PRESSURE VESSEL Handbook - Eugene F. Megyesy 12th 2001

PRESSURE VESSEL HANDBOOK

PRESSURE VESSEL HANDBOOK Twelfth Edition

with foreword by

Paul Buthod Professor of Chemical Engineering University of Tulsa Tulsa, Oklahoma

Eugene F. Megyesy

PRESSURE VESSEL PUBLISHING, INC. P.O. Box 35365· Tulsa, Oklahoma 74153

Copyright © by Eugene F. Megyesy Copyright 1972, 1973 by Pressure Vessel Handbook Publishing, Inc. All rights reserved. No part of this book may be reproduced in any form or by any means including information storage and retrieval systems - without permission of the publisher. Library of Congress Control Number: 2001130059 ISBN 0-914458-21-3 COPYRIGHT© 1972,1973,1974,1975,1977,1979,1981,1982,1983, 1986, 1989, 1992,1995,1998,2001 Printed and bound in the United States of America.

NOTE: This new edition of the Pressure Vessel Handbook supersedes all previous editions, effective July 1,2001. The changes over the previous Eleventh Edition have been made necessary by the revision of Codes, Standards, Specifications, etc.

FOREWORD

Engineers who design equipment for the chemical process industry are sooner or later confronted with the design of pressure vessels and mounting requirements for them.

This is very often a frustrating

experience for anyone who has not kept up with current literature in the field of code requirements and design equations.

First he must familiarize himself with the latest version of the applicable code.

Then he must search the literature for techniques

used in design to meet these codes. Finally he must select material properties and dimensional data from various handbooks and company catalogs for use in the design equations. Mr. Megyesy has recognized this problem.

For several years he

has been accumulating data on code requirements and calculational methods.

He has been presenting this information first in the form

of his "Calculation Form Sheets" and now has put it all together in one place in the Pressure Vessel Handbook. I believe that this fills a real need in the pressure vessel industry and that readers will find it extremely useful.

Paul Buthod

PREFACE This reference book is prepared for the purpose of making fonnulas, technical data, design and construction methods readily available for the designer, detailer, layoutmen and others dealing with pressure vessels. Practical men in this industry often have difficulty finding the required data and solutions, these being scattered throughout extensive literature or advanced studies. The author's aim was to bring together all of the above material under one cover and present it in a convenient fonn. The design procedures and fonnulas of the AS ME Code for Pressure Vessels, Section VIII Division I have been utilized as well as those generally accepted sources which are not covered by this Code. From among the alternative construction methods described by the Code the author has selected those which are most frequently used in practice. In order to provide the greatest serviceability with this Handbook, rarely occurring loadings, special construction methods or materials have been excluded from its scope. Due to the same reason this Handbook deals only with vessels constructed from ferrous material by welding, since the vast majority of the pressure vessels are in this category. A large part of this book was taken from the works of others, with some of the material placed in different arrangement, and some unchanged. The author wishes to acknowledge his indebtedness to Professor Sandor Kalinszky, Janos Bodor, Laszl6 Felegyhazy and J6zsef Gyorfi for their material and valuable suggestions, to the American Society of Mechanical Engineers and to the publishers, who generously pennitted the author to include material from their publications. The author wishes also to thank all those who helped to improve this new edition by their suggestions and corrections. Suggestions and criticism concerning some errors which may remain in spite of all precautions shall be greatly appreciated. They contribute to the further improvement of this Handbook. Eugene F. Megyesy

7

ASME CODE vs. THIS HANDBOOK The ASME BOILER AND PRESSURE VESSEL CODE - 2001, Sect. VIII, Div. 1 The American Society of Mechanical Engineers set up a Committee in 1911 for the purpose of formulating standard rules for the construction of steam boilers and other pressure vessels that will perfonn in a safe and reliable manner. The Code comprises these rules. It's scope includes vessels: 1. made of nonferrous materials, cast iron, high alloy and carbon steel, 2. made by welding, forging, bracing, and 3. applying a wide variety of construction methods and details. It includes all vessels where the question of safety is concerned.

PRESSURE VESSEL HANDBOOK2001, Twelfth Edition The Handbook covers design and construction methods of pressure vessels: 1. made of carbon steel, 2. made by welding 3. applying construction methods and details which are the most economical and practical, which are in accordance with the Code rules, and thus generally followed by the industry. The vast majority of the pressure vessels today fall into this category. For construction rules and details which are excluded from the scope of the Handbook, references are made to the applicable Code paragraphs to avoid neglecting them.

