Pvelite Seminar Notes

vRJlJlifJ~L I ::luLl OIrz..oo3

C·.I~' ENGINEERING

SOFTWARE

Pressure Vessel Design and AnalVsis Seminar Notes

Revised 6/2001

Table of Contents

Table ofContents Chapter 1: II/trodl/ctioll to the Semil/a/' & the Sofhl'are Purpose of the Seminar 1 Structure of the Seminar 1 Seminar Da)' I / Semillar Da)' 2 2 Seminar Da)' 3 2 Overall Notes 2

About the Software

2

CodeCalc: The PVElite Component Anal)'sis Pmgram Features 2 SUlJlmalY ojApplications 3

Chapter 2: Ol'erl'iew ofthe ASME Code History of the Code 1 Organization of the Code 2 Scope and Limitations of the Code 3 Allowable Tensile Stress per the Code 4 Internal pressure C)'lindCl's

011

shells and heads

5

5

Spheres, Elliptical Heads, Torispherieal Heads 7 Important Terms for analyzing cylinders and heads Geometry for Pressure Vessel Heads 8 Elliptical Heads

8

Hemispherical Heads

9

Geometry for Pressure Vessel Heads Torispherical Heads

10

/0

Press lire Vessel Design am/ Af1al)'sis

8

2

Chapter 3: Usil/g the Program-The 1I1ail/ Mel/II File Menu

1

New 2 Open 2 Save 2 Save As 3 Prillt 3 Prilll Prel';ew 3 Prilll Setup J

Exit 3 Previous FOlll' Files

Edit Menu

3

4

Title Page

4

Project Data

4

Insert New Item 4 Delete Current Item

4

Se/ecl All 4 Dese/ecl All 4

Analyze Menu

5

Browse

5

Analyze Selected Items

5

Analyze Current Componenl SlImmll1Y 5

Choose Analysis Type

5

5

Output Menu 6 Tools Menu 7 Configuratioll Options 7 Comput3tion Control Tab 7 Miscellaneous Options 9 Set Unit 9 Make UI/il File 10 Calculator I J Edit/Add Materials Dialog

Diagnostics Menu

JJ

13

CRC Check /3 Build Version Check J3 DLL Versio/1 Check 13

ii

Pressure Vessel Design find Analysis

View Menu 14 ESL Menu 15 Phone Updmc

J5

GClIemle Fax Codes

J5

ReceiJ'c lind Enler Fax Codes View ESL In/ormation J 5

Help Menu

/5

16

Camel/IS

J6

Tip of Ihe Da)' Illfo 16

16

Chapter 4: Example Problem l-A Simple Dl'lIm Problem 1 Specifications Brittle Fracture - Minimum Design Metal Temperature External Pressure on shells aud heads 21 Stiffening 27

10

What do )'011 need to kllow 10 analyze cylinders alld heads for external 28 External Pressure OUU1 28 Diameter 28

pressure?

Actual (or assumed) Thickness 29 Design Length for the Vessel or Vessel Segment 29 Width and Thickness of Reinforcing Rings 29

Nozzle Reinforcement and Failure Path Calculations Hillside and Off-Angle Nozzle Angles 45

36

What do you need 10 know /0 pelform nozzle reinforcement calClllatiolls? Required Thickness of Head or Shell and N07..z1e 50 Geometry ofNol.7.le and Shell 50 Diameter Limit, Thickness Limit 51 Is the Nozzle ill a Seam? 51 Details of Nozzle Welds 51 Largc Nozzle Considerations 51 Mallway or Access Opening 51

Press"re Vessel Design and Analysis

50

iii

Chapter 5: Cones and Conical Sections Guidelines for Cones

I

Typical Geometry for a Simple Cone

2

What do yOll need (0 know to perform cone calculations?

5

Dimensions of/he COile am/the Cylinders at Either End Dimensions o/Trallsifioll Klluckles allY exist) 5 HalfApex Angle of fhe COliC 5 Axial Forces 011 'he COile 5 Width alld 11/ickncss o.rCone Reinforcement 5

ar

Chapter 6: Welt/ed Flat Heads Guidelines for Welded Flal Heads

1

Whal do you need to know to analyze welded nat heads? Af/achme,,' Dctails 4 Required lIlId AClllal Thickness o/Ihe Shelf 4 Large ami Small Dimensions for the Head 4

Chapter 7: Half-Pipe Jackets Guidelines for Half-Pipe Jackets Flange Design and Analysis Gaskets 7

7

Gasket Materials and Gasket Factors 8 Other Gasket Types 12 Facing Sketches 12 Flange Types 13 Flange Behavior 16

Flange Stresses 20 Stress Analysis 20 Blind Flanges and Channel Covers Large Central Openings

38

46

Chapter 8: Floating Heads ami Spherically Dished Covers Types of Spherically Dished Covers Description Exmnple

iv

1

2

Pressure Vessel Design find Analysis

1

4

5

Chapter 9: Heat Exchallger Tubesheets TEMA Tubesheets ASME Tubesheets

1 17

Chapter 10: Expallsioll Joillts Flanged and Flued Expansion loints 1 Prc~surc Vessel Design and Analysis Seminar Component Design Problem Metal Bellows Expansion Joints 16

5

Pressure Vessel Design and Analysis Seminar Cornponcllt Design Problem

19

Chapter 11: Stresses Stresses in Shells due to Loads on Attachments

Discussion of Results

3

WRC 107 Stress Calculations 3 Why are the Stresses al Edge of the Pad the Same as at the Edge of the NOlzle? 3 What arc the Allowable Stresses? 4 IYRe/O? Stress SUlIlmatiollS 5 ASME Section VIII Division2-Elaslic Analysis a/Nozzle 6

Stresses in Horizontal Pressure Vessels

20

Chapter 12: Tall Vertical Towers Allowable Stresses on Tall Towers 1 Analyzing Tall Vertical Process Towers Design Procedure

2

3

Wind Load Computations 3 Wind Pressl/re Complllation 4 Earthquake Load Computation 5

Skirt and BaseRing Design

11

Basering Thickness Calculations 11 Thickness of Basering llnder Tension 14 Thickness of Top Ring under Tension 14 Basering Design Selections 15 Calculation of Required Area for Each Bolt 15

Selection of tile Bolt Size 15

PreSSl/re Vessel Design and Analysis

v

Selection of Preliminary l3asering Gcometly 15 Analysis of Preliminary Uasering Geometry 15 Sclection of Final Bascring Gcomctry 15 Analysis of Uasering Thickncsses 16

Skirt 11lickness Calculations

/6

13asic Skirt Thickness 16 Stress in Ski'1 due to Gussets or Top Ring 16

Chapter 13: Vessel Legs, Support Lugs, and Lifting Lugs Vessel Legs I Support Lugs 3 Lining Lugs 4

Bibliogmphy

vi

Pressure Vessel Design alltl Analysis

Pressure Vessel Design and Analysis - Seminar Notes

Chapter 1: Introduction to the Seminar & the Software

PURPOSE OF THE SEMINAR Weleome to COADE's Pressure Vessel Design and Analysis Seminar. The purpose of this course is to help you become comfortable with the guidelines for designing and analyzing pressure vessels, and to make you more effective at your job by introducing you to computerized design tools for pressure vessels-the CodeCale program and the PVElile program. The intent of this course is IlQ1 to have you memorize the exael rules of the ASME Code for pressure vessel design. Instead, we want you to become morc familiar with the ASME Code, be able to find what you need to know in it. "'nd above all be able to spot unrealistic results or questionable designs, whether gcncrated by you, by a computer. or by some other engineer. In a nutshell, we want to teach you what you need to know to successfully apply the ASME Code to pressure vessel design and analysis.

STRUCTURE OF THE SEMINAR Our approach to this goal will be to look at pressure vessels on a component by component basis. In other words, we will study separately each of the many pressure vessel components-sheils, heads, nozzles, flanges, tubesheets, etc. This approach is widely practiced by engineers as they design or analyze pressure vessels. It also allows us to start with relatively simple components and progress to more complicated ones. Most of the lecturcs in the course will have the following fonnat: Introduction to the theory of analysis for a particular component. Detailed review of ASME Code rules associated with a particular component. Design of the component using the CodeCale or PVElile programs.

Seminar Day 1 The first area to cover includes the history nnd structure of the ASME Code, calculation of nllowable stresses using Code rules, and review of the scope of vessels covered by the Code. This section of the seminar also inCludes a discussion of the design of some of the most basic components of pressure vessels: shells, heads, and nozzles. We will begin with rules for internal pressure design of shells and heads, along with associated mles for weld efficiency and brittle fracture. We will also learn how to use the CodeC~lc program, especially the SHELL program for internal pressllre analysis.

Introduction to the Seminar & the Software

1-1

Abollt the Software

Pressure Vessel Design and Analysis· Semin
We will then stndy the mles for external pressnre design of shells and heads. If time permits we will also study nozzle reinforcement and the design of Oal head and conical sections. This will complete our study of components typically associated with drums.

Seminar Day 2 On day two we will study components associated with heat exch~lI1gers. \\le will extellsively study flanges and bolted flanged connections such as those in TEMA channel covers, ASM E blind flanges, and floating heads ofheat exchangers. We will look at tubeshccts, including floaling, stationary, and fixed tubesheet designs, as well as metal bellows expansion joints and flanged and nued expansion joints, commonly lIsed in heat exchangers to absorb differential thermal stress.

Seminar Day 3 On day three we will focus on tall vertical pressure vessels and vessel supports. We will review the loadings on tall vessels, such as wind and earthquake. We will also review the design of vessel skirts and base rings, horizontal vessels on saddle supports, and legs or support lugs which arc commonly used for smaller yessels and exchangers. We will also study yessel-shell stresses that arc caused by loads on attachmcnts, including nozzles and support lugs.

Overall Notes This notebook is arranged to allow yOll to work step-by-step through the course, and to work example problems of each type of calculatiou. Plenty of space has been left to take additional notes. We want to encourage you to ask questions and make comments during the course so that we can cover the material you 1110st want to learn. A three day course is really 100 short to fully explore both the theoretical and the practical aspects of the ASME Code, but we do want to cover as much of these topics as possible. Your input will make this easier and more effective.

ABOUT THE SOFTWARE CodeCalc: The PVElile Component Analysis Program The CodeCale program is a package of seyenleen applications for the design and analysis of pressure vessels and heat exchangers. The purpose of the program is to provide the mechanical engineer with easy 10 use, technically sound, well documented calculations that will speed and simplify the task ofyessel design or re-rating. Calculations in the CodeCalc program are based on the latest editions of national codes such as the ASME Boiler and Pressure Vessel Code, or industry standards such as the Zick method of analysis for horizontal drums. The CocicCalc program offers exceptional case of usc, which results in dramatic improvement in efficiency for both design and re-rating. One expert estimates that the time he needs (0 rcrate an exchanger has been reduced from eight hours to two hours.

