Intensity Values for VIII-2 Except for bolting (ASME II-D)...3
1.3 Stress Values for SA-515 and SA-516 Materials ...4
1.4 Allowable Stress Values for Welded Connections...5
1.5 Maximum Allowable Efficiencies for Arc- and Gas-Welded Joints (ASME VIII-1) ...6
El.1 Stress Categories...8
1.6 Assignment of Materials to Curves (ASME VIII-1) ...10
1.7 Minimum Design Metal Temperatures in High Alloy Steels Without Impact Testing ...18
2.1 Tabular Values for Fig. 2.4 ...36
3.1 Factor KOfor an Ellipsoidal Head With Pressure on the Convex Side...64
3.2 Values of D for Junctions at the Large Cylinder Due to Internal Pressure...72
3.3 Values of D for Junctions at the Small Cylinder Due to Internal Pressure...73
3.4 Values of D for Junctions at the Large Cylinder Due to External Pressure...80
E3.14 Allowable Stress and Pressure Data ...90
6.1 Example of Pressure Used for Design of Components...163
6.2 Closure Detail Requirements for Various Types of Jacket Closures...169
6.3 Penetration Detail Requirements ...171
7.1 Joint Efficiencies frfor various tube-to-tubesheet attachments (ASME VIII-1)...228
8.1 Primary Stress Category ...234
8.2 Structural Discontinuity...235
8.3 Thermal Stress ...235
8.4 Stress Categories and Their Limits (ASME VIII-2)...236
8.5 Classification of Stresses (ASME VIII-2) ...237
8.6 Some Stress Concentration Factors Used in Fatigue ...239
E8.4 Summary of Finite Element Output ...244
B.1 Carbon Steel Plate ...235
B.2 Chrome-Moly Steel Plate Specifications, SA-387 ...235
B.3 Chrome-Moly Steel Forging Specifications, SA-182...235
B.4 Chrome-Moly Steel Forging Specifications, SA-336...256
1
B
ACKGROUND
I
NFORMATION
1.1 INTRODUCTION
In this chapter some general concepts and criteria pertaining to Section VIII are discussed. These include allowable stress, factors of safety, joint efficiency factors, brittle fracture, fatigue, and pressure testing. Detailed design and analysis rules for individual components are discussed in subsequent chapters.
Since frequent reference will be made to ASME Section VIII Divisions 1 and 2, the following designation will be used from here on to facilitate such references. ASME Section VIII, Division 1 Code will be desig-nated by VIII-1. Similarly, VIII-2 will designate the ASME Section VIII, Division 2 Code. Other ASME code sections such as Section II Part D will be referred to as II-D. Equations and paragraphs referenced in each of these divisions will be called out as they appear in their respective Code Divisions.
Many designs rules in VIII-1 and VIII-2 are identical. These include flange design and external pressure requirements. In such cases, the rules of VIII-1 will be discussed with a statement indicating that the rules of VIII-2 are the same. Appendix A at the end of this book lists the paragraph numbers in VIII-1 that pertain to various components of pressure vessels.
Section VIII requires the fabricator of the equipment to be responsible for its design. Paragraphs UG-22 in VIII-1 and AD-110 in VIII-2 are given to assist the designer in considering the most commonly encoun-tered loads. They include pressure, wind forces, equipment loads, and thermal considerations. When the designer takes exceptions to these loads either because they are not applicable or they are unknown, then such exceptions must be stated in the calculations. Similarly, any additional loading conditions considered by the designer that are not mentioned in the Code must be documented in the design calculations. Paragraphs U-2(a) and U-2(b) of VIII-1 give guidance for some design requirements. VIII-2, paragraph AD-110 and the User’s Design Specifications mentioned in AG-301 provide the loading conditions to be used by the manufacturer.
Many design rules in VIII-1 and VIII-2 are included in the Appendices of these codes. These rules are for specific products or configurations. Rules that have been substantiated by experience and used by industry over a long period of time are in the Mandatory Appendices. New rules or rules that have limited applica-tions are placed in the Non-Mandatory Appendices. Non-Mandatory rules may eventually be transferred to the Mandatory section of the Code after a period of use and verification of their safety and practicality. However, guidance-type appendices will remain in the Non-Mandatory section of the Code.
The rules in VIII-1 do not cover all applications and configurations. When rules are not available, Paragraphs U-2(d), U-2(g), and UG-101 must be used. Paragraph U-2(g) permits the engineer to design com-ponents in the absence of rules in VIII-1. Paragraph UG-101 is for allowing proof testing to establish maxi-mum allowable working pressure for components. In VIII-2 there are no rules similar to those in UG-101, since VIII-2 permits design by analysis as part of its requirements. This is detailed in Paragraphs AD-100(b), AD-140, AD-150, and AD-160 of VIII-2.
Many pressure vessel designers and vessel users are of the opinion that a vessel or component will be thin-ner, and subsequently lower cost, using the VIII-2 rules compared with using the VIII-1 rules. This opinion
is usually based on the design factor on tensile strength of 3.5 for VIII-1 while the factor for VIII-2 is lower at 3.0. The design factor on yield strength for both VIII-1 and VIII-2 is 1.5. However, as can be seen from a comparison of some design calculations for the same geometry and loadings, the thickness required for a VIII-2 component may be greater than that required for a VIII-1 component. (For example, compare the cal-culations of an opening in a 2:1 ellipsoidal head given in Example Design Problems 5.1 and 5.5). Often, this is caused by differences in design formulas and analysis methods as well as the design criteria and allowable stresses. For many materials, the allowable stress is set by the yield strength criteria rather than by the ten-sile strength criteria and, consequently, the allowable stress value at a temperature may be the same.