The Code - as it is stated in paragraph UG2 - "does not contain rules to cover all details of design and construction ... " "where details are not given, it is intended that the Manufacturer ... shall provide details of design and construction."

Details of design and construction not covered by the Code are offered by the Handbook including: Design of tall towers, wind load, earthquake, vibration, eccentric load, elastic stability, deflection, combination of stresses, nozzle loads, reaction of supports, lugs, saddles, and rectangular tanks.

"The Code is not a handbook." "It is not intended that this Section be used as a design handbook" as it is stated in the Foreword of the Code.

The updated and revised Code is published in three years intervals. Addenda, which also include revisions to the Code, are published annually. Revisions and additions become mandatory six (6) months after the date of issuance, except for boilers and pressure vessels contracted for prior to the end ofthe 6 month period. (Code Foreword)

The aim of this Handbook is to be easily handled and consulted. Tables, charts eliminate the necessity of calculations, Geometry, layout of vessels, piping codes, API storage tanks, standard appurtenances, painting of steel surfaces, weights, measurements, conversion tables, literature, definitions, specification for vessels, design of steel structures, center of gravity, design of welded joints, boIted connections, boiler and pressure vessel laws, chemical resistance of metals, volumes, and surfaces of vessels, provide good serviceability. The Handbook is updated and revised in three years intervals, reflecting the changes of Code rules, new developments in the design and construction method, and includes the revisions of its sources.

8

THE ASME CODE ASME Boiler and Pressure Vessel Code, Section VIII, Division 1 An internationally recognized Code published by The American Society of Mechanical Engineers. PRESSURE VESSEL - is a containment of solid, liquid or gaseous material under internal or external pressure, capable of withstanding also various other loadings. BOILER - is a part of a steam generator in which water is converted into steam under pressure. RULES OF DESIGN AND CONSTRUCTION - Boiler explosions around the turn of the century made apparent the need for rules governing the design and construction of vessels. The first ASME Code was published in 1914. ISSUE TIME - The updated and revised Code is published in three years intervals (2001 and so on). Addenda, which also include revisions to the Code, are published annually. Revisions and additions become mandatory 6 months after the date of issuance, except for boilers and pressure vessels contracted for prior to the end of the 6 month period. (Code Foreword) SCOPE OF THE CODE - The rules of this Division have been formulated on the basis of design principles and construction practices applicable to vessels designed for pressures not exceeding 3000 psi. Code U-l (d) Vessels, which are not included in the scope of this Division, may be stamped with the Code U Symbol if they meet all the applicable requirements of this Division. Code U-2(g) THE DESIGN METHOD - The Code rules concerning design of pressure parts are based on the maximum stress theory, i.e., elastic failure in a ductile metal vessel occurs when the maximum tensile stress becomes equal to the yield strength ofthe material. OTHER COUNTRIES' Codes deviate from each other considerably, mainly because of differences in the basic allowable design stresses. The ASME Code's regulations may be considered to be at midway between conservative and unconservative design. COMPUTER PROGRAMS - Designers and engineers using computer programs for design or analysis are cautioned that they are responsible for all technical assumptions inherent in the programs they use and they are solely responsible for the application of these programs to their design. (Code, Foreword) DESIGN AND CONSTRUCTION NOT COVERED - This Division ofthe Code does not contain rules to cover all details of design and construction. Where complete details are not given, it is intended that the Manufacturer shall provide details which will be as safe as those provided by the rules of this Division. Code U-2(g)

CONTENTS

PART I

Design and Construction of Pressure Vessels ............. 11

PART II

Geometry and Layout of Pressure Vessels .............. 257

PART III

Measures and Weights ............................................ 321

PART IV

Design of Steel Structures ....................................... 447

PART V

Miscellaneous .......................................................... 465

11

PART I. DESIGN AND CONSTRUCTIONS OF PRESSURE VESSEL 1. Vessels Under Internal Pressure ............................................ Stresses in Cylindrical Shell, Definitions, Formulas, Pressure of Fluid, Pressure-Temperature Ratings of American Standard Carbon Steel Pipe Flanges.