Features The following arc features of the CorieCalc/PVElile program:

1-2

Introduction to the Seminar & the Softwmc

Pressure Vessel Design and Analysis. Seminar Notes

Abollt the SoOwarc

Thc Windows cnvironmcnt provides cxtcnsive on-line help at a keystrokc, and manages multiple analysis files so that thc USCI' CHn define a whole pressure vcssel in a single file. The USCI' Cfln define his own unit systcm, opening CodeCfllc 10 thc world of metric and Sf calculations. Internally, however, calculations continuc to be in the English systcm of units, assuring continued compliance with ASME Code requirements. The CodeC"le progmm has " complete m"teri"llibmry including over 2000 t"bles of allowable stress versus temperature and 49 external pressure charts. The CodcCalc program also includes a component library which contains diameter and wall thickness for all standard pipc sizes, pressure vs. temperature charts for ANSI B 16.5 nanges, "nd section properties for AISC beam sections. A sUlllmary capabi lity allows evaluation of all the components of a pressure vessel or hcat exchanger. Design pressure, temperature, material, and Maximuill Allowable Working Pressure are shown for each component. Printed output from the CodcCalc program is exceptionally clear and complcte, with user definable headings on each pagc. User comments and additions Illay bc inserted at any point in the output. The ability to save any analysis to disk m
Summary of Applications The following applie"tions are avail"ble in the CodeC"le Program: SHELL-lntern"1 and external pressure design of vessels and exeh"ngers using the ASME Code, Section VIII, Division I mles. Components include cylinders, conical sections, elliptical heads, tori spherical heads, nat heads, and spheric"1 shells and heads. This program calculates required thickncss and maximum allowable intern
Introduction to the Scminar & the Softwmc

1-3

About the Sofiwarc


lations technique allowed by the Code (Soehren's ealeulation) is also imple-

mented by this program. FLANGE-Stress analysis and geometry selection for all types of flanges using Ihe ASME Code, Section VIII, Division I mles. This program bOlh designs and analyzes the following types of flanges: Weld neck flanges and all integral flange types Slip on flanges and all loose flange Iypes with hubs Ring type flanges and all loose flange types without hubs

Blind flanges, both circular and

non~circlliar

TEMA channel covers

Reverse geometry weld neck flanges Flat faced flanges with full face gaskets LGCENTER-Stress analysis for flat heads with a large, central circular opening based on ASME, Section VIII, Division I, Appendix 14. The program calculates the stresses at both the OD of the head and the location of the opening. TUI3SHT-Analysis of aillypes of lubesheets using Ihe Seventh Edition of the Standards of the Tubular Exchanger Manufaelurers Association. The program takes full account of the effects of tubesheets extended as flanges, and'for fixed lubesheets also includes the effects of differentialthemlal expansion and the pres-

ence of an expansion joint. ASMETUI3E-Analysis of several types of tubesheets using the mles from Appendix AA of the ASME Code, Section VlIl, Division I. This appendix provides alternate mles for tubesheet design, and may result in thinner tubesheets than the TEMA program. HORIZVES-Slress analysis of horizontal dmms on saddle supports using the method of L.P. Ziek. Results include stresses at the saddles, the midpoint of the vessel, and in Ihe heads. Stiffening rings used inlhe design of the vessel arc also evaluated. LEG&LUG-Analysis of vessel support legs, support lugs, and lifling lugs. This anal-

ysis is based on industry standard calculation techniques, and the resulting stresses are compared to the AISC Handbook of Steel Construction or the ASME Code. A full table of AlSC beams, channels and angles is included in the program. PIPE&PAD-Required wall thickness and maximum allowable working pressure for

two pipes, and branch reinforcement rcquirements for the same two pipes considered as a branch and a header. I3ased on ANSI 1331.3 mles, this program includes tables of outside diamcter and wall thickness for all nominal pipe diameters and schedules. WRC 107-Stresses in cylindrical or spherical shells due to loading on an attachment, using the method of P.P. I3ijlaard as defined in Welding Research Council Bulletin 107. I3ASERING-Thickness calculations and design for annular plate base rings, lop

rings, bolting, and gussets. Thesc calculations are performed using industry stan(!cud calculation techniques.

1-4

Introduction to the Seminar & the Sofiware

Pressure Vesscl Dcsign and Analysis - Seminar Notes

About the Sofiware

THINJNT-Calculates stress in a metal bellows expansion joint of the type typically

used in piping and heat exchangers. The program does elastic stress analysis for stresses due to internal pressure and opening or closing of the joint, and calculates the cycle life of the joint based on the ASME Code, Section VIII, Division I, Appendix BB. THICKJNT-Calculates stresses in a heat exchanger expansion joint fabricated from relatively thick plate, also called flanged and filled expansion joints, using the mles of the TEMA Seventh Edition, Paragraph RCB-8. The analysis is based on the equivalent geometry used in "Expansion Joints for Heat Exchangers" by S. Kopp and M. f. Sayre, with slight modifications. RECTVES-Perfonns stress calculations and Maximum Allowable Working Pressure calculations for the rectangular, obround, and circular vessels described in the ASM E Code, Section VIII, Division I, Appends 13. The calculations arc taken from Sections 13-6 through 13-13. SUMMARY-Description and evalnation of all the components ofa pressure vessel

or heat exchanger. Design pressure, temperature, material, actual thickness, and Maximum Allowable Working Pressure are shown for each component.

Introduction to thc Scminar & thc Software

t -5

About the Software

1-6


Introduction to lhe Seminar & the Soflwarc

Pressure Vessel Design and AllCllysis - SeminClr Notes

Chapter 2: Overview of the ASME Code HISTORY OF THE CODE The Boiler Codc has becn in existence for almost 90 years. During the 1800's there wcre numerous catastrophic failures of prcssure vessels and boilers, resulting in thousands of deaths. Public and professional concern resulted in the release of the first mles for power boilers in 1915. The first pressure vessel Code was issued by the ASME in 1925. In 1968 the Code was divided into two subsections: Section VIII, Division 1 - Rules for Constmction of Pressure Vessels. Section VIII, Division 2 - Alternative Rules for Pressure Vessels. The main differences between Division 1 and Division 2 are Division 2 has higher allowable stresses in most cases - resulting in thinner vessels. Division 2 also has more sophisticated design requirements, requiring more extensive stress analysis than is used in Division J. Division 2 frequcntly requires a fatigue analysis. Division 2 frequently requires more extensive inspection and record keeping than Division In 1992, the Code moved the allowable stress tables and external prcssure charts for materials from Section VIII, Division 1 to Section II, Part D. This new publication contains all of the material data from previous editions of Section Vlll, Division I, plus a few new reference tables from other sections. Division 3, whieh is a new division intended fOf high prcssufe vessels, has been in preparation for sevcral years, and is now issued. Section VIII, Division I is by far the most widely used of the two current divisions. Almost all the pressure vessels constmeted in the U.S. are constmcted to Division 1. In this course, unless specifically noted otherwise, "The Code ll is Section VIII, Division I. ASME approves Code Case 2290 which increases Division I allowable stresses.

Overview orlhe ASME Code

2-1

Organization of the Code


ORGANIZATION OF THE CODE The Code is divided into three Subseelions, pins Mandatory Appendices and NOl1Jnandatory appendices

Thc three subsections cover A - General Requiremcnts

B - Requirements Pertaining to Methods of Fabrication of Pressure Vessels Part UW - Requirements for Pressure Vessels rabricated by Welding Pari ur - Requirements for Pressure Vessels rabricaled by rorging Part UB - Requiremenls for Pressure Vessels rabricated by Brazing. C - Rcquircmcnts Pertaining to Classes of Materials

Part Part Part Part Part Part ParI Part

UCS - Carbon and Low Alloy Steels UNr - Nonferrous Materials UCI - Cast lron UCL - Clad and Lined Vessels UCD - Cast Ductile Iron UHT - Heal trealed ferritie steels ULW - Layered Construclion ULT - Materials with higher allowable stresses at low temperature

Mandatory appendices cover snbjects nol covered in the main body of the Code. The requirements of these appendices are mandatory when the subject covered is appropriate 10 the eonslmction oflhe vessel.

Nonmandatory appendices provide information and suggest good practices relative to prcssure vessel construction. They also scrve as a place where new design rules 3rc introduced and tested before moving into the mandatory requirements.

2-2

Overview of the ASME Code


Scope and Limitations of the Code

SCOPE AND LIMITATIONS OF THE CODE Section VIII, Division I applies to all pressurized containers, but with Ilumerous exceptions. Some of the exceptions listed in paragraph U-I include Vessels within the scope of other sections (i.e. power boilers). Fired process tubular heaters (furnaces) Pressure containers that arc part ora machine (i.e. pumps) Piping or piping components Pressurized water storage up to 300 psi Ileated water storage up to 210°F Vessels with design pressure 15 psi or less Vessels with a maximum cross section dimension 6 in. or less Vessels for Human Occupancy Division I rules are applicable to vessels not exceeding 3000 psi design pressure. You can usc Division I above 3000 psi, but special precautions to avoid fatigue and other additions to the mles are recommended. The scope of Division I includes the nozzles and attachments to the vessel. Unfired steam boilers may be constmeted to Division 1 or Section I. Some classes, such as evaporators or vessels in chemical plants fife required to meet Division I. SOIIle small vessels are exempt from inspection:

Up to 5 cubic feet at 250 psi Up to 1.5 cubic feet at 600 psi ANY Vessel that meets ali the requirements of Division I may be stamped with a U stamp even though exempted by one of the above limitations.


2-3

Allowable Tensile Stress per the Code

Pressure Vessel Design anu Analysis - Seminar Notes

ALLOWABLE TENSILE STRESS PER THE CODE Division I mles arc based on a maximum principle stress failurc theory. While not very accuratc, this thcory is simple to understand and apply.

By way of contrast, Division 2 is based on a maximum shear stress theory, which is more accuratc. Somc other Codcs are based on the even more accurate theory of maximum distortion encrgy (Von mises stress). The allowable tensile strcss for materials in Division I is the minimnm of the following

stresses: 1/3.5 ofthc specified minimum tensile strength at ambient temperature 1/3.5 of the tensile strength at the design temperature

2/3 of the specified minimum yield strength at ambient temperature 2/3 of the yield strength at the design temperature 100% of the average stress to produce a creep rate of 1% in 10,000 hours. 67% of the average stress to produce rupture in 100,000 hours. 80% of the minimum stress to produce rupture in 100,000 hours. In 1999, the ASME changed the 1/4 to 1/3.5. In the temperature range in which tensile strength or yield strength set the allowable

stresses, higher allowable stresses arc permitted for austenitic stainless steels and nickelalloy materials where greater defomlation is not objectionable. In this case the critcrion of 2/3 yield strength at temperature may bc increased to 90% of yield strength at temperature. Bolting materials whose strength has been enhanced by heat treating or strain hardening are limited to 1/5 of tensile and 1/4 of yield.