1.2 ALLOWABLE STRESSES
Except for UCI, UCD, and ULT material and for bolting, the criteria for establishing allowable stress values for VIII-1 is detailed in Appendix 1 of II-D and summarized in Table 1.1.1. For UCI, UCD, and ULT mate-rial, the criteria for establishing allowable stress values for VIII-1 is detailed in Appendix P of VIII-1 and summarized in Table 1.1.2. The criteria for establishing the design stress intensity values for VIII-2 is detailed in Appendix 2 of II-D and summarized in Table 1.2. The criteria for establishing allowable stress values for bolting for both VIII-1 and VIII-2 is detailed in Appendix 2 of II-D. The allowable stress at design temperature for most materials is the lessor of 1/3.5 the minimum effective tensile strength or 2/3 the mini-mum yield stress of the material for temperatures below the creep and rupture values. The controlling allow-able stress for most bolts is 1/5 the tensile strength. The minimum effective tensile stress at elevated temperatures is obtained from the actual tensile stress curve with some adjustments. The tensile stress value obtained from the actual curve at a given temperature is multiplied by the lessor of 1.0 or the ratio of the minimum tensile stress at room temperature obtained from ASTM Specification for the given material to the actual tensile stress at room temperature obtained from the tensile strength curve. This quantity is then mul-tiplied by the factor 1.1. The effective tensile stress is then equal to the lessor of this quantity or the
mini-TABLE 1.1.1
CRITERIA FOR ESTABLISHING ALLOWABLE STRESS VALUES FOR VIII-1 EXCEPT FOR UCI, UCD, AND ULT MATERIAL AND FOR BOLTING (ASME II-D)
Nomenclature
RT= ratio of the average temperature dependent trend curve value of tensile strength to the room temperature tensile strength
RY= ratio of the average temperature dependent trend curve value of yield strength to the room temperature yield strength
SRavg= average stress to cause rupture at the end of 100,000 hr
SRmin= minimum stress to cause rupture at the end of 100,000 hr
SC= average stress to produce a creep rate of 0.01 %/1000 hr
ST= specified minimum tensile strength at room temperature, ksi
mum tensile stress at room temperature given in ASTM. This procedure is illustrated in example 4.1 of Jawad and Farr (reference 14, found at back of book).
The 1.1 factor discussed above is a constant established by the ASME Code Committee. It is based on engi-neering judgment that takes into consideration many factors. Some of these include increase in tensile strength for most carbon and low alloy steels between room and elevated temperature; the desire to maintain a constant allowable stress level between room temperature and 500°F or higher for carbon steels; and the adjustment of minimum strength data to average data. Above approximately 500°f or higher the allowable stress for carbon steels is controlled by creep-rupture rather than tensile-yield criteria. Some materials may not exhibit such an increase in tensile stress, but the criterion for 1.1 is still applicable to practically all materials in VIII-1.
Table 1.1.1 also gives additional criteria for creep and rupture at elevated temperatures. The criteria are based on creep at a specified strain and rupture at 100,000 hours. The 100,000 hours criterion for rupture cor-responds to about eleven years of continual use. However, VIII-1 does not limit the operating life of the equipment to any specific number of hours.
The allowable stress criteria in VIII-2 are given in II-D of the ASME Code. The allowable stress at the design temperature for most materials is the smaller of 1/3 the tensile strength or 2/3 the yield stress. The design temperature for all materials in VIII-2 is kept below the creep and rupture values. Table 1.2 summa-rizes the allowable stress criteria in VIII-2.
A sample of the allowable stress Tables listed in Section II-D of the ASME Code is shown in Table 1.3. It lists the chemical composition of the material, its product form, specification number, grade, Unified
TABLE 1.1.2
CRITERIA FOR ESTABUSHING ALLOWABLE STRESS VALUES FOR VIII-1 FOR UCI, UCD, AND ULT MATERIAL (ASME VIII-1)
TABLE 1.2
CRITERIA FOR ESTABLISHING DESIGN STRESS INTENSITY VALUES FOR VIII-2 EXCEPT FOR BOLTING (ASME II-D)
Numbering System (UNS), size, and temper. This information, with very few exceptions, is identical to that given in ASTM for the material. The Table also lists the P and Group numbers of the material. The P num-bers are used to cross reference the material to corresponding welding processes and procedures listed in Section IX, “Welding and Brazing Qualifications,” of the ASME Code. The Table also lists the minimum yield and tensile strengths of the material at room temperature, maximum applicable temperature limit, External Pressure Chart reference, any applicable notes, and the stress values at various temperatures. The designer may interpolate between listed stress values, but is not permitted to extrapolate beyond the pub-lished values.
Stress values for components in shear and bearing are given in various parts of VIII-1, VIII-2, as well as II-D. Paragraph UW-15 of VIII-1 and AD-132 of VIII-2 lists the majority of these values. A summary of the allowable stress values for connections is shown in Table 1.4.
Some material designations in ASTM as well as the ASME Code have been changed in the last 20 years. The change is necessitated by the introduction of subclasses of the same material or improved properties. Appendix B shows a cross reference between older and newer designations of some common materials.
The maximum design temperatures allowed in VIII cannot exceed those published in Section II-D. VIII-1 defines design temperature as the mean temperature through the cross section of a component. VIII-2 defines design temperature as the mean temperature in the cross section of a component, but the surface temperature cannot exceed the highest temperature listed in II-D for the material. This difference in the definition of tem-perature in VIII-1 and VIII-2 can be substantial in thick cross sections subjected to elevated temtem-peratures.