15

2. Vessels Under External Pressure ............................................ Definitions, Formulas, Minimum Required Thickness of Cylindrical Shell, Chart for Determining Thickness of Cylindrical and Spherical Vessels under External Pressure when Constructed of Carbon Steel.

31

3. Design of Tall Towers ............................................................ Wind Load, Weight of Vessel, Seismic Load, Vibration, Eccentric Load, Elastic Stability, Deflection, Combination of Stresses, Design of Skirt Support, Design of Anchor Bolts (approximate method), Design of Base Ring (approximate method), Design ofAnchor Boltand Base Ring, Anchor Bolt Chair for Tall Towers.

52

4. Vessel Supports........... ............... ....... .............. ...................... Stresses in Large Horizontal Vessels Supported by Two Saddles, Stresses in Vessels on Leg Support, Stresses in Vessels Due to Lug Support, Lifting Attachments, Safe Loads for Ropes and Chains.

86

5. Openings ......................................... :...................................... Inspection Openings, Openings without Reinforcing Pad, Opening with Reinforcing Pad, Extension of Openings, Reinforcement of Openings, Strength of Attachments, Joining Openings to Vessels, Length of Couplings and Pipes for Openings.

122

6. Nozzle Loads ..........................................................................

153

7. Reinforcement at the Junction of Cone to Cylinder ...............

159

8. Welding of Pressure Vessels ................................................. Welded Joints, Butt Welded Joint of Plates of Unequal Thicknesses, Application of Welding Symbols.

170

9. Regulations, Specifications .................................................... Code Rules Related to Various Services, Code Rules Related to Various Plate Thicknesses of Vessel, Tanks and Vessels Containing Flammable and Combustible Liquids, Properties of Materials, Description of Materials, Specification for the Design and Fabrication of Pressure Vesels, Fabrication Tolerances.

181

12

10. Materials of Foreign Countries ..............................................

194

11. Welded Tanks ........................................................................

203

12. Piping Codes ..................... ................ ......................... ............

208

13. Rectangular Tanks .................................................. ................

213

14. Corrosion ................................................................................

221

15. Miscellaneous ....................................................................... Fabricating Capacities, Pipe and Tube Bending, Pipe Engagement, Drill Sizes for Pipe Taps, Bend Allowances, Length of Stud Bolts, Pressure Vessell Detailing, Preferred Locations, Common Errors, Transportation of Vessels.

232

16. Painting of Steel Surfaces .....................................................

247

IN REFERENCES THROUGHOUT THIS BOOK "CODE" STANDS FOR ASME BOILER AND PRESSURE VESSEL CODE SECTION VIII, DIVISION 1 - AN AMERICAN STANDARD.

2001 EDITION

13

STRESSES IN PRESSURE VESSELS Pressure vessels are subject to various loadings, which exert stresses of different intensities in the vessel components. The category and intensity of stresses are the function of the nature ofloadings, the geometry and construction of the vessel components. LOADINGS (Code UG-22) a. Internal or external pressure b. Weight of the vessel and contents c. Static reactions from attached equipment, piping, lining, insulation, d. The attachment of internals, vessel supports, lugs, saddles, skirts, legs e. Cyclic and dynamic reactions due to pressure or thermal variations f. Wind pressure and seismic forces g. Impact reactions due to fluid shock h. Temperature gradients and differential thermal expansion 1. Abnormal pressures caused by deflagration. STRESSES (Code UG-23)

a.

Tensile stress

b. Lingitudinal compressive stress c. General primary membrane stress induced by any combination of loadings. Primary membrane stress plus primary bending stress induced by combination of loadings, except as provided in d. below. d. General primary membrane stress induced by combination of earthquake or wind pressure with other loadings. Seismic force and wind pressure need not be considered to act simulta neously.

MAXIMUM ALLOWABLE STRESS Sa = Maximum allowable stress in tension for carbon and low alloy steel Code Table UCS-23; for high alloy steel Code Table UHA-23., psi. (See properties of materials page 186-190.) The smaller of Sa or the value of factor B determined by the procedure described in Code UG 23 (b) (2)

1.5 Sa

Sa = (see above)

1.2 times the stress permitted in a., b., or c. This rule applicable to stresses exerted by internal or external pressure or axial compressive load on a cylinder.