2-4

Overview orllle ASME Code

Pressure Vessel Design

Analysis ~ Seminar Notes

Internal pressure on shells
INTERNAL PRESSURE ON SHELLS AND HEADS Cylinders Theoretical derivation of stress for thin walled cylinder: Hoop Strcss: (circumferential)

Force

~

PXD;XL

A reo = 2 X Force Area

--

/ X

L

I'D; orl 21

-

~

I'D; 2S

Axial Stress: (longitudinal)

Force

Area

Pxnx

=

=

n x D;

Force Area

I'D; 4/

X

(-iD)2 t

or 1

I'D;

4S

For thick-walled cylinders the theoretical stress is expressed by the Lame equations. The formulation of the Lame equations is as follows, for internal pressure only. The maximum stress at the ID surface is:

Hoop

Rodial

~

-I'

ASME has a slight variation on the first formula, making it fit fairly closely the rcsult of thc sccond fommla:

PR SE - 0.61'

The

aD basis form of the samc equation is:

SE + 0041' The ID and OD formulas do not yicld exactly thc same results: for relatively thick walled

cylinders, the effect can bc noticeable. The following graph shows the relative accuracy of these three formulas. The Lame equation is exact for all geometries. Thc simple approximation becomes pretty bad for thick-


2-5

Internal pressure

011

shells and heads

Pressure Vessel Design and Analysis - Seminar Notcs

walled cylilHJers. The ASME equation is much closer to the exact solution than the simple approximation.

2M

T

10"0 -

-

to

~" "....., - -- - -"'" ~

~

~

.10

1.0

,.5

2.0

2"S

--

--

EO.ls"9l EQ(8.1

EO.(s.l)

35

3.0

4.0

"~ figure '.6

"

Compori.on ol fom,vlai !of Op

Il'ml

r.

0

cy!i,*kol

Votl ••

The Joint Efficicncy in this (and all othcr) ASME Code formulas is a measure of the inspeclion quality on Ihe weld seam. In general, weld seams that receive full radiography have a joint efficicney of 1.0. Weld scams that receive spot radiography have a joint efficiency of 0.85. Weld seams that receive no radiography have a joint efficiency of 0.7. Seamless components have ajoint efficiency of 1.0.

In addition to the basic mles described above, the Code requires that no two seams in the same vessel differ in joint efficiency by more than one category of radiography. For example, ifcireumferential seams receive no radiography (E=0.7) then longitudinal scams have a maximum E of 0.85, even if they receive full radiography. The practical outworking of this is that circumferential seams, which are usually less highly stressed, may be spot radiographed (E=0.85) while longitudinal seams arc fully radiographed. This provides the sallle metal thickness at some savings in inspection costs.

2-6

Overview or the ASME Code

Splleres, Elliptical I-leads, Torispherical Heads


SPHERES, ELLIPTICAL HEADS, TORISPHERICAL HEADS In Basis

00 Basis

~ p(D12

+ CAl SE- 0.61'

Cylinder

Elliptical Hcad

~ p[K(D + CAl]

Spherical Head

DI2 + CA) f ~ l' ( 2S£-0.2P

Torispherical Head

f ~ p[M(L + CAl] 2SE-0.lp

2S£ - 0.21'

l' (D/2) SE + 0041'

~ 1'(

~

I

Pa

Pa

2SE(I - ea) [K(D + 2ea) + 0.2(1 - ea)]

Spherical

Pa

SE(I-ea) (DI2 + ea) + 0.2(1 - ea)]

Pa ~

Pa ~

SE(I - ea) [(D/2 + ea) + 0.6(1 - ea)]

Elliptical

Torispherical

1'[

KD ] 2SE+2p(K-OI) D/2 ) 2SE+ 0.81' ML ] 2SE+ 0.785P(M - 02)

OD Basis

In Basis Cylinder

1'[

f -

SE(I-ea) [M(D + ea) + 0.1 (1- ea)]

Pa

Pa Pa

SE(f-ea) [D/2-0A(I-ca)] 2SE(I- cal [KD-2(1 + ea)(K - 0.1)1 SE(I - ea) [D12 - 0.8(1 - ea)]

SE(t-ea) [(MD - (I - ea))(K - 0.2))

The fonnulas for elliptical and torispherical heads are general. The factors M and K are semi-empirical adaptations of more complicated shell theories. K and M have the following fommlas: Where h ~ depth of head L r ~ knuckle radius

~

crown radius

For the special case of a 2: I elliptical head,

K~1.

For the special case of a nanged & dished (6%) torispherieal head, M ~ 0.885 and the

inside crown radius equals thc outside diameter of the vcssel. The shape of elliptical and torispherieal heads produces compressive stresses at the knuckles.

ror thin torispherical heads the eqliatiol1 in Division I is unconscrvativc-dimpling can occur even in vesscls that meet the Code requircments. Division 2 contains a more complicated equation that should be checked when the value of rlt is large.


2-7

Important Terms for analyzing cylinders and heads


IMPORTANT TERMS FOR ANALYZING CYLINDERS AND HEADS ALLOWABLE STRESS-Sclect the allowable slrcss from the appropriatc Code table at thc design temperature. JOINT Erf'ICIENCY-Sclcct Ihejoinl cfficiency from lable UW-12 DIAMETER (INSIDE OR OUTSIDE)-ID formulas are found in paragraphs UG-27 and UG-32 00 formulas arc found in Appendix I CORROSION ALLOWANCE-Subtractthc corrosion allowance from the actual thickncss, and [liso increase the inside diamcter to account for corrosion. ACTUAL THICKNESS ANDIOR DESIGN PRESSURE-You can calculate maximum allowable working pressure if you know the thickness of the component. Otherwisc, use the design pressure to calculate the required componcnt thickness. ASPECT RATIO f'OR ELLIPTICAL HEADS-This is typically 2:1, but may range from I: I to 3: I CROWN RADIUS AND KNUCKLE RADIUS FOR TORISPHER1CAL HEADSThe ratio ofcTOwn radius to knuckle radius may not bc less than 1 nor grcater than 16.66

GEOMETRY FOR PRESSURE VESSEL HEADS Elliptical Heads

Major A~is

(Head Diameter)

Minor Axis (1/2) Head Diameler

(Aspect mtio = ratio of major nxis to minor axis, Iypically 2.0)

2-8

Overview oftbe ASME Code

Pressure Vcssel Design ;:md Analysis - Seminar Notes

Geomctry for Pressure Vessel lIeads

Hemispherical Heads

/""""""---

--------

-'-'~-........ "

/

"""\.

/

,I

I

\

I

\ ---,I

!

I


- -

HCild Di;nnctcr

--~

\

2-9

Geomelry ror Pressure Vesscilleads


GEOMETRY FOR PRESSURE VESSEL HEADS Torispherical Heads

-r------'" I

. . , . . . . --.. . . .

I

_/I

IIEAO DIAMETER

). /

KNUCKLE \ RADIUS Ir)

I

/

I I

_I

til

[$1

q I

"I

"'/

i!

",/

vi I

I

/

I

I !

I

The typical torispherieal head, also known as Flanged & Dished, has a crown radius equal to the outside diameter of the cylinder, and a knuckle radius equal to six percent of the cyl-

inder diameter.

2-tO

Overview or tile ASME Code


Chapter 3: Using the ProgramThe Main Menu CodcCalc always starts with the Vessel Data Input Screen. Across the top of this screen is a line of itcms that is called the Main Menu. The Main Menu controls the major fUllctiolls of the program. \Ve will review the functions available in each of these menu items. The items in the Main Menu - file, Edit, Analyze, Output, Tools, Diagnostics, View, ESL, and Help - may be selected with a mouse click or by pressing the underlined character while pressing the Ait key. for example, the Output processor may be selected by pressing the Ait and 0 keys simultaneously. First, we will begin by going over each of the Main Menu items.

FILE MENU The File Menu controls the general operations of CodeCalc files. Options that are displayed in the menu with an ellipsis ( ... ) cause a file manage window to appear when selected.

Figure 1--The File Menu

Using the Program- The Main Menu

3-1

File Menu

Pressure Vessel Design and Analysis

~

Scminar Noles

New Starts a new file. [jo t~

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o c& "fa nOlO

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Figure 2--File New

Open Opens a previously created file. When the Open option is chosen, the user is prompted to select an existing job file. Files ortype '.ee; will be displayed for selection. ~

?X

Figure 3--0pen Dialog

Save Saves the current file in its present condition.

3-2

Using the Program- Thc Main Menu

Pressure Vessel Design and Analysis - Seminar Nolcs

file

MCIIU

Save As Saves a file that has not been previously namcd or saves the current file under another namc.

fW·i ::!:lY.o~_11,cd

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Figure 4--Save As Dialog

Print Sends the current vcsscl graphic image directly to a postscript or laser jet printer.

Print Preview Displays the page that will be sent to the printer (sec above).

Print Setup Brings up the standard Windows printer setup screen,

Exit Exits CodeCalc. A message window will appear to give the user a last opportunity to save any modifications to the current job.

Previous Four Files The File Menn also lists the last fOUf vessel inpullilcs accessed from your computer. Any of these liles may be opened with a mouse click.

Using the Program- The Mnin Menu

3-3

Edit Menu


EDIT MENU Once a filc is selectcd, the Edit Mcnu indicates the options available for ediling. A3Irujll'.I."juOnli,Nj 1.i,.l.tmiif.!:·1.it1,3i

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Figure 5-- Tile Edit Menu

Title Page Allows the user to enler report titles for this group of rep orIs.

Project Data Allows the user to enler up to 3 Ii tie lines, which appear at Ihe lop of each page of the prillted reports.

Insert New Item Inserts a new element after the current element.

Delete Current Item Deletes the current clement.

Select All Selects all of the items in the browse window.

Deselect All Deselects all of the items in the browse window.

3-4

Using the Program- Thc M"in Mcnu


illHJ

Analysis - Seminar NOles

Analyze Menu

ANALYZE MENU The Analyze options cause the program to quit the input process and enter the analysis process. CodeCalc will first save the current job to the input file with the same filename; Ihcn il will process the analysis.

Browse Allows the seleclion of certain components in Ihe input lile to be analyzed.

Analyze Selected Items Performs calculations for selecled analysis Iypes. The calculations will be saved in a binary file and will be ready for display or printing.

Analyze Current Component Performs calculations for the current analysis type. The analysis program looks for appropriate data in the current analysis file and performs calculations, saving the results in a text file. The results oflhc analysis will then be ready for display or printing.

Summary Looks through all the data in the current analysis file and prepare a brief summary of cach analysis.

Choose Analysis Type Selects Ihe Iype of componenl you wish to work on.

Figure 5--Choose Analysis Type Menu

The analysis types chosen from this menu can also be selected from Ihe Analysis Tool Bar by simply clicking on the icon.

Figure 7-- The Analysis Types Toolbar


)-5

Output MCllu

Pressurc Vessel Design and Analysis - Scminar Notes

OUTPUT MENU The Output Menu allows the user to review the analysis results and print (hc graphics of

the vessel. The following option is available under Output:Review - allows the user to review the analysis results of the clIrrcJ1tjob, if those results are available.