1.3 JOINT EFFICIENCY FACTORS
In the ASME Boiler Code, Section I, as well as in VIII-2, all major longitudinal and circumferential butt joints must be examined by full radiography, with few exceptions. VIII-1, on the other hand, permits various levels of examination of these major joints. The examination varies from full radiographic to visual, depend-ing on various factors specified in VIII-1 and by the user. The degree of examination influences the required
TABLE 1.3
STRESS VALUES FOR SA-515 AND SA-516 MATERIALS
Line Nominal Product Spec Type/ Alloy Desig./ Class Cond./ Size/ Group No. Composition Form No. Grade UNS No. Temper Thick, in. P-No. No.
19 CS Plate SA-515 70 K03101 … … 1 2
20 CS Plate SA-516 70 K02700 … … 1 2
Applic. & max. Min. Min. Temp. Limits
Tensile Yield (NP = Not Permitted) External Notes Line Strength Stress (SPT = Supports only) Pressure
No. ksi ksi I III VIII-1 Chart No.
19 70 38 1000 700 1000 CS-2 G10, S1.T2
20 70 38 850 700 1000 CS-2 G10, S1.T2
Maximum Allowable Stress, ksi, for Metal Temperature,°F, Not Exceeding Line –20 to
No. 100 150 200 300 400 500 600 650 700 750 800 850 900
19 20.0 20.0 20.0 20.0 20.0 20.0 19.4 18.8 18.1 14.8 12.0 9.3 6.7
20 20.0 20.0 20.0 20.0 20.0 20.0 19.4 18.8 18.1 14.8 12.0 9.3 6.7
Note:
thickness through the use of Joint Efficiency Factors, E. The Joint Efficiency Factors, which are sometimes referred to as Quality Factors or weld efficiencies, serve as stress multipliers applied to vessel components when some of the joints are not fully radiographed. These multipliers result in an increase in the factor of safety as well as the thickness of these components. In essence, VIII-1 vessels have variable factors of safety, depending on the degree of radiographic examination of the main vessel joints. As an example, fully radi-ographed longitudinal butt-welded joints in cylindrical shells have a Joint Efficiency Factor, E, of 1.0. This factor corresponds to a safety factor of 3.5 in the parent material. Nonradiographed longitudinal butt-welded joints have an E value of 0.70. This reduction in Joint Efficiency Factor corresponds to a factor of safety of 5.0 in the plates. This higher factor of safety due to a nonradiographed joint results in a 43% increase in the required thickness over that of a fully radiographed joint.
TABLE 1.4
ALLOWABLE STRESS VALUES FOR WELDED CONNECTIONS VIII-1
Component Type of Stress Stress Value Reference
Fillet weld tension 0.55S* UW-18(d)
Fillet weld shear 0.49S UW-15(c)
Groove weld tension 0.74S UW-15(c)
Groove weld shear 0.60S UW-15(c)
Nozzle neck shear 0.70S UG-45(c)
Dowel bolts shear 0.80S II-D
Any location bearing 1.60S II-D
* S = allowable stress for VIII-1 construction
VIII-2
Component Type of Stress Stress Value Reference
Fillet weld tension 0.5Sm* AD-920
Fillet weld shear 0.5Sm* AD-920
Groove weld tension 0.75Sm AD-920
Groove weld shear 0.75Sm AD-920
Nozzle neck shear 0.6Sm AD-132.2
Any location bearing Sy AD-132.1
* Sm= stress intensity values for VIII-2 construction
FIG. 1.1
T
ABLE 1.5
MAXIMUM ALLO
W
ABLE JOINT EFFICIENCIES
1,5
FOR ARC- AND GAS-WELDED JOINTS (ASME
VIII-1)
T
ABLE 1.5
ASME VIII-1 identifies four joint categories that require E factors. They are Categories A, B, C, and D as shown in Fig. 1.1. Category A joints consist mainly of longitudinal joints as well as circumferential joints between hemispherical heads and shells. Category B joints are the circumferential joints between various components as shown in Fig. 1.1, with the exception of circumferential joints between hemispherical heads and shells. The attachment of flanges to shells or heads is a Category C joint. The attachment of nozzle necks to heads, shells, and transition sections is categorized as a Category D joint.
The four joint categories in VIII-1 do not apply to items such as jacket closure bars, tubesheet attachments, and ring girders. The degree of examination of the welds attaching these components to the shell or head is not covered in VIII-1. Most designers assign an E value of 1.0 when calculating the shell or head thickness at such junctions. This is justified since in most cases the strain in the hoop direction, and hence hoop stress, is close to zero at the junction due to the restraint of tubesheet or bars.
The type of construction and joint efficiency associated with each of joints A, B, C, and D is given in Table 1.5. The categories refer to a location within a vessel rather than detail of construction. Thus, a Category C weld, which identifies the attachment of a flange to a shell, can be either fillet, corner, or butt welded, as illus-trated in Fig. 1.2. The Joint Efficiency Factors apply only to the butt-welded joint in sketch (c). The factors do not apply to sketches (a) and (b) since they are not butt welded.
The Joint Efficiency Factors used to design a given component are dependent on the type of examination performed at the welds of the component. As an example, the Joint Efficiency Factor in a fully radiographed longitudinal seam of a shell course is E = 1.0. However, this number may have to be reduced, depending on the degree of examination of the circumferential welds at either end of the longitudinal seam. Appendix B shows some typical components and their corresponding Joint Efficiency Factors.