14

STRESSES IN CYLINDRICAL SHELL

Unifonn internal or external pressure induces in the longitudinal seam two times larger unit stress than in the circumferential seam because of the geometry of the cylinder. A vessel under external pressure, when other forces (wind, earthquake, etc.) are not factors, must be designed to resist the circumferential buckling only. The Code provides the method of design to meet this requirement. When other loadings are present, these combined loadings may govern and heavier plate will be required than the plate which was satisfactory to resist the circumferential buckling only. The compressive stress due to external pressure and tensile stress due to internal pressure shall be detennined by the fonnulas: FORMULAS LONGITUDINAL JOINT

CIRCUMFERENTIAL JOINT

S _ PD 2t

2 -

D p 51 52 t

= =

= = =

NOTATION Mean diameter of vessel, inches Internal or external pressure, psi Longitudinal stress, psi Circumferential (hoop) stress, psi Thickness of shell, corrosion allowance excluded, inches

EXAMPLE

Given

D p t

=

= =

96 inches 15 psi 0.25 inches

PD SI = 4t

=

15 x 96 4 x 0.25

15 x 96

=

1440 psi

=

2880 psi

2 x 0.25 For towers under internal pressure and wind load the critical height above which compressive stress governs can be approximated by the formula:

H

=

PD 12t

where H = Critical height of tower, ft.

15

INTERNAL PRESSURE 1.

OPERA TING PRESSURE

The pressure which is required for the process, served by the vessel, at wh ich the vessel is normally operated. 2.

DESIGN PRESSURE

The pressure used in the design of a vessel. It is recommended to design a vessel and its parts for a higher pressure than the operating pressure. A design pressure higher than the operating pressure with 30 psi or 10 percent, whichever is the greater, will satisfy this requirement. The pressure of the fluid and other contents of the vessel should also be taken into consideration. See tables on page 29 for pressure of fluid.

3.

MAXIMUM ALLOWABLE WORKING PRESSURE

The internal pressure at which the weakest element of the vessel is loaded to the ultimate permissible point, when the vessel is assumed to be: (a) (b) (c) (d)

in corroded condition under the effect of a designated temperature in normal operating position at the top underthe effect of other loadings (wind load, external pressure, hydrostatic pressure, etc.) which are additive to the internal pressure.

When calculations are not made, the design pressure may be used as the maximum allowable working pressure (MA WP) code 3-2. A common practice followed by many users and manufacturers of pressure vessels is to limit the maximum allowable working pressure by the head or shell, not by small elements as flanges, openings, etc. See tables on page 28 for maximum allowable pressure for flanges. See tables on page 142 for maximum allowable pressure for pipes. The term, maximum allowable pressure, new and cold, is used very often. It means the pressure at which the weakest element of the vessel is loaded to the ultimate permissible point, when the vessel: (a) is not corroded (new) (b) the temperature does not affect its strength (room temperature ) (cold) and the other conditions (c and d above) also need not to be taken into consideration. 4.

HYDROSTATIC TEST PRESSURE

At least 1.3 times the maximum allowable working pressure or the design pressure to be marked on the vessel when calculations are not made to determine the maximum allowable working pressure. Ifthe stress value of the vessel material at the design temperature is less than at the test temperature, the hydrostatic test pressure should be increased proportionally. Hydrostatic test shall be conducted after all fabrication has been completed.

16

In this case, the test pressure shall be: 1.5 X

Stress Value S At Test Temperature Max. Allow. W. Press. (Or Design Press.)

X

Stress Value S At Design Temperature

Vessels where the maximum allowable working pressure limited by the flanges, shall be tested at a pressure shown in the table: Primary Service Pressure Rating

1501b

300lb 400lb

600lb

9001b 1500lb

Hydrostatic Shell Test Pressure

425

1100

2175

3250

1450

5400

25001b 9000

Hydrostatic test of multi-chamber vessels: Code UG-99 (e) A Pneumatic test may be used in lieu of a hydrostatic test per Code UG-I 00 Proof tests to establish maximum allowable working pressure when the strength of any part of the vessel cannot be computed with satisfactory assurance of safety, prescribed in Code UG-I 0 1. 5. MAXIMUM ALLOWABLE STRESS V ALVES