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f

OUC1f,6orld~t.eI Sedi:tl: ISPH£RlCAlliEAO

Figure 8--Ti,e Output Menu

}-

3-6

Using the

Program~

The Main Menu

Pressure Vessel Design and Analysis - Seminar Noles

Tools Menu

TOOLS MENU The Tools Menu controls the utility processors as summarized here.Configuration - This option allows the user to dcfine a variety of system variables for the program. The first screen of the Configuration mcnu looks like this:

!

pic5lHljlf.lllliS 1i0!!!ij4B H.'i'iMl.• f1r1.i!'. 33.1 . 1m O~lb:.I i,SL ~~ .u~~ ! 0 cHi I e;J,' -f: '(0:(9'1"L -',: ]-1Il1l3 ill ,De £.li. hWfli .D~

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Figure 9-- Tile Tools Menu

Configuration Options Computation Control Tab The Computation Control Tab in thc Configuration dialog leis some specific program computation control parameters be set. These controls Icl you set some options in some programs that control the results of some computations.

Figure 1a--Configuration Options

Following is a description of the options: Computc Increascd Nozzle Thickness? In many cases pressure vessels are designed and built long before the piping system is attached to them, This means that the nozzle loadings are unknown. lfthis field is checked, thcn your minimum nozzlc thickness (tm) will be the maximum of trn Irn

= (.134,trn for internal pressure) less than or equal Nps 18 = (DD/ISO,trn for internal pressure) greater than Nps 18

By using such a requirement in addition 10 UG-45, the piping dcsigners will have some additional metal to work with to satisfy thermal bending stresscs in systems these vessels are designed for.

Using the

Progral11~

The Main Menu

3-7

Tools Menu

Pressure Vessel Design and Analysis· Seminar Notes

Note

These fonnulae are not in the ASME Code. They are used in industry.

You can also specify the minimum wa)) thickness of the nozzle (Tm) in the Nozzle input. If you do so, that will override this calculation. Calculate F iu Flohead if the Pressure is Zero? In the design of noating heads, a factor F is computed. The factor F is a direct function of the internal pressure. If the internal pressure is 0, then F is equal to 0. However, some interpret the Code to mean that F should always be computed regardless of which case we arc analyzing. Typically, the case in question is the flange bolt-up case. When the uni' is being bolted up, it has 110 internal pressure. That is why the defanlt is not checked. (fyou wish F to always be considered in the thickness cales, then check this box. This is the conservative method of calculation. Use P iustead ofMAWP for UG-99B? The Code paragraph UG-99(b) discusses the subject of hydrostatic test pressure on vessels. The equation that wonld nonnally be used is as follows: Test Pressure = 1.3' MAWP • StestfSdesign The code in note 34 states that the MAWP may be assumed to be the same as the design pressure when calculations arc not made to delcnnine the MAWP. This will allow for lower test pressures. This directive should be used with caution. Perform Area Calculatious for Small Nozzles? The Code paragraph UG-36 discusses the requirement of performing aTea placement calculations when srnall nozzles arc involved. The Code States Openings in vessels not subject to rapid fluctuations in pressure do not require reinforcement other than that inherent in the constmetion under the following conditions: 3.5-in. finished opening in a shell or head .375 in. thick or less 2.375-in. finished opening in a shell or head greater than .375 in. If your geometry meets this criteria and this box is nol checked, then no area of reinforcement calculations will be perfonned. Priut Water Volume jn Gallons? Normally the volumes computed by the program arc in diameter units. ]fyou want to use US gallons instead of cubic diameter units, check this directive. Otherwise, the program will use cubic units as the default value. Use Calculated Value of M for Torispherical Heads in UG-45 bI? The Code in paragraph UG-45 requires a calculation of the required head thickness at the location of the nozzle. This may lead one to believe that the thickness Inay be computed per paragraph UG-37. However a recent code interpretation states that the thickness should be computed by the mles of paragraph UG-32 or by the rules in Appendix I. Thus, this directive should always be checked. The second screen of the Configuration Menu looks like this:

3-8


Tools Menu


Miscellaneous Options The Miscellaneous Options of the Configuration MCllulets the user select some miscellaIleous directives. These directives control some printout style options and others. ?Ixl

• ~[>1Jom3!,,·rlWr.RCNn1 nA~~fio~rl~?"-

Oda<.iUr.sfle Kr.;lllhll

iJ1

"-

Figure 11--Miscellaneous Options

Following is a description of the options: Report Content. This directive allows thc uscr to change the length of the printcd reports. When the summary option is checked. the formulas and substitutions will not be printed out. Thus, this option will generate less paper and more compact reports. Whcn thc dctailcd option is chccked, the reports will be the normallcngth. External Printont in Rows? There are two choices for thc style of printing extemal pressure results: rows and columns. Printing the values by row tends to reduce the length of the printouts. This is the default. If you wish to print by column, do not check this directive.

Set Unit This option allows the user to change the current job's units system. Once this option is selected, a File Open dialog will appear and allow the user to select a new units file. Thcsc

Using the Progmm- The Main Menu

3-9

Tools Menu


units files have the extension ,fil. English, Metric and SJ units are available in the system subdirectory. After you select a units file, the following window will appear: ~,lllJfe

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"SystemUnl'

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Figure 12..Unil Window

If the units selection is acceptable, press the OK bullon; otherwise, press Cancel. When OK is selected, the current units will be overlayed with the selected units.

Make Unit File This option allows the creation of a cuslom units file. Simply pull down the appropriate conversion constant or label and the corresponding unit or label will change accordingly. If your conversion constant is not one ofthe choices, type in the label and constant for your particular unit. (The program will continue to use English units internally).

Figure 13..Make Unit Dialog

This window presents a table of items, the internal units used for each item, a conversion factor, and the user lin its. The conversion factor is used to obtain the user units from the internal units. The lip and down arrow keys can be used to move lhe selection to the desired item. If a desired unit conversion is not available as a default program selection, it can be entered manually by typing it in. Ensure that your conversion constants arc correct and that your labels go with the constants. Once all units have been set, press OK to exit

3·10


!

Pressure Vessel Design ;lnd Analysis - Seminar Notes

Toob Menu

this screen and save the new units file. A safe place to save it would be in the system subdirectory where Ihe supplied units files arc stored.

After you have saved the new units file, you will need to overlay the current units in your job file with Ihe new units. This option is Ihe Set Unit option. After you sel your file with

the new units, all of the entered data will be converted into the new set of units immedialely.

Calculator This option allows the user to perform simple calculations and paste the results in the input field in which the cursor resides. .

Figure 14--The Calculator Option

You can use the calculator to compute a number and transfer that number into CodeCalc by using the Edil, Copy fealure. From the desired field, right click and choose the Paste

option. Before pasting, ensure that the field's current contents have been removed. Import Nozzle Dala - Imports nozzle information from a PVElilc inpul file (.pvi) for use in the WRC 107 Module. Edit!Add Materials - This opliou allows the user to add materials 10 Ihe COADE Material dalabase. The screen appears as follows:

Edit/Add Materials Dialog To use this processor, fill in all of the values in all cells. If more than one material is to be

entered, usc the Next button to enter the new material. After all materials have been


3-11

Tools Mellu


entered, save the file with the Save button. Finally, press the Merge key to join the userdefined material database with the supplied material database. xl

Figure 15--Ma/erial Editor

• "

3-t2


Pressurc Vessel Design ;lIId Analysis. Scminar Noles

Dillgnostics Menu

DIAGNOSTICS MENU The Diagnostics Menu helps to troubleshoot problem installations.

e'''''j.. ·''iliji'.I:!liWj4 f.!"l.!M\i¥ fI"i,I,,_,!

IlemUI.ITbl:'f:~ ! Dtw'~" Shel Sedl«ljSFH[Rl~ HEAD

)

Figure 16--Diagnoslics Menu

CRC Check Performs a cyelic redundancy check (CRC) on each of the supplied CodeCale files.

Build Version Check Checks the revision level of the CodeCalc executablc files.

DLL Version Check Checks to make sure thc CodcCale .DLL files are currcnl.

Note

If the DLLs are not current, the program may behave in an unusual manner or may not mn at all.

Using the Progralll- The Main Menu

3-13

View Menu


VIEW MENU The View Menu allows the user to specify the tool bars to be displayed.

.~ .

,

Figure 17--The View Menu

The following options are available: File Toolb.r

Figure 18--File Toolbar

Analysis Toolbar

Figure 19--Analysis Toolbar

)

Status Bar -

~-

Figure 20--8ta/us Bar (a/ the bottom of/he screen)

;

;

3-14


E51 Menu

Pressure Vessel Design lind Analysis - Seminar Notes

ESL MENU The ESL Menu provides utilities lhal interact with the Extenlal SoOw",e Lock (ESL).

e .. U1J....\IlIju.':M,'lj.J.!1 1.!@4f.!li',.HI

"",--,:::;->q~p:~=--.:::=:::;,~y_ l:itb '''''--l''h?MUpd~e

1-'---.:....=-,'----.:....--'-----4l

1i~alfI Fa:< Co.:le-\~' B~~[rtl'"h;cCode:l

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Figure 21--ESL Menu

Phone Update Allows update authorization information or other ESL changes to be obtained over the phone.

Generate Fax Codes Provides the uscr with access codes for rcmote ESL updating. These access codes should be sent to COADE for authorization codes.

Receive and Enter Fax Codes Allows yOll to enter the remole authorization codes you receivcd from COADE. Each set of four codes will make one change to the data stored on your ESL.

View ESL Information Displays the data stored on the ESL.


3-15

Ilclp Menu

Pressure Vessel Design and Antilysis· Seminar Noles

HELP MENU The Ilelp Menu displays on-line Help and infonnation on how to obtain technical support for CodeCale. II

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. tcir

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lY",*e QoJpA. IIX4 Obt;Jl~l- ~Sl

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11-00').) . . . . . . . oiI.

Figure 22--Help Menu

Contents Starts the Help facility.

Tip of the Day Provides tips for running CodeCale.

Info Provides information on the best ways to contact COADE personnel for teclmical support, and provides a link to COADE's website.

3-16

Using thc Progmm- Thc Main MCllu

Pressure Vcssel Design and Analysis - Scmin:lr NOlcs

Chapter 4: Example Problem 1A Simple Drum PROBLEM The drawing on the following page shows a simple horizontal pressure vessel that we will use for our first example problem. In this case the preliminary sizing of the vessel and its attachments has been completed, and we are asked I) to select thicknesses for the pressure components, nozzles, and reinforcemcnt, and 2) to check the vessel for stresses when it is full of liquid.

SPECIFICATIONS Design conditions, as shown on the drawing, arc 230 psig (1.586 N/1l1l11 2) ii;ternal pressure and fnll vacuum at 450°F (232°C). Materials arc carbon stecl, SA-516, 70, Nonnalized. The vessel is subject to full radiography, and has a 1/8-in. (3.175 mm) corrosion allowance. The length of the vessel is 244 in. (6198 mm) between langent lines (the heads have a 2-in. (51 mm) straight flange, making the weld-to-weld length of the vessel 240 in. (6096 mm)). We will require the following programs in this analysis: Internal pressure on shells and heads - SHELL program, External pressure on shells and heads - SHELL program. Nozzle thickness and reinforcement - NOZZLE program. Follow these steps to complete this first example problem: I.