Example 1.1 Problem
Determine the category and Joint Efficiency Factor of the joints in the heat exchanger shown in Fig. E1.1. The channel side is spot radiographed. The longitudinal seam, b, is a single-welded butt joint with a backup bar. The shell side is not radiographed. The longitudinal and circumferential seams m and 1 are single-welded butt joints with backup bars. The jacket longitudinal seam, n, is a single-single-welded butt joint, without backup bar.
Solution
The joint categories of the various joints can be tabulated as given in Table E1.1: TABLE E1.1
STRESS CATEGORIES
Joint Location Category Joint Efficiency
(a) Channel-to-flange connection C Does not apply
(b) Longitudinal channel seam A 0.80
(c) Channel-to-tubesheet weld C Does not apply
(d) Nozzle-to-channel weld D Does not apply
(e) Flange-to-nozzle neck C Does not apply
(f) Pass partition-to-tubesheet weld None Does not apply. See also UW-15(c) and UW-18(d)of VIII-1
(g) Tube-to-tubesheet weld None Does not apply. See also UW-20 of VIII-1
(h) Shell-to-tubesheet weld G Does not apply
(i) Jacket bar-to-inner-shell weld None Does not apply
(j) Jacket to bar-to-outer-shell weld None Does not apply
(k) Nozzle-to-jacket weld D Does not apply
(l) Longitudinal shell seam A 0.65
(m) (m)Head-to-shell seam B 0-65
(n) Longitudinal jacket seam A 0.60
1.4 BRITTLE FRACTURE CONSIDERATIONS
Both VIII-1 and VIII-2 require the designer to consider brittle fracture rules as part of the material and design selection. The rules for carbon steels are extensive and are discussed first. VIII-1 has two options regarding toughness requirements for carbon steels. The first is given in Paragraph UG-20(f) and allows the designer to exempt the material of construction from impact testing when all of the following criteria are met:
1. The material is limited to P-No.1, Gr. No. 1 and 2, and thickness shall not exceed that given in (a) or (b) below:
FIG. 1.2 CATEGORY C WELD
(a) 1/
2in. for materials listed in Curve A in Table 1.6.
(b) 1 in. for materials listed in Curve B, C, or D in Table 1.6.
2. The completed vessel shall be hydrostatically tested per UG-99(b) or (c) or to 27-4.
3. Design temperature between –20°F and 650°f, incl. Occasional operating temperatures colder than –20°f lare acceptable when due to lower seasonal atmospheric temperatures.
4. The thermal and mechanical shock loadings are not a controlling design requirement. 5. Cyclical loading is not a controlling design requirement.
The above requirements are intended for, but not limited to, relatively thin carbon steel vessels operating in a service that is neither severe in thermal and pressure cycling nor in extreme cold temperatures. Vessels
TABLE 1.6
of low alloy steel or those with carbon steel operating beyond the scope of Paragraph UG-20(f) require an evaluation for brittle fracture in accordance with the rules of UCS-66. The procedure consists of
1. Determining the governing thickness in accordance with Fig. 1.3.
2. Using Fig. 1.4 to obtain the temperature that exempts the material from impact testing. If the specified Minimum Design Metal Temperature, MDMT, is colder than that obtained from the figure, then impact testing in accordance with Fig. 1.5 is required. The specified MDMT is usu-ally given by the user, while the calculated MDMT is obtained from VIII-1. The calculated MDMT is kept equal to or colder than the specified MDMT.
3. The temperature obtained from Fig. 1.4 may be reduced in accordance with Fig 1.6 if the component operates at a reduced stress. This is detailed in Paragraph UCS-66(b) of VIII-1.
FIG. 1.3
At a ratio of 0.35 in Fig. 1.6, the permitted temperature reduction drops abruptly. At this ratio, the stress in a component is about 6000 psi. At this stress level, experience has shown that brittle fracture does not occur regardless of temperature level.
4. The rules in VIII-1 also allow a 30°F reduction in temperature below that obtained from Fig. 1.4 when the component is post-weld heat treated but is not otherwise required to be post-weld heat treated by VIII-1 rules.
FIG. 1.3 (CONT’D)
The toughness rules for ferritic steels with tensile properties enhanced by heat treatment are given in Paragraph UHT-6 of VIII-1. The rules require such steels to be impact tested regardless of temperature. The measured lateral expansion as defined by ASTM E-23 shall be above 0.015 in.
The toughness rules for high alloy steels are given in Paragraph UHA-51 of VIII-1. The permissible Minimum Design Metal Temperature for base material is summarized in Table 1.7. Similar data are given in VIII-1 for the weld material and weld qualifications. Thermally heated stainless steels may require impact testing per the requirements of UHA-51(c).
The rules for toughness in VIII-2 are different than those in VIII-1. However, the concepts of exemption curves and Charpy impact levels are similar in VIII-2 and VIII-1. The toughness requirements for carbon and low alloy steels are given in Paragraph AM-218 of VIII-2. High alloy steels are covered in Paragraph AM-213. Example 1.2
Problem
Determine the Minimum Design Metal Temperature, MDMT, for the reactor shown in Fig. E1.2. Let the shell, head, pad, and ring material be SA-516 Gr. 70 material. Flange and cover material is SA-105. Pipe
FIG. 1.3 (CONT’D)
material is SA-106. The required shell thickness is 1.75 in., and the required head thickness is 0.86 in. The required nozzle neck thickness is 0.08 in. Assume a joint efficiency of 1.0 and no corrosion allowance. Solution
Shell
SA-516 specifications require the material to be normalized when the thickness exceeds 1.5 in. Thus, from Table 1.6, Curve D is to be used for normalized SA-516 Gr. 70 material. Using Fig. 1.4 and a governing
FIG. 1.4
thickness of 2.0 in., we get a minimum temperature of –5°F. The ratio of required thickness to actual thick-ness is 1.75/2.0 = 0.88. Using Fig. 1.6 for this ratio, we obtain 12°F. Hence, MDMT = –5–12 = –17°F.