The maximum allowable tensile stress values permitted for different materials are given in table on page 189. The maximum allowable compressive stress to be used in the design of cylindrical shells subjected to loading that produce longitudinal compressive stress in the shell shall be determined according to Code par. UG-23 b, c, & d. 6. JOINT EFFICIENCY The efficiency of different types of welded joints are given in table on page 172. The efficiency of seamless heads is tabulated on page 176. The following pages contain formulas used to compute the required wall thickness and the maximum allowable working pressure for the most frequently used types of shell and head. The formulas of cylindrical shell are given for the longitudinal seam, since usually this governs. The stress in the girth seam will govern only when the circumferential joint efficiency is less than one-half the longitudinal joint efficiency, or when besides the internal pressure additional loadings (wind load, reaction of saddles) are causing longitudinal bending or tension. The reason for it is that the stress arising in the girth seam pound per square inch is one-half of the stress in the longitudinal seam. The formulas for the girth seam accordingly:

PR 1 =

2SE

See notation on page 22.

+ O.4P

P =

2SEI R - 0.41

17 NOTES

18

INTERNAL PRESSURE FORMULAS IN TERMS OF INSJDE DIMENSIONS NOTATION P

5

E

=

Design pressure or max. allowable working pressure psi = Stress value of material psi. page 189

A

R

= Joint efficiency. page 172 = Inside radius. inches

D = Inside

diameter. inches Wall thickness. inches CA. = Corrosion allowance. inches t

=

CYLINDRICAL SHELL (LONG SEAM) 1

r

/

\

\

t

+--.f-- l~ '~

PR SE-O.6P

t

R

P=

SE~_ R+O.6t

I.

Usually the stress in the long seam is governing. See preceding page.

2.

When the wall thickness exceeds one half of the inside radius or P exceeds 0.385 SE, the formulas given in the Code Appendix 1-2 shall be applied.

B

SPHERE

6

and

HEMISPHERICAL HEAD

PR 2SE-O.2P

t

P

2SE t R+O.2t

~"","~

l

R

f

I.

For heads without a straight flange. use the efficiency of the head to shell joint if it less than the efficiency of the seams in the head.

2. When the wall thickness exceeds 0.356 R or P exceeds 0.665 SE. the formulas given in the Code Appendix 1-3, shall be applied. C

2: 1 ELLIPSOIDAL HEAD

hS-~ I. ]\1 0

h = 0/4

t

I.

PD 2SE-O.2P

P=

2SEt D+O.2t

For ellipsoidal heads, where the ratio of the major and minor axis is other than 2: I, see Code Appendix 1-4(c).

19

EXAMPLES E

DESIGN DATA: p = 100 psi design pressure S = 20,000 psi stress value of SA 515-70 plate @ 500°F E = 0.85, efficiency of spot-examined joints of shell and hem is. head to shell

1.00, joint efficiency of seamless heads R = 48 inches inside radius* D = 96 inches inside diameter* t = required wall thickness, inches CA. = 0.125 inches corrosion allowance * in corroded condition greater with the corrosion allowance.

SEE DESIGN DATA ABOVE

SEE DESIGN DATA ABOVE

Determine the required thickness, t of a shell

Determine the maximum allowable working pressure P for 0.500 in. thick shell when the vessel is in new condition.

100 X 48.125 . t=20,000 X 0.85 -0.6XIOO =0.284 m.

+CA.

=

P=20,000 X 0.85 X 0.500 48 + 0.6 X 0.500

0.125 in. 0.409 in.

176 psi

Use 0.500 in. plate

SEE DESIGN DATA ABOVE

SEE DESIGN DATA ABOVE

The head furnished without straight flange. Determine the required thickness, t ofa hemispherical head. t

100X48.125 =0.142 in. 2 X 20,000 X 0.85 -0.2 XI 00 +C A.

0.125 in. 0.267 in.

Determine the maximum allowable working pressure, P for 0.3125 in. thick head, when it is in new condition. P 2 X 20,000 X 0.85 X 0.3125 -221 48+0.2XO.3125 -

. pSI

Use OJ 125 in. plate

SEE DESIGN DATA ABOVE Determine the required thickness ofa seamless ellipsoidal head. 100 X 96.25 . t 2 X 20,000 X 1.0-0.2 X 100 =0.241 m.

+CA.

0.125 in. OJ66 in.