Analyze the cylindrical shell under internal pressure, and pick a design thickness for

it. Next analyze the 2: I elliptical head, using the same basic thickness. 2.

Analyze the cylinder under full vacuum conditions. Also,. analyze the heads for external pressure.

3.

Analyze the nozzle reinforcement using the NOZZLE program. Select appropriate reinforcing pads for each nozzle.

4.

After the entire vessel has been analyzed using CodeCale, model the same vessel under PYElite and review the results.

Example Problem 1- A Simple Dnllll

4-1

i

144.0'10

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230 PSIG & Full Vacuum Design Temperature: 450 F

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........

Pressure Vessel Design and Allalysis· Seminar Noles

Specillcnlions

Pressure Vessel Design and Analysis Seminar

Component Design Problem

D-101 Cylindrical Shell:

Design Pressure

230 psi (1.586 N/nlln 2)

Design Temperature

450°F (232°C)

Material

SA 516,70

Joint Efficiency Corrosion Allowance

Inside Diameter

Notes:

Questions:

What is the required thickness?

What thickness will you usc?

DO NOT TURN THE PAGE UNTIL YOU HAVE COMPLETED YOUR ANALYSIS

Example Problcm 1- A Simplc Drum

4-3

Pressure Vt:sscl Design nnd Annlysis· Scminllr Noles

Spcciliclllions

COAoE Engineering Soft.ware pVEl i t.e 4 . 00 Licensee: COADE I llC., l.oca 1 Hhj te Lock PileName : Seminar ----- •••. --- ...••• --------------- .••••• Page 2 Shell Analysis: 0101 CYLINDER Item: 1 9:51il Sep 21,2000 Input Echo, Component

"

Design Internal Pressure l'empen\ture for Internal Pressure

Design Length of Section Length of Cylinder [or Volume Cales. Inside Diameter of Cylindrical Shell

i

Description: 0101 CYLINDER p

230.00 450.00

psig p

NO

Include Hydrostatic Ilead Components I'laterial Specification (Normalized) Allowable Stress At Temperature Allowable Stress At Ambient Curve Name for Chart UCS 66 Joint efficiency [or Shell Joint

,

S SA

SA-516 70 20000.00 20000.00

g

1. 00

L

244.0000 244.0000 144.0000

in.

1.0000 0.1250

in. in.

psi psi

D

CYLLEN [)

T

f·linimum Thickness of Pipe or Plate Corrosion Allowance

CA

Type of Element:

Cylindrical Shell

in. in.

INTERNAL PRESSURE RESULTS, SHELL NUI·IBER 1, Dese.: 0101 CYLINDER ASHE Code, Section VIII, Division 1, 1998, A-99 Thickness Due to Internal Pressure (TR): (P*(D/2 .. CA)/(S*E-0.6*P) per UG-27 (c) (1) (230.00*(144.0000/2 .. 0.1250))/(20000.00"1.00·0.6*230.00) 0.8352 in.

.•>'

Nax. All. working Pressure at Given Thickness (HAI1P); (S*E* (T·CA) 1 (D/2+CA) +0. 6* (T·CA)) per UG-27 (c) (1) (20000.00*1.00*(0.8750))/«144.0000/2+0,1250)+0.6*0.8750) 240.88 psig "laximum Allowable Pressure, New and cold Uo1l\PNC): (SA*EIT)/(D/2.0.6*T) per UG·27 (e) (1) (20000.00*1.00*1.0000)/(144.0000/2tO.6*1.0000) 275.48 psig Actual stress at given pressure and thickness (p* ((0/2tCA) to. 6" (T-CA»)) 1 (E· (T·CA»

(Sact):

(230.00'«(144.0000/2+0.1250).0.6*(0.8750))/(1.00*(0 8750) 19096.57 pai SUJoIlo1ARY OP INTERNAL PRESSURE RESULTS: Required Thickness plus Corrosion Allowance, Trca Actual Thickness as Given in Input Maximum Allowable Working Pressure '·lAWP P Design Pressure as Given in Input

0.9602 1.0000 240.88 2]0.00

in. in. psig psig

HYDROSTATIC TEST PRESSURES ( f·leasurcd at High Point ):

4-4

Example Problem 1- A Simple Drum


eOADE Engineering Software PVE) ite 4.00 Licensee: eOADE Inc.,

Speci ficilliol1s

Local

~Ihite

Lock

PileName Seminar ---------- •• --------------------------page) Shell Analysis: 0101 CYLINDER Item: 1 9:51a Sep 21,2000 Hydro. per UG-99Ibl; 1.3 Hydro. per UG- 991el ; 1.3

~ ~

HAWP • SalS r·1APNC

Hin. f·letal Temp. wlo impact per Pig. UeS-66 IHn. f·letal Temp. at Req'd thk. (per ues 66.1) IHn. f·letal Temp. wlo impact per UG-20 (f)

313.15 358.13

-30 -35 -20

WEIGHT Volume \'/eight Inside I'/eight

and VOLUNE RESULTS, ORIGINI\L THICKNESS: of Shell Component VOLI·lET of Shell Component Wf·1ET Volume of Component VOLIO of I'latel- in Component ImAT

111149.8 31455.4 3973788.0 143497.9

I'/EIGHT Volume I'leight Inside

AND VOLW1E RESULTS, CORRODED THICKNESS: of Shell Component, Corroded VOLNETCA of Shell Component, Corroded 1'11·1ETCA Volume of Component, Corroded VOLIDCA of l'later in Component, Corroded HHATCA

97339.7 27547.1 3987598.0 14]996.6

~leight

The PV Elite Program,


psig

P S1 9 p F

,-

in.~~3

lb. in. "'3

lb.

in.

~ ~

lb. in .• lb.

3

~)

Ie) 1989-2000 by COADE Engineering Software

4-5

z "

o ;;

Pressure Vesscl Design anti Analysis - Scminar Notes

Specifications

Pressure Vessel Design and Analysis Seminar Component Design Problem

D-101 Elliptical Hcad: Starting from the cylinder analysis, design an elliptica' head for the dnnl1.

Notes:

Qnestions:

WhHt is the aspect ratio of the head? What is the required thickness?

What thickness will you usc?

As the pressure increases, the required thickness increases: Linearly

Almost linearly

As the square

DO NOT TURN THE I'AGE UNTIL YOU HAVE COMPLETED YOUR ANALYSIS


4-7

Specifications

Pressure Vcssel Design and Analysis - Scminar Notes

COADE Engineering Software PVElitc 01.00 Licensee: COADE Inc., i.ocal \'Ihite Lock FileName: Seminar -----------------.- ... -.- .. -----------Page 01 Shell Analysis: DIOI ELLIPTICAL ltem: 2 9:51a Sep 21,2000 Input Echo, Component

2,

Design Internal Pressure Temperat\lre for Internal Pressure

Description: 0101 ELLIPTICAl, p

230.00 450.00

Include Hydrostatic Head Components "laterial Specification (Normalized) Allowable Stress At Temperature Allowable Stress At Ambient Curve Name for Chart UCS 66 Joint efficiency for Head Joint Inside

Diameter of Elliptical Head

f-linimum Thickness of Pipe or Plate Corrosion Allowance Aspect Ratio i.ength of Straight Flange Type of Element:

psig F

NO

SA

SA-516 70 20000.00 20000.00

E

1.00

o

144.0000

in.

','

1.0000 0.1250

in. in.

AR

2.0000 2.0000

in.

S

psi psi

0

eA

S1'RTFLG

Elliptical Head

INTERNAL PRESSURE RESULTS, SHELL NUI·IBER 2, Dese.: 0101 ELLIPTICAl. Asr·1E Code, Section VIII, Division 1, 1998, 1\-99 Thickness Due to Internal Pressure (TR): (P' (D.2 l eAl lK)/ (2·S'E-O.2'P) Appendix 1-4 (c) (230.00' (144.000012'0.1250)·1.00)/(2·20000.00'1.00-0.2'230.00) 0.B304 in. r~ax.

All. l'lorking Pressure at Given Thickness (t·IAWP): (2·S·E· (T-CA)) /(K' (Ot2'C/\) +0.2' (T-CA)) per Appendix 1-4 (c) (2.20000.00·1.00·(0.8750))/(1.00·(144.0000+2'0.1250)+0.2'(0.8750)) 242.34 psig

Maximum Allowable Pressure, New and Cold (I·t1\PNC): (2·SA·E·T)/(K*O.0.2·T) per Appendix 1-4 (c) (2·20000.00·1.00·1.0000)/(1.00·144.0000tO.2·1.0000) 277.39 psig Actual stress at given pressure and thickness (Sact): (P' (K· (O.2 l CA).0. 2' (T-CAl) ) / (2·E· (T-CAl) (230.00*(1.00·(144.0000t2·0.1250).0.2·(0.8750»)/(2*1.00-(0.8750» 18981.57 psi Sm~MARY OF INTERNAL PRESSURE RESULTS: Required Thickness plus corrosion Allowance, Trca Actual Thickness as Given in Input Maximum Allowable Working Pressure I>1AWP Design Pressure as Given in Input P

4-8

0.9554 1.0000 242.34 230.00

in.

in. psig psig


"

j

Specificiltions

Pressure Vessel Design and Anillysis - Seminar Notes

COI\DE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local \'Illite Lock FileName: Seminar ----------------------~---------------- Page 5 Shell Analysis: 0101 ELLIPTICAI~ Item: 2 9:51a Sep 21,2000 HYDROSTATIC TEST PRESSURES ( t,leasured at High Point I, 315.04 Hydro. per UG-991b); 1.3 f·1A11P , sals 360.61 Hydro. pel' UG-99(c); 1.3' f·1APNC

p si 9 psig

Nin. I-letal Temp. wlo impact per r~ig. UCS-66 Nin. loIetal Temp. at Req'd thk. (per UCS 66.1) IUn. "'etal Temp. wlo impact per UG-20(f)

F F F

-30 -35 -20

\'IEIGllT and VOLUHE RESULTS, ORIGINAL THICKNESS: VOLI·1ET Volume of Shell Component Wt'lET Weight of Shell Component Inside Volume of Component VOLID Weight of \1atcr in Component \'11'11\1' Inside vol. of 2.00 in. Straight VOLSCA Total Volume for Head I Straight VOLTOT

26074.9 7379.2 390864,4 14114.5 32572.0 423436.4

in.") lb. in. t ' ) lb. in.") in.")