Head
For a 1 in. thick head, SA-516 specifications permit a non-normalized material. Thus, from Table 1.6, Curve B is used. Using Fig. 1.4 and a governing thickness of 1.0 in., we get a minimum temperature of 30°F. The
FIG. 1.5
CHARPY IMPACT-TEST REQUIREMENTS FOR FULL SIZE SPECIMENS FOR CARBON AND LOW ALLOY STEELS WITH TENSILE STRENGTH OF LESS THAN 95 KSI (ASME VIII-1)
ratio of re required thickness to actual thickness is 0.86/1.0 = 0.86. Using Fig. 1.6 for this ratio, we obtain 14°F. Hence, MDMT = 30 –14 = 16°F.
Stiffener
For a 0.75-in. stiffener, Curve B of Table 1.6 is to be used. Using Fig. 1.4 and a governing thickness of 0-75 in., we obtain a minimum temperature of 15°F. Since stresses cannot be established from VIII-1 rules, the MDMT = 15°F.
Pad
The material will be normalized since it is 2.00 in. thick. Curve D of Table 1.6 is used. From Fig. 1.4 and a governing thickness of 2.0 in., we obtain a minimum temperature of –5°F. Since stresses cannot be estab-lished from VIII-1 rules, the MDMT = –5°F.
FIG. 1.6
Nozzle Neck
From Table 1.6, Curve B is to be used for a nozzle neck of 0.258-in. thickness. From Fig. 1.4, minimum tem-perature is –20°F. The ratio of required thickness to actual thickness is 0.08/0.258 × 0.875 = 0.36. Using Fig. 1.6 and this ratio, we get 130°F. Hence, MDMT = –20 – –130 = –150°F.
Flange
Since the flange is ANSI B16.5, it is good to –20°F.
Cover
From Fig. 1.3(c), the controlling cover thickness is 2.5/4 = 0.625 in. Curve B applies for this material, and the MDMT = 5°F.
Therefore, the MDMT for this reactor is governed by the head with a value of 16°F. A colder value can be obtained by impact testing the various components. Thus, assuming a specified MDMT of –15°F is required, then the head, stiffener, pad, and cover need impact testing.
TABLE 1.7
MINIMUM DESIGN METAL TEMPERATURES IN HIGH ALLOY STEELS WITHOUT IMPACT TESTING
1.5 FATIGUE REQUIREMENTS
Presently, VIII-1 does not list any rules for fatigue evaluation of components. When fatigue evaluation of a component is required in accordance with UG-22 or U-2(g) of 1, the general practice is to use the VIII-2 fatigue criteria as a guidance up to the temperature limits of VIII-VIII-2. At temperatures higher than those given in VIII-2, the rules of III-H are followed for general guidance. Other fatigue criteria, such as those given in other international codes and ASME B31.3, may also be considered as long as the requirements of U-2(g) of VIII-1 are met.
VIII-2 contains detailed rules regarding fatigue. Paragraph AD-160 gives criteria regarding the need for fatigue analysis. The first criterion is listed in Paragraph AD-160.1 and is based on experience. Vessels that have operated satisfactorily in a certain environment may be cited as the basis for constructing similar ves-sels operating under similar conditions without the need for fatigue analysis.
The second criterion for vessel components is based on the rule that fatigue analysis is not required if all of Condition A or all of Condition B is satisfied, as noted below.
Condition A
Fatigue analysis is not required for materials with a tensile strength of less than 80 ksi when the total num-ber of cycles in (a) through (d) below is less than 1000.
a. The design number of full range pressure cycles including startup and shutdown.
b. The number of pressure cycles in which the pressure fluctuation exceeds 20% of the design pressure,
c. Number of changes in metal temperature between two adjacent points. These changes are mul-tiplied by a factor obtained from the following chart in order to transform them to equivalent cycle number.
d. Number of temperature cycles in components that have two different materials where a differ-ence in the value (α1– α2)∆T exceeds 0.00034. Where, α is the coefficient of thermal
expan-sion and ∆T is the difference in temperature. Condition B
Fatigue analysis is not required when the following items (a) through (f) are met:
a. The number of full range pressure cycles, including startup and shutdown, is less than the num-ber of cycles determined from the appropriate fatigue chart, Fig. 1.7, with an Savalue equal to
3 times the allowable design stress value, Sm.
b. The range of pressure fluctuation cycles during operation does not exceed P(1/3)(Sa/Sm), where
P is the design pressure, Sais the stress obtained from the fatigue curve for the number of
sig-nificant pressure cycles, and Smis the allowable stress. Significant pressure cycles are defined
as those that exceed the quantity P(1/3)(S/Sm). S is defined as
S = Sataken at 106cycles when the pressure cycles are ≤ 106.