Use 0.375 in. min. thk. head

SEE DESIGN DATA ABOVE

Determine the maximum allowable working pressure, P for 0.250 in. thick seamless head, when it is in corroded condition. P

2 X 20,000 X 1.0 X 0.250 -103 . 96.25 + 0.2 X 0.250 pSI

20

INTERNAL PRESSURE FORMULAS IN TERMS OF INSIDE DIMENSIONS

D = Inside diameter, inches

NOTATION

a = One half of the included (apex) angle, degrees L = Inside radius of dish, inches r = Inside knuckle radius. inches I = Wall thickness, inches C.A. = Corrosion allowance. inches

P = Design pressure or max. allowable working pressure psi S = Stress value of material psi, page 189

E = Joint efficiency, page 172 R = Inside radius. inches

D

CONE

AND

CONICAL SECTION

PD

t=~~----~~~~:7

2 cos a (SE-O.6P)

P= 2SEt cos a D+ 1.2t cos a

I. The half apex angle, a not greater than 30° 2. When a is greater than 30 ~ special analysis is required. (Code Appendix 1-5(g))

ASME FLANGED AND DISHED HEAD

E

(TORISPHERICAL HEAD)

0.885PL t

SEt

SE-O.IP

P

=-=0--=.8=85-=-=L=-+-0-=-."'-1t

When lIr less than 16 2/3

\ When the min. t~nsile strength of material exceeds 70,000 psi. see Code UG-32(e)

PLM

p=

t = -=-2=SE=--=0-=.2-=P

2SEt

LM+0.2t

VALUES OF FACTOR "M" L/r M

L/r M

*

1.00

1.50 1.25

1.00

1.06 1.08

1.03 7.00

8.00 1.46

2.00

2.25

1.10

2.50 2.75 1.15

1.13

1.17 10.0

9.00 8.50

7.50 1.41

I. 75

19·50

3.00 1.18

3.25 1.20

11.0 10.5

3.50

4.00

1.22 1.25 12.0

11.5

4.50 1.28

5.00 1.31

14.0 13.0

5.50 6.00 1.34 1.36 16.0

15.0

2 16]"

1.54 1.50 1.58 1.62 1.69 I. 75 1.48 1.52 1.56 I.~Q 1.65 1.72 1.77 THE MAXIMUM ALLOWED RATIO: L - D + 2t (see note 2 on facing page) 1.44

6.50 1.39

•

21

EXAMPLES DESIGN DATA: P = 100 psi design pressure S = 20,000 psi stress value of SA 515-70 plate @ 500°F E = 0.85, efficiency of spot-examined I joints ' E = 1.00, joint efficiency of seamless heads SEE DESIGN DATA ABOVE cos 30°= 0.866 Determine the required thickness, t of a cone

L = 96 inches inside radius of dish* D = 96 inches inside diameter* t = required wall thickness, inches a = 300 0ne half ofthe apex angle CA. = 0.125 inches corrosion allowance * in corroded condition greater with the corrosion allowance

SEE DESIGN DATA ABOVE

I

!

100 X 96.25 -0 "28 . t 2XO.866 (20,000 X 0.85 _0.6XlOO) - . J m.

+C.A.

Q 125 jn

Determine the maximum allowable working pressure, P for 0.500 in. thick cone, when the vessel is in new condition. P 2X20,000XO.85XO.500XO.866 96 + 1.2 X 0.500 X 0.866 152psi

0.453 in.

Use 0.500 in. plate SEE DESIGN DATA ABOVE

SEE DESIGN DATA ABOVE Llr = 16~ Determine the required thickness, t of a seamless ASME flanged and dished head. 0.885 x 100 x 96.125 . t=20,OOO x 1.0-0.1 XI 00 0.426 m.

+CA.

0.125 in. 0.551 in.

Determine the maximum allowable working pressure, P for 0.5625 in. thick seamless head, when the vessel is in new condition. P=

20,000 X 1.0 X 0.5625 132 psi 0.885 X 96 + 0.1 X 0.5625

Use 0.5625 in. plate SEE DESIGN DA T A ABOVE Knuckle radius r = 6 in. Llr = ~ = 16 M= 1.75 from table. Determine the required thickness t of a seamless ASME flanged and dished head. 100 X96.125 X 1.75 . t= 2 X 20,000 - 0.2 X 100 -0.421 m. 0.125 in. 0.54610. Use 0.5625 in. min. thick head

+CA.