WEIGHT /\NO VOLU1·1E RESULTS, CORRODED THICKNESS: Volume of Shell Component, Corroded VOLI,lETCA weight of Shell Component, CO)"roded I'.'NETCI\ Inside Volume of Component, Corroded VOLIDCA \'Ieight of Water in Component, Corroded WNATCA Inside Vol. of 2.00 in. Straight, Carr. VOLSCA Total volume for Head • Straight Corroded VOLTCA

22815.5 6456.8 392903.7 14188.2 32685,2 425588.9

in. ,*) lb. in. '*3 lb. in." ) in.*')


Example Problem 1- A Simple Dnlln

(c) 1989-2000 by COADE Engineering Software

4-9

Brilll~

fracture - Minimulll Design f\'lclnl Temperalure

Pressure Vessel Design and Analysis - Semillnr Noles

BRITTLE FRACTURE - MINIMUM DESIGN METAL TEMPERATURE

)

New niles in Ihe 1987 Addenda to the Code make il imperalive 10 uuderstand brittle fracture and its relationship to material selection. What is brill/efractllre? It is a sudden and ealastrophie growth of cracks in sleel. How does brittle/roc/lire occllr? The stress at the tip ofa discontinuity in the stcel (crack, inclusion, weld, cte.) reaches a critiefll valuc and the crack begins to propagatc. The total energy ofpropagalion exceeds the energy absorbing capability of the sleel. The crack continues to grow at the speed of sound.

The requirements for brittle fracture are low toughness, high stress, and a discontinuity. Carbon and Low Alloy sleels exhibillow toughness allow lemperalure. AI higher lemperatures the toughness improves dramatically. Charpy impact tests arc one way to measure the fracture toughness of steels. Thc following chart shows a typical variation of impact energy with tcmperature.

00.

1

~I

.----.-------..- - - . - ? I l.\:IlEr sre

~

40 f---

.----...--------/-----.--.----.---.----

W

::1)'---------..- ---.

Iu

20 1----·--.-.-.. -/.----·..-----···----·.. --..··-..·-·

,;

w

101----·.

01 -40

I aWEooor

I -20

--../.-.--------------...--...- ..--.....-

I

o

I 20

I

I

40

fi)

I 00

TerpEi"Olu'e. De;roos F Figure 1-- Typical Variation of Impacl Energy with Tempera/ure

As a function of the grain fineness of the steel, fracture toughness also increases dramatically wilh reduced size of the plale. i\SME has applied Ihesc general principles 10 the impact requiremenls of carbon sleels using a Minimum Design Metal Tcmperature approach: The Minimum Design Metal Tcmperature is the lowest design temperature at which a given steel can be used to construct [\ pressure vessel without impact testing the steel. Figure UCS-66 shows the variation of minimum design temperatUl'e with plate thickness and material. Each of the four curves (A, B, C, D) correspond to several materials.

,

t 4-10

Example Problelll 1- A Simple DnJlll


Brillic FraclUre - Minimulll Design Mel;!1 TemperalUl'C

For a givclll1latcrial, lhe bnsic MDMT is a function only ofthickncss. Howevcf, the MDMT can be rcduccd drnmatically by nOfmalizing or normalizing and tempering many of the carbon stecl materials. (Toughness is rclated to grain size). The MDMT can be red need if the slress in the malerial is lower than thc allowable slress. (Remember, briltle fracture requires a certain energy leveL) Usc figure UCS-

66.1 to calculatc the reduction in Minimum Design Metal Tcmperatufe for a givcn stress level. Most carboll steels lip to I in. are exempt from impact testing down to -20°F. When impact testing is needed, used UG-84 levels.

10

delermine the required energy

Hydrostatic tesling ofa pressure vessel is Ihe best available method for determining maximum tolerable defect size. If a thick pressure vesscl is hydrotested al a prcssure that is

50% greater than the design prcssure, the critical KI (fracturc toughncss) is KIC=sF, whcrc F is a CHICk shape factor Assuming an intemal circular naw of dimcnsion

3,

the maximum KIC immedintely aftcr

successful hydrotesting is KIC

~

1.5*Sm*(2*SQRT(a/pi»

Maximum defect size x at the design prcssure is given by 1.5*Sm*(2*SQRT(a/pi)) = Sm*(2*SQRT(x/pi» or x = 2.25a

Hence, a crack that is discovcrcd after hydrotesting can grow 2.25 times its original size before causing failure. This fact illustrates the importance of hydrotesting and is based on

a hydrostatic temperature that is the same as the lowest opcrating temperature of the vessel.

EXllmplc Problcm 1- A Simple Dnnll

4-11

Pressure Vessel Design and Allalysis - Seminar Notes

Brillic Fracture· Minimum Design Metal Temperature

, }

I

)

!

140

I 120

!

I

100

g.

'0

,~

60

! :;

>!

II I I

'0 20

.~!

E

~

C ~

o

~I J

- '0

I

/ L

;-

I

0.39<1

V ./

1---

....

cj.--

~

Vr-

__

D

_

~

_

f-b-:::~

~

V r

/

t=

. ~--

__

-----

V

/ /V

.r.V_r-

~

V

iV V

-65 -60 _ 60

/

i L/

- 20

V

/

I / I

.§.

~

A/

......

_

t--....

'1 "T' "q"IT

--

--

PK

•

2

,

3

6

Nomlll.1 Thlckllo».IIl. Illmll0d 10 .. Ill. for Woldod Consltuclton)

Gonenl Not,s and Nolos tollV'N on nell1 pago

FIG. UCS·66 IMPACT TEST EXEMPTION CURVES [SEE NOTES (1) ANO (2ll [SEE UCS·66(a))

Figure 2--1998 Section VIII, Division 1

.' 4-12

Example Problem

I~

A Simple Drum


and

Annlysis - Seminnr

I3rillle Fracture - Minimum Design Mewl TcmpcrallJrc

Noles

FIG. UCS-66 (CONT'OJ

or

GENERAL NOTES ON ASSIGNMENT MATERIALS TO CURVES: (a) Curve A applies to: (1) all carbon and all low alloy steel plates, structural shapes, and bars not listed In Curves B, C, and 0 below; (2) SA·21b Grades WCB and WCC if normalized and tempered or water-quenched and tempered; SA·217 Grade WCb if normalized and tempered or waler-quenched and tempered. (bl Curve 0 applies to: (1) SA-21b Grade WCA if normalized and tenlpered or water· quenched and tempered 511.·216 Grades WCD and WCC for thicknesses not exceeding 2 In., If produced to fine grain practice and water-quenched and tempered 511.-217 Grade WC9 if normalized and tempered 511.-205 Grades A and 0 511.-414 Grade A SA-SIS Grade 60 511.-516 Grades b5 and 70 if not normalized 511.-612 If 1I0t normalized SA-b62 Grade B if not normalized; (Zl except fOr cast steels, alt materials of Curve A if produced to fine grain practice and nOrmalized which are not listed In Curves C and 0 below; (3l aU pipe, fittings, forglrl9s and tubing not listed fOr Curves C and 0 below; (4) parts permitted under UG-ll shall be Included In Curve 8 even when fabricated from plate that otherwise would be assigned to a different curve. (c) Curve C (1) SA·182 Grades 21 and 22 If normalized and tempered 511.-302 Grades C and 0 511.-336 F21 and F22 if normalized and tempered 511.-307 Grades 21 and 22 if normalized and tempered 511.-516 Grades 55 and 60 if not nOrmalized 511.·533 Grades Band C SA-6b2 Grade A; (2l all material of Curve B If produced to fine grain practice and normalized and not listed for Curve 0 below. (d) Curve 0 511.·203 511.-500 Grade 1 511.·516 If normalized 511.·524 Classes 1 and 2 511.-537 Classes 1, 2, and 3 SA-bl2 If normalized SA-bb2 If nOrmalized SA-nO Grade A (e) For bolting and nuts, the following Impact lest exemption temperature shaH apply: Bolting

Spec. No. SA·193

Grade

85

SA-193

07 (2 1/ 2 In. dla. and under) (Over 2 1/2 In. to 7 In., IncU 87M

511.·193

016

511.·193

Impact Test Exemptlon Temperature, -F

-20 -55

-40 -55 -20 -20

SA-307

o

511.·320

L7, L7A, L7M, L43

SA-325

1, 2

-20

00

.20 -20 .10

511.·354 SA·354 511.·449 51\-540

oe

023n4

Impact tested

o

GMeral Notes ilnd !loin conl/nue on ned

Pd~

Figure 3--General notes on assignment of materials to cUlYes

EXntllplc Problcm 1- A Simplc Dnllll

4-13

Pressure Vessel Design fmd Analysis - Seminar Notes

Briltle fracture· Minimulll Design Metal Temperature

-~

I

1001<

I

I

I

I

..

--.~-_.~

I

I

I

I

I

I

I

I

I

.g ~

cc

~

.~

~

E 0.80

.1!
.2 ~ ~

;; uc 0.80, ~

E 0

z ~

~

Ul

<; I

-'". ~

0.40 ' 0.35

•

~

.2 ;; cc

0.20 .

0.00

CCC(

o

r

(o? ({(CCA c/( ..1//(tl/!~''I/V/4/v///v/c (v(( ( r ( / / l ( ( ( r ( ( ( 4 ( ( / 6

20

40

60 80 'F ISee UCS·66(bjJ

100

120

140

.,i

Nomenclaturo (Note references to General Notes of Fig. UCS·66.2.) tr :< required thickness of the component under consideralion In the corroded condition for ell applicable loadings IGeneral Note (21J. based on tho applicable joint efficiency E IGeneral Note (311. in. t n '" nominal thickness of the component under consideration before corrosion allowance Is deducted. in. c • corrosion allowance. in. E' as defined in General Note (3). AHcrnalive Ratio S· E* divided by the product ollhe maximum allowable stress value from Table UCS·23 times E. whore S· is tho applied genoral primary membrane tensile stress and E and E· are as definod in General N01e(3).

FIG. UCS-66.1 REDUCTION IN MINIMUM DESIGN METAL TEMPERATURE WITHOUT IMPACT TESTING

Figure 4--Reduction in Minimum Design Me/ai Tempera/ure

1

.'

4-14


Prcs~urc

Vessel Design and Analysis - Seminar NOles

Brillle fmClure - Minimum Design Melal Temperature

ESlablish nominallhicknt'lsscs (General Note (111 of welded par1s, nonwelded parts, and allachments under consideration both before and after corrosion allowance is deducted (In and tn - c, respectively), and other pertinent data applicable to tho nominal thicknesses such as: All i1pplicable loadings (General Nolo (2)) and coincident minimum design melal temperature (MDMTI Materials of construction E .. joint etriciency IGoneral Nolo (3)J .. nominal noncorrodcd thickness IGeneral Note (1)1. in. " • required thickness in corroded condition for all applicable loadings IGeneral Noto (2)1, " based on the applicable joint eU;ciency IGenaral NOle f311, in. Applicable curve(s) 01 Fig, UCS·66 c .. corrosion allowance, in.

Stop 1

l Select MDMT from Fig, UCS·66 (Genoral Note (4)1 ror each nominal noncorrodod governi~g thickness lGeneral Note (5)).

StOll 2

r------ -'- ------, 'I

StOll 3

I

Determine Ratio:

t,P

tn-c IGeneral Noles 13>, f6), (7), and (8))

II

J

~====== [~~=~~~~I

I I I

Stop 4

Using Ratio from Step 3 to cnler ordinate of Fig. UCS-66.1, determine reduction in Slep 2 MDMT IGenr.rel NolO (9)1.

I I

L-------I-------~ r------I I

Stop 5

-------,

Delormine adjusted MOMT for governing Ihickness under consideration.