S = Sataken at actual number of cycles when the pressure cycles are > 106.
c. The temperature difference between adjacent points during startup and shutdown does not exceed Sa/(2Eα), where Sais the value obtained from the applicable design fatigue curve for
the total specified number of startup and shutdown cycles.
d. The temperature difference between adjacent points during operation does not exceed
Sa/(2Eα), where Sais the value obtained from the applicable design fatigue curve for the total
number of significant fluctuations. Significant fluctuations is defined as those exceeding the quantity S/(2Eα), where S is as defined in (b) above. Adjacent points are defined in AD-160.2, Condition A, Paragraph (c) of VIII-2.
e. Range of significant temperature fluctuation in components that have materials with different coefficient of expansion or modulus of elasticity and that do not exceed the quantity Sa/[2(E1
α1– E2α2)], where α is the coefficient of thermal expansion and E is the modulus of elasticity.
Significant temperature fluctuation is that which exceeds the value S/ [2(E1α1– E2α2)], where
S is as defined in (b) above.
f. Range of mechanical loads does not result in stress intensities whose range exceeds the Savalue
obtained from the fatigue chart.
The third criterion for nozzles with nonintegral reinforcement is given in Paragraph AD-160.3 of VIII-2 and is very similar to Conditions A and B detailed above.
Example 1.3 Problem
A pressure vessel consisting of a shell and two hemispherical heads is constructed from SA 516–70 carbon steel material. The self-reinforced nozzles in the vessel are made from type SA 240–304 stainless steel material. The vessel is shut down six times a year for maintenance. At start-up, the full pressure of 300 psi and full temperature of 400°F are reached in two hours. The maximum ∆T between any two points during start-up is 250°F. At normal operation, the ∆T is negligible. At shutdown, the maximum ∆ T is 100°F.
Metal Temperature Differential,°F Factor
50 or less 0 51 to 100 1 101 to 150 2 151 to 250 4 251 to 350 8 351 to 450 12 Higher than 450 20
FIG.
1.7
F
A
TIGUE CUR
VES FOR CARBON,
LO W ALLO Y , SERIES 4XX, HIGH ALLO Y STEELS, AND HIGH TENSILE STEELS FOR TEMPERA TURES NO T EXCEEDING 700 °F (ASME VIII-2)
Determine the maximum number of years that this vessel can be operated if a fatigue evaluation is not per-formed. Let the coefficient of expansion for carbon steel be 6.5 × 10–6in./in./°F and that for stainless steel
be 9.5 × 10–6in./in./°F.
Solution
From Condition A, determine the number of cycles in one year. a. Number of full pressure cycles for one year is 6. b. This condition does not apply for this case.
c. From the chart, the 250°F difference in temperature during start-up corresponds to 4 cycles. The 100°F difference in temperature during shutdown corresponds to 1 cycle. Thus total equiv-alent cycles due to temperature in one year is (4 + 1) 6 = 30 cycles.
d. At nozzle attachments, the quantity (9.5 × 10–6– 6.5 × 10–6) 400 is equal to 0.0012. Since this
value is greater than 0.00034, the equivalent cycles per year = 6. Total cycles per year due to (a), (c), and (d) = 6 + 30 + 6 = 42.
Number of years to operate vessel if fatigue analysis is not performed = 1000/42 = 23.8 years. Example 1.4
Problem
A pressure vessel has an inside diameter of 60 in., internal pressure of 300 psi, and design temperature of 500°F. The shell thickness is 1/2 in. at an allowable stress level of 18,000 psi (material tensile stress = 70 ksi). The thickness of the hemispherical heads is 1/4 in. at an allowable stress level of 18,000 psi. Integrally reinforced nozzles are welded to the shell and are also constructed of carbon steel with an allowable stress of 18,000 psi. At start-up, the full pressure of 300 psi and full temperature of 500°F are reached in eight hours. The maximum ∆T between any two points during start-up is 60°F. At normal oper-ation, the ∆T is negligible. At shutdown, the maximum ∆T is 50°F. Determine if the shell and heads are adequate for 100,000 cycles without the need for fatigue analysis. From II-D, the coefficient of expansion for carbon steel is 7.25 × 10–6in./in./°F and the modulus of elasticity is 27.3 × 106psi. Use Fig. 1.7 for a
fatigue chart. Solution
Condition B is to be used.
a. Three times allowable stress at the nozzle location is = 3(18,000) = 54,800 psi. Using Fig. 1.7 for this value gives a fatigue life of 4200 cycles.
b. This condition does not apply.
c. From Fig. 1.7, with 100,000 cycles, the value of Sa = 20,000 psi. The value of Sa/(2Eα) =
20,000/(2 × 37.3 × 106× 7.25 × 10–6) = 51°F. Since this value is less than 60°F, the specified
cycles are inadequate. The designer has two options in this situation. The first is to perform fatigue analysis, which is costly. The second option, if it is feasible, is to reduce the ∆T at startup to 51°F.
d. This condition does not apply. e. This condition does not apply. f. This condition does not apply. Example 1.5
Problem
In Example 1.4, determine the required thickness of the shell and heads for 1,000,000 cycles without the need to perform fatigue analysis.
Solution
From Fig. 1.7, with a cycle life of 1,000,000, the value of Sa= 12,000 psi. From Condition B, subparagraph
(a), the maximum stress value for the shell is (12,000/3) = 4,000 psi. The needed shell thickness = 0.5 ×
18,000/4000 = 2.25 in. The required head thickness = 0.25 × 18,000/4000 = 1.13 in. The maximum ∆T at start-up or shutdown cannot exceed Sa/(2Eα) = 12,000/(2 × 27.3 × 106× 7.25 × 10–6) = 31°F, otherwise a
fatigue analysis is necessary.