SEE DESIGN DATA ABOVE Knuckle radius r = 6 in. Llr = 9£ = 16 M= 1.75 from table Determine the maximum allowable working pressure, P for a 0.5625 in. thick seamless head when the vessel is in corroded condition.

p= 2x20,000xl.OxO.5625 104 . 96.125x1.75+0.2 x0.4375 pSI

NOTE: When the ratio of Llr is greater than 161, fupn-Code construction) the values of M may be calculated by the formula: M = Y4 (3 + Wr)

22

INTERNAL PRESSURE FORMULAS IN TERMS OF OUTSIDE DIMENSIONS

NOTATION

= Joint efficiency, page 172 = Outside radius, inches D = Outside· diameter, inches t = Wall thickness, inches C.A. = Corrosion allowance. inches E

P = Design pressure or max. allowable

R

working pressure psi

S

= Stress va'ue of material psi, page 189

A

CYLINDRICAL SHELL (LONG SEAM)I

PR 1-

SE

+ 0.4P

SEI

P - R - O.4t

1. Usually the stress in the long seam is governing. See page 14 2. When the wall thickness exceeds one half of the inside radius or P exceeds 0.385 SE, the formulas givencin the Code Appendix 1-2 shall be applied.

B

SPHERE and HEMISPHERICAL HEAD

PR 1 -

2SE

+ 0.8P

2SEI P - R - 0.81

1. For heads without a straight flange, use the efficiency

of the head to shell joint if it is less than the efficiency of the seams in the head. 2. When the wall thickness exceeds 0.356 R or P exceeds 0.665 SE, the formulas given in the Code Appendix 1-3, shall be applied.

C

2:1 ELLIPSOIDAL HEAD PD

t=""'2"""'S=E-+--:"1-".8~P

2SEt P= ..- - D· -1.8t

1. For ellipsoidal heads, where the ratio of the major and minor axis is other than 2:1, see Code Appendix 1-4(c).

h = 0/4

23

EXAMPLES DESIGN DATA: P = 100 psi design pressure S = 20,000 psi stress value of SA 515-70 plate @ 500°F E = 0.85, efficiency of spot-examined joints of shell and hemis. head to shell

R D t

CA.

1.00, joint efficiency ofseamless heads = 48 inches outside radius = 96 inches outside diameter = Required wall thickness, inches = 0.125 inches corrosion allowance =

SEE DESIGN DATA ABOVE

SEE DESIGN DATA ABOVE Determine the required thickness, t of a shell 100 X 48 0.283 in t 20,000 X 0.85 - 0.4 X 100 +C.A.

E

0.125 in. 0.408 in.

Determine the maximum allowable working pressure, P for 0.4375 in. thick shell when the vessel is in new condition. 20,000 X 0.85 X 0.4375 p= 155psi 48 - 0.4 X 0.4375

Use: 0.4375 in. thick plate

SEE DESIGN DATA ABOVE

SEE DESIGN DATA ABOVE Head furnished without straight flange. Determine the required thickness, t of a hemispherical head. 100X48 0.141 in. 2 X20,000XO.85 +0.8 X 100

t

+CA.

0.125 in. 0.266 in.

Determine the maximum allowable wOFking pressure, P for 0.3125 in. thick head, when the vessel is in new condition.

p=2 X 20,000 X 0.85 X 0.3125 222 psi 48-0.8 XO.3125

Use: 0.3125 in. min. thick head

SEE DESIGN DA T A ABOVE

SEE DESIGN DATA ABOVE

Determine the required thickness t of a seamless ellipsoidal head.

Determine the maximum allowable working pressure, P for 0.375 in. thick head, when it is in new condition.

t

100X96 0.239 in. 2 X 20,000 X l.0+ l.8X 100 +CA.

Use 0.375 in. min. thick head

0.125 in. 0.364 in.

P

2 X 20,000 X 1.0 X 0.375 96-1.8 X 0.375 157psi

24

INTERNAL PRESSURE FORMULAS IN TERMS OF OUTSIDE DIMENSIONS NaTATION

D = Outside diameter, inches a = One half of the included (apex) angle. degrees L = Outside radius of dish, inches r = Inside knuckle radius. inches t = Wall thickness, inches C.A. = Corrosion allowance. inches

P = Design pressure or max. allowable working pressure psi S = Stress value of material psi, page 189

E = Joinl efficiency. page 172 R = OUlside radius, inches

D

CONE

AND

CONICAL SECTION

~

!

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