: I

L-----~-l------~-~

Repest for all governing thicknesses [General Note 1511 and take warmestvaluo as the lowest allowable MOMT to be marked on nameplate for Iho zono under consideration {General Note 11011. Sec UG·116. See UG·99 Ih) for coldest recommended met'll temperature during hydrostatic testlGonoral Note (611. Sec UG·100(c) for coldost motal temperature permitted during pneumatic tcstlGencral Notc 1611.

Stop 6

legend

I

I

Requirement

,r----l.

------

Optional

General Not6S foflow on noxt pago

FIG. UCS-66.2 DIAGRAM OF UCS-66 RULES FOR DETERMINING LOWEST MINIMUM DESIGN METAL TEMPERATURE IMDMT) WITHOUT IMPACT TESTING

Figure 5··Diagram of UGS-66 RlIles for Determining Lowest Minimllm Design Me/al Tempera/lire (MDMT) Wi/holl/lmpac/ Testing

Example Problem 1- A Simple Dnml

4-15

Brittle Fracturc - Minimulll Design Mclal Tcmperallll"c

Prcssme Vcsscl Dcsign and Analysis - Scminar Notes

"

i Fig. UCS·66.2

1998 SECfION VIU -

DIVISION 1

FIG. UCS-66.2 (CONT'O) GENERAL tWTES: 11) For pipe where a mill undertolerance Is allowed by the material specifIcation, the thickness after mill undertolerance has been deducted shall Ix> taken as the noncorroded nominal thickness tit for determination of the MDt.n to be stamped on the nanwplate. LIkewise, for formed heads, the minimum specified thickness after forming shall be used as I". (2) loadings, Including those listed In UG·22, which result in general primary membrane tensile stress at the coincl~nt MOMT. ()) E is the joint efficiency (Table UW·l2) used In the calculation of tf ; e has a value equal to E exceplthat E" shall not be less than 0.60. For castings, use quality 'ador or Joint efficiency EYAlichevtr governs design. (4) The construction of Fig. UCS'66Is such that the MDMT so selected Is considered to occur coincidentally wilh an applied qeoeral primary membrane tensile stress at the maximum allowable .stress value In tension from Table IA of Section II Part 0, Tabular values for Fig. UCS-M are shown In Table UCS·6b. (5) See UCS-bMa)(I), (2), and 0) for definitions of governing thickness. (6) If the basis for calculated test pressure Is greater than the design pressure (UG·99(c) test], a Ratio based on the I, determined from the basis for calculated test pressure and associated appropriate value of '" - c shall be used to determine the recommended coldest metal temperature during hydrost.1tlc tut and the coldest metal temprrature permitled during the pneumatic test. See UG·Q9(h) and UG-100{c). (7) Alternath-ely, a Ratio of S' ~ divided by the produ<.t of the maximum allowable stress value in tension from Table IA of Section 11 Part o times E may be used, where ~ is the applied qeneral primary membrane tensile stress and E and e are as defined in General Note (31. (8) For UCS-6b(b)(IHb) and (1)(2), a ratio of the maximum design pressure at the MOMT to the maKimum allowable pressure (MAP) at the MOMT shall be used. The MAP 15 defined as the hignest permissible pressute as determined by the design formulas for a component using the nominal thickness less cOrloslon allowance and the maKlmum allowable stress value from the Table lA of Section II, Part 0 at the MDMT. For ferrille Sleel flanges defined In UCS-6b(c), the flange rating at the warmer of the MOMT or 100"F may be used as the MAP. (9) For reductions In MOMT up to and Including 40"F, the reduction can be determined by; reduction in MDMT t= 11 - Ratlo)lOO"F. no) A colder MDMT may be obtained by selective use of impact tested materials as appropriate lo the llCed (see UG·8tll. See also UCS-6B(c).

Figure 6..General Noles (or Fig. UGS-66.2

}

, ,, 4·16


Pressure Vessel Design ;md Analysis - Seminar Notes

13rillle fracllIrc - Minimum Design Metal Temperature

1998 SECTION VIn -

Fig. UG·84.1

OJVISION 1

! I

0.39'1 In.

I

50

i I

I I

40

, 30

'0

1"

!

/'

,/"

I

I

./

I

,

"'u:>

65 ksi

r

•E

£ •l'

Minimum spe-cilied ylald J1ranglh

I

I

:!

•

I

'0

!

./

I

15

I

'0

!

~

I

i

./

l:::::::

0

55 ksi

/

60 ks,

~

45 ksl

~ t.---

<

38 ksi

,

r

o

/""

/

: 1.0

'_0

Mallimum Nomlnel Thickness of

;> Mal~rlel

3.0

or Weld. In.

GENERAL NOTES:

Cal Inlarpolallon batwean yield stHlngths shown i, permlltod. Cbl The minimuM Implcl enolgy for ono spacimon shall nol ba less lhan '1J3 of lhe average energy requlled fOI th,oo spoclmans. (cl Malarlals produced and IMpacl tested In accordanco with $A·320. SA·333. SA·334. SA·350, SA·351, SA·,nO and SA·165 do not havo 10 satisfy lhose onergy valuel. Thoy ara acceptable for use al minimum design motal tomperalure not colder than lha lest temperatura when lho anergy values loqulred by tho applicablo spflciflcatlon .ra sllllsliod, ldl For matorlalJ having I spaclfied mInimum tllnsllo slrength of 95 ksl or more. soe UG·a4IcU'llbl.

FlG_ UG-84_1 CHARPY V-NOTCH IMPACT TEST REQUIREMENTS fOR fULL SIZE SPECIMENS fOR CARBON ANO LOW ALLOY STEELS, IIAVING A SPECIFIED MINIMUM TENSILE STRENGTH Of LESS TIIAN 95 ks;, LISTED IN TABLE UCS-23

Figure 7--Charpy V-Notch Impact Test Requirements

Example Problem 1- A Simple DnUll

_

4-17

3

2

0

-5;;-

>

I

~

c§"

0

:r

;;-

"C

~

~

'"3

ov.

v.

;;

0

z

~

g

:;:;.

'A

.:;:

c..

>

~.

0

~

v.

V.

~

<

Co

~

;f .,.,.

Prcssure Vessel Dcsign and AUillysis - Scminar Noles

nrillic Fracture - Minimulll Design Metal Temperature

Pressurc Vessel Design
Minimum Design Met
MOMT of Cylinder MOMT of Ellipse What is Ihe MOMT if the sleel is nol normalized?

MOMT of Cylinder -

MOMT of Ellipse-

DO NOT TURN THE I'AGE UNTIL YOU HAVE COMPLETED YOUR ANALYSIS

EXill1lplc Problem 1- A Simple Drum

4-t9

OrilLl!.: FrnClure - Minimulll Design Ivklal Temperature

Pressure Vessel Dcsign "lid Analysis. Scminar Notes

Answers:

I-ill. SI\-516,70, Ilormolized, MOMT

= -30

F (-34 C) Curve 0

I-ill. SI\-516,70, lIollllormolized, MOMT = 30 F Curve 13 bllt per UG20, this COil be -20 F (-29 C)

4-20


Extern;}1 Pressure on :,hdls and hcnds

Pressure Vessel Design and Analysis - Scminar Noles

EXTERNAL PRESSURE ON SHELLS AND HEADS Pressure vessels under extern'll pressure buckle r'lther than deforming elastically or yieldmg. A cylinder deforms into lobed shapes as shown in Figure 9.

Buckling Mode

0 f

I

\

Ch~n"cleriIIJcl

k· Number of lobes (k is the number 01 lull sine waves eround (he periphery)

coll~pse

I

J

-'

0 .,

0 (No elastic buckling OCCUrl. only side'wile dilpJacemenlof undislorloo drciel

- .,'

CJ 0

, ',

,

--'

I

Pc" Crilical prenure lhe elallic

buckling or

,

\

,

.... ,

2

_ .. I

3

,~

--

4

3Et

8Et

15EI

7

7

--;>

Fig. 8.2. Buckling or CoU~ple of Cylindriaal Rin.!: Under EXlernal

PrCHU[C

~,

Showing Various Modes

Figure B--Buckling or Collapse of Cylindrical Ring

The buckling of a cylinder can be predicted based on the strain due to external pressure. A

given thickness, length, and diameter will buckle at a predictable strain. With an appropriate factor of safety, the Code uses this strain to calculatc the allowable

stress under external pressure. The Code calls the critical strain factor 'N and the allowable stress 'B'.

For heads or for allowable compressive stresses in shells, A is a function only of the thickness and diameter. The geometry chIlli in the Codc is derived from the theorctical buckling strain. Above a certain length/diamcter ratio (typically 10), the result of the buckling calculation is independent of the Icngth of the cylinder. Thus, decreasing the free length of a cylinder

docs not begin to increase the allowable pressure until the free rength is less than 10 times the diameter.

Example Problem 1-- A Simple Dnnl1

4-21

Pressure Vcssel Design alUl Analysis - Scminar Noles

External Pressure on shells and heads

0" I r 0" 1 f

~

4

____

5

•

~~ y'r': •

0"

o·

•

•

1/

M

/

o

V

I--'

10

I"

I,

V V

~

~

I-'

0~/tl'8

-.

~

o·

V

Do /t • 6

~

,l

V-

N

v Il'

~

~ ~

00

1/ •

•

IS

I I I

f-

D: ,,'.,), D:/,'·,l D:"'. JO I I

~

/

V :/

/

V

/

L../

/

" /

V

I

V

V V'"

"V

"

"

~

~

V V

l-

60

/

/

/

.100

l.-

V V

V

I/

V

V'"