1.6 PRESSURE TESTING OF VESSELS AND COMPONENTS
1.6.1 ASME Code Requirements
Pressure vessels that are designed and constructed to VIII-1 rules, except those tested in accordance with the requirements of UG-101, are required to pass either a hydrostatic test (UG-99) or a pneumatic test (UG-100) of the completed vessel before the vessel is U-stamped. Pressure vessels that are designed and constructed to VIII-2 rules also are required to pass either a hydrostatic test (Article 3) or a pneumatic test (Article T-4) before the U2-stamp is applied. Each component section of the ASME Boiler and Pressure Vessel Code has a pressure test requirement that calls for a pressure test at or above the maximum allowable working pres-sure indicated on the nameplate or stamping and in the Manufacturer’s Data Report before the appropriate Code stamp mark may be applied.
Under certain conditions, a pneumatic test may be combined with or substituted for a hydrostatic test. When testing conditions require a combination of a pneumatic test with a hydrostatic test, the requirements for the pneumatic test shall be followed. In all cases, the term hydrostatic refers not only to water being an accept-able test medium, but also to oil and other fluids that are not dangerous or flammaccept-able; likewise, pneumatic refers not only to air, but also to other nondangerous gases that may be desirable for “sniffer” detection.
1.6.2 What Does a Hydrostatic or Pneumatic Pressure Test Do?
There is always a difference of opinion as to what is desired and what is accomplished with a pressure test. Some persons believe that the pressure test is meant to detect major leaks, while others feel that there should be no leaks, large or small. Some feel that the test is necessary to invoke loadings and stresses that are equiv-alent to or exceed those loadings and stresses at operating conditions. Others feel that a pressure test is needed to indicate whether a gross error has been made in calculations or fabrication. In some cases, it appears that the pressure testing may help round out corners or other undesirable wrinkles or may offer some sort of a stress relief to some components.
1.6.3 Pressure Test Requirements for VIII-1
1.6.3.1 Hydrostatic Test Requirements. A hydrostatic pressure test is the preferred test method. A pneu-matic test or a combination of pneupneu-matic/hydrostatic test is conducted only when a hydrostatic test cannot be done. Except for certain types of vessels that are discussed later, the hydrostatic test pressure at every point in the vessel shall be at least 1.3 times the maximum allowable working pressure multiplied by the ratio of the allowable tensile stress value at test temperature divided by the maximum allowable tensile stress value at design temperature. As an alternative, a hydrostatic test pressure may be determined by calculations agreed upon by the user and the manufacturer. In this case, the MAWP (maximum allowable working pressure) of each element is determined and multiplied by 1.3 and then adjusted for the hydrostatic head. The lowest value is used for the test pressure, which is adjusted by the test temperature to design temperature ratio. In any case, the test pressure is limited to that pressure which will not cause any visible permanent distortion (yielding) of any ele-ment. The metal temperature of the vessel or component to be tested is recommended to be at least 30°F above the MDMT to be marked on the vessel but need not exceed 120°F, to minimize the risk of brittle fracture. The test pressure shall not be applied until the vessel and its contents are at about the same temperature. If the test temperature is above 120°F. it is recommended that inspection be delayed until the temperature of the vessel is
120°F or less. Also, it is recommended that a liquid relief valve set to 11/
3time the test pressure be installed in
the pressure test system in those cases where the vessel under test may warm up while test personnel are absent. After the test pressure is reached and the other items noted above are satisfied, the pressure is reduced to the test pressure divided by 1.3. At that time welded joints, connections, and other areas are visually examined for leaks and cracks. The visual examination may be waived if a gas leak test is to be applied, if hidden welds have been examined ahead of time, and if the vessel will not contain a lethal substance. Venting shall be provided at all high locations where there is a possibility of air pockets forming during the filling of the vessel for testing. The general rules for hydrostatic testing do not call for a specific time for holding the vessel at test pressure. The length of this time may be set by the Authorized Inspector or by a contract specification.
1.6.3.2 Pneumatic Test Requirements. For some vessels, it is necessary to apply a pneumatic test in lieu of a hydrostatic test. This may be due to any number of reasons, including vessels designed and sup-ported in such a manner that they cannot be safely filled with liquid and vessels that cannot tolerate any trace of water or other liquids. If a vessel is to be pneumatically tested, it shall first be examined according to the requirements of UW-50. This paragraph requires that welds around openings and attachments be examined by MT or PT before testing. Except for certain vessels, the pneumatic test pressure at every point in the ves-sel shall be 1.1 times the maximum allowable working pressure multiplied by the ratio of the allowable ten-sile stress value at test temperature divided by the allowable tenten-sile stress value at design temperature. For pneumatic testing, the metal temperature of the vessel or component shall be at least 30°F above the MDMT to be marked on the vessel.
The test pressure shall be gradually increased to no more than half of the full test pressure and then increased in steps of one-tenth of the test pressure until the full test pressure is reached. After that, the pres-sure shall be reduced to the test prespres-sure divided by 1.1 and all areas are to be examined. All other require-ments for hydrostatic testing shall be observed, including the waiving of the visual examination, provided the same limits are met.
1.6.3.3 Test Requirements for Enameled or Glass-lined Vessels. The maximum test pressure for enameled and glass-lined vessels does not have to be any greater than 1.0 MAWP unless required by the Authorized Inspector or by a contract specification. Higher test pressure may damage the enameled or glass coating. All other rules for hydrostatic testing apply.
1.6.3.4 Test Requirements for Vessels Built to the Rules of Parts UCI or UCD. For those vessels designed and constructed to the rules of Part UCI for Cast Iron and Part UCD for Cast Ductile Iron, where the factor of safety on tensile strength to set the allowable tensile stress values is 10 and 5, respectively, the multiplier for the hydrostatic test pressure is set differently. For Part UCI, the test pressure shall be 2.5 MAWP, but is not to exceed 60 psi for a design pressure less than 30 psi and 2.0 MAWP for a design pres-sure equal to or greater than 30 psi. For Part UCD, the test prespres-sure shall be 2.0 MAWP. With these changes, the remaining rules of UG-99 are followed.
1.6.3.5 Test Requirements for Vessels Built to the Rules of Part ULT. Alternative rules for the design and construction of vessels to operate at cold temperatures as low as –320°F are given in Part ULT. These rules permit the use of increased allowable tensile stress values at temperatures colder than ambient temper-ature to as low as –320°F for 5%, 8%, and 9% nickel steels, 5083 aluminum alloy, and Type 304 stainless steels. Other materials listed in both Section II and Subsection C may be used for vessels and parts for design at cold temperature with the allowable tensile stress values set by the value at 100°F. When the vessel is designed and constructed to Part ULT rules, special hydrostatic testing requirements are necessary due to the fact that the material is stronger at design temperature than at ambient test temperature. The vessel shall be hydrostatically tested at ambient temperature with the test pressure held for 15 minutes and either of the fol-lowing criteria may be applied:
a. A standard hydrostatic test as described in 1.6.3.1 is used, but with the ratio of allowable stresses not applied and the test pressure shall be 1.4 MAWP, if possible, instead of 1.3 MAWP. b. In applying (a), the membrane stress in the vessel shall not exceed 0.95 of the specified mini-mum yield strength nor 0.5 of the specified minimini-mum tensile strength. In complying with these stress limits, the ratio of hydrostatic test pressure divided by the MAWP may be reduced below 1.4, but it shall be not less than 1.1 MAWP. If the value comes out less than 1.1 MAWP, a pneu-matic test shall be conducted using the rules of UG-100, but omitting the adjustment for the allowable tensile stress ratio.
A vessel to be installed vertically may be tested in the horizontal position, provided the test pressure is applied for 15 minutes at not less than 1.4 MAWP, including a pressure equivalent to the liquid head in oper-ating position.
1.6.3.6 Proof Testing to Establish MAWP. In addition to the hydrostatic or pneumatic pressure test of the completed vessel, a pressure proof test is permitted to establish the MAWP of vessels and vessel parts for which the strength cannot be calculated with assured accuracy. The rules for such a pressure proof test are given in UG-101 of VIII-1 and may be based on yielding or on bursting of the vessel or vessel part. Proof tests must be witnessed by the Authorized Inspector, who indicates acceptance by signing the Manufacturer’s Data Report Form. Duplicate or similar parts to that part which has had its MAWP established by a proof test according to the requirements of UG-101(d) of VIII-1 may be used without a proof test of their own, but shall be given a hydrostatic or pneumatic pressure test as part of the completed vessel pressure test.
1.6.4 Pressure Test Requirements for VIII-2
1.6.4.1 Hydrostatic Test Requirements. Except for glass-lined and enameled vessels, the hydrostatic test pressure at every point in the vessel shall be 1.25 times the design pressure (or MAWP) to be marked on the vessel multiplied by the ratio of the design stress intensity value at test temperature divided by the design stress intensity value at design temperature. For glass-lined or enameled vessels, the hydrostatic test pressure shall be at least equal to, but need not exceed, the design pressure (or MAWP). Similar to the alternative in VIII-1, the hydrostatic test pressure may be determined by calculations agreed upon between the User and the Manufacturer and shall be described in the Design Report.
The hydrostatic test pressure shall not exceed a value that results in the following:
a. A calculated primary membrane stress intensity Pmof 90% of the tabulated yield strength Syat
test temperature;
b. A calculated primary membrane plus primary bending stress intensity Pm+ Pbnot to exceed
the limits given below:
[Pm+ Pb] ≤ 1.35 Sy, when Pm≤ 0.67 Sy (1.1)
[Pm+ Pb] ≤ 2.35 Sy– 1.5 Pm, when 0.67 Sy< Pm≤ 0.90 Sy (1.2)
1.6.4.2 Pneumatic Test Requirements. A pneumatic test is permitted only when one of the following prevails:
a. Vessels cannot be safely filled with water due to their design and support system; b. Vessels in which traces of testing liquid cannot be tolerated.
When a pneumatic test is permitted in lieu of a hydrostatic test, except for glass-lined and enameled ves-sels, the pneumatic test pressure at every point in the vessel shall be 1.15 times the design pressure (or
MAWP) to be marked on the vessel multiplied by the ratio of the design stress intensity value at test tem-perature divided by the design stress intensity value at design temtem-perature. For glass-lined or enameled ves-sels, the pneumatic test pressure shall be at least equal to, but need not exceed, the design pressure (or MAWP).
The pneumatic test pressure shall not exceed a value that results in the following:
a. A calculated primary membrane stress intensity Pmof 80% of the tabulated yield strength Syat
test temperature;
b. A calculated primary membrane plus primary bending stress intensity Pm+ Pbnot to exceed
the following limits:
[Pm+ Pb] ≤ 1.20 Sy, when Pm≤ 0.67 Sy (1.3)