V ",V

V

V-

~ ~

§

~

L-

1

V

f

V

D~,.I. ,&,

L.-

D~"I:,~

~~~/

V /

~,,'. ,:" ~

N

,

0~1I1. ,~

C!

<

M

00lt o l25

~ ~

u

"

•

V

V f

'l: 0.

0

~

V V

g

a "«

•

o~lIl.sb 0 0 1/

<

III

,

0 9 11 -50 Oolf

M

N

°0/r.40

I-

•

001

~

~~ ~

C! ~

0. ~

V-

/'

~

i" • M

1/

001

00

~ ~ ~. ~ ~

:;l-

N

~

0 01

0. •

II!

M

C!

M

~

"

owoo-.

N": ..:

..:

N

..:

~

°0 /7 .. maW9!O 0PISlno ... lll6u81 Figure g··Geometric Chart for Factor A

4·22

Examplc Problcm J- A Simplc Drum

Extern
Pressure Vessel Dcsign and Analysis - Seminar Noles

·,,• OO~~'~Z~£~f~y ~ 0/ ~ o~ m <

~.,o \

):

::...t

1\

pO

<

1\:,A--,\*

ol

V

/

7

,

V

1\*

/

,o·.'~

V

..-

V

V

V

1/

V

V

V

...-

V

l/

V V/

V

V

V

yl/v V V

.-

~
Y

/'

.-

...-

/...-

..- V

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...-..-

V

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·,,

o'-j-- 'd'

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m <

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om~'N°2RO~~

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-:

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q

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°0 / 1 .. JalawgrO BPIslnO + 416u'll

Figure 10--Geometric ClJart for Components under External or Compressive Loadings

Since B is the stress associated with the strain A for a given material, the materials charts in the external pressure section (ASME, 1992, Section II, Parl D, Subpart 3) are really just stress strain diagrams for a given material, adjusted to reflect a specific factor of safety against buckling. The Code has also provided a tabular form for the charts in Subpart 3.

Example Problcm 1- A Simple Drum

4-23

External Pressure Ull shells aud heads

Pressure Vessel Design and Analysi:, - Seminar Nute:,

The external pressure charts have a straight section which corresponds to the clastic portion of the stress strain curve. Inlhis straight section, B is always equal to AE/2 where A is lhc strain and E is Young's modulus for the material.

1-

1 I

I~)',

.",

L--

V.

t= f= -

-

1.".~.'O· t"

.

"""

3

-

n,,, 111 1

rmi'"

4 5 6 1 III

f'

II/II

71///

IiSfVI 3

..",

--

,""00 ·

/ ~~I

I--

.-

.

, 11(;

. ..

-

.

,&,(IoXl

, ,

.~~

'-~

OCO:u

-

""'

-_. --

•.1.")

·

·1- 1-- --/l!Jf- .

1-190 .. 10' I" 110. 10(

l-

V

:e.- . -

'~J»O

J,,':t!.Ou' J,!

GENERAL NOTE; See Tablo CS·2101 tGbulil ... 31 Ullt,

-

I,(m .~o

w ...

.

· .

_,CY.Jl

·

1,1'"

-

- _. 4 5 6 18'

"" ,,,crOll "

3

456/111

34~'719

"

""

, ""

FIG. CS-2 WART FOR DETERMINING SUEll TlUCI
Figure 11·· Typical Materials Chart for Factor B

Up untillhe 1992 edition of the Code, external pressure charts (and Ihe tabular form) were in Appendix 5 of Section VllI, Division I, In 1992 these were moved to Part D of Seelion II (Malerials). They are found in Subparl 3 of Part D. At the same time Ihe names of the charts were changed. The following table shows the new names, the old names, and all abbreviated title for each chart: New Name

Table-Garboll Steel Materials Old Name Title

CS-I CS-2 CS-3 CS-4 CS-5 CS-6

UCS-28.1 UCS-28.2 UCS-28,3 UCS-28.4 UCS-28.5 UCS-28.6

New Name

Table...J4eat-TI'eated Materials Old Name Title

HT-1 HT-2

UHT-28.1 UHT-28.2

New Name

4·24

Carbon and Low Alloy, Sy<30000 Carbon and Low Alloy, Sy>30000 Carbon and Low Alloy, Sy>38000. SA-537 SA·508, SA-533, SA-54 I SA-562 or SA-620

SA-517 and SA-592 A, E, and f SA-508 Cl. 4a, SA-543,13,C

Table-8lainless Steel (High Alloy) Materials Old Name Title


External Pressure

Pressure Vessel Design ,1Ild Analysis - Scminnr Notes

HA-I HA-2 IIA-3 HAA HA-5

UIIA-28.1 UIIA-28.2 UHA-28.3 UHA-28.4 UIIA-28.5

011

shells ,11ld heads

Type 304 Type 3 16, 321, 347, 309, 310, 430B Type 304L Type 316L, 317L Alloy S31500

Table-l'/{)n Ferrons Materials New Name

Old Name

Title

NFA-I NFA-2 NFA-3 NFA-4 NFA-5 NFA-6 NFA-7 NFA-8 NFA-9 NFA-IO NFA-II NFA-12 NFA-13 NFA-14 NFC-I NFC-2 NFC-3 NFC-4 NFC-5 NFC-6 NFN-I NFN-2 NFN-3 NFN-4 NFN-5 NFN-6 NFN-7 NFN-8 NFN-9 NFN-IO NFN-II NFN-12 NFN-13 NFN-14 NFN-15 NFN-16 NFN-17 NFN-18 NFN-19

UNF-28.2 UNF-28.J UNF-28.4 UNF-28.5 UNF-28.IJ UNF-28.14 UNF-28.17 UNF-28.18 UNF-28.19 UNF-28.20 UNF-28.2J UNF-28.JO UNF-28.J I UNF-28.J2 UNF-28.9 UNF-28.10 UNF-28.11 UNF-28.12 UNF-28.4J UNF-28.48 UNF-28.1 UNF-28.6 UNF-28.7 UNF-28.8 UNF-28.15 UNF-28.24 UNF-28.25 UNF-28.27 UNF-28.29 UNF-28.JJ UNF-28.34 UNF-28.J6 UNF-28.J7 UNF-28.J8 UNF-28.J9 UNF-28.40 UNF-28.44 UNF-28.45 UNF-28.46

ALJ003,0 and HI 12 ALJ003, HJ4 ALJ004, 0 and H 112 ALJ004, 1134 AL5154,0 and H 112 C61400 (Aluminum Bronze) ALl060,0 AL5052,0 and HI 12 AL5086, 0 and H 112 AL5456,0 AL5083, 0 and H 112 AL6061, T6, T65 I, T65 10 and T6511 AL6061, T4, T451, T4510 and T4511 AL5454, 0 and H J 12 Annealed Copper Copper-Silicon A and C Annealed 90-10 Copper Nickel Annealed 70-30 Copper Nickel Welded Copper Iron Alloy Tube SB-75 and SB-Ill Copper Tube Low Carbon Niekcl Ni N i Cu Alloy Annealed Ni Cr Fe Ni Mo Alloy B Ni Mo Cr Fe Ni Mo Cr Fe Cu Ni Fe Cr Alloy 800 . Ni Fe Cr Alloy 800H Ni Moly Chrome Alloy N I0276 Ni Cr Fe Mo Cu Alloys G and G-2 Cr Ni Fe Mo Cu Co, SB-462, 463, etc. Ni Fe Cr Si Alloy 330 Ni Cr Mo Grade C-4 Ni Mo Alloy X Ni Mo Alloy B-2 Ni Cr Mo Co N06625 (Alloy 625) Ni Mo Cr Fe Cn (Grade G-3) Ni Mo Cr Pe Cu (Grade G-3, >3/4)

Example Problem 1- A Simple DnllTI

4-25

E.\lt:rnal Pressure

011

shells alld hcruls

NFN-20 NI-T-I NFT-2 NFT-3 NFZ-I NFZ-2

Pressure Vessel Design and Analysis· St:lllinar Notes

UNF-28.47 UNF-28.22 UNF-28.28 UNF-28.42 UNF-28.35 UNF-28.41

1-'

Work Hanlcncd Nickel Unalloyed Titaniulll, Grnde I Unalloyed Titanium, Grade 2 Til"nium, Gradc I Zirconium, Alloy 702 Zirconium, Alloy 705 t

Moment axis of ring ~~~

L -.;..- L -

L

h/3

--+---

00 I

-+-L~-L

I

I

1113 h

a

depth 01 hcad-!4, • •

Ih--

L --,J.,-

..j

L-i-l I I

L

,r

Do

,

Figure 12-- Typical Geometry showing variables used for External Pressure Analysis

4-26

EXClIllJllc Problem 1- A Simple Drum

Pressure Vessel Design rllld Annlysis - Seminar Noles

Stiffening

STIFFENING /\ section ora vcsscl held ill a circular shape by a stifTening ring will not buckle into lobes, and will prevent the sections ncar it on either side frolll buckling into lobes. for a given length of cylinder, the Code requires that the stiffener have a known moment of incrtia, to cnsure that it will hold the section round for the calculated external load. The reinforcement required in a stiffener increases as a function of the length between stiffeners. However, increasing a stiffener size does not change the maximum length between stiffeners at all. When calculating available stiffener moment of inertia, the Code allows you to decide whether or not to take the shell into account in your calculation. Thc Codc allows you to include a width of shell in the calcutation cqual to 1.1 * SQRT(Do*t). This is the dccay length of the cylinder.

1

0.' 0.0

Decay Length:!: 1,56·SQRT{rl)

0.7 0.6

0.' 0.' 0.' 0.2

_::~H

~~

I

==+---;'

-0.2 -0.3 ~O.4

-O.!l -0.6

-0.7 -0.6 -0 • ~,

I

I

I

'=...",/

I

I

04.00

2.00

0.00

o

(Xp(eX)

x

COS(BX)

v

I

I

6,00

EXP(BX)'COS(8X)

Figure 13--Decay Length for Cylinder Local Bending The required value of the moment of inertia will be higher when you include the shell in your actual moment ofincrtia.ln cffect the Code credits something for the shell when you chose not to use it in your calculation.

Example Problem 1- A Simple DnllJl

4-27


Stifrening

r1rY' i

T

j

1_

,"

l{J

i ~/

-l,

L

,

'l'

I 1--11

.

I~

- 21

Ibl

I NOIII (II .nd (:Ill

--l

T r 1, J L

l~ i L L

L

'lc - tI

Irlol.1311

Ie - 21 [Notu OJ 3M l:lll

'Cdl

'.1

INoI. (3))

'"

INotll (3)\

'NOTES: C11 VlII.n 1"'- conl.to-cv'inder Of' thl knuckl"l(~eVnnderJunction I. not, line ollUpport. tho nomln,l thlcknell of Ihe cone. knuckle. 01 1000ieonlc.t uetlon lh.1I not be 111I1han Ih. minimum reQulre-d thlckneu of Ih. adiKllnl evllndlle.1 ,hill. 121 Calculetlon, ch,U be made u,lng tho dllmeter end cOHupandlrog thlcknen of nch ,e-ctlon ..,llh dlmen.lon L III 'hown. IJI Vlhln \h. con..to-" pt'ovldloJ In llGCOrd,nCl with 1-8.

Figure 14--Geometry Showing Design Length for External Pressure for Cylinders

What do you need to know to analyze cylinders and heads for external pressure? External Pressure Chart The tables of allowable tensile stress fol' materials also give the name of the external pressure c1wrt to use for the analysis. Many external pressure charts have both lower and upper limits on temperature. If your design temperature is below the lower limit, use values chosen at thc lower limit. If your temperature is above the upper limit thc component should not be designcd for vacuum conditions.

Diameter Thc diamclcr for external pressure considcrations is ;]Iways the ollis ide diamcter.

4-28

Example Problem 1-/\ Simple Drum

Pressurc Vcssel Design and Analysis - Seminar Notes

SlilTcning

Actual (or assumed) Thickness The strain factor is based on the ratio of diameter to thickness. Therefore to find the required thickness for a given pressure, you have to guess a thickness, calculate the allowable pressure associated with that thickness, and keep modifying it until the allowable pressure is equal to the desired pressure.

Design Length for the Vessel or Vessel Segment The design length orthe section is typically the length orthe vessel pills one third the depth of the heads or, alternately, the distance between stinening rings. For a vessel with 2 elliptical heads nnd no intermediate stiffeners, the design length is the tangent length plus the diameter/G. for a vessel with 2 spherical hcads and no intermediate stiffencrs, the design length is the tangent length plus the diametcr!3. For a vessel with 2 flanged and dishcd heads and no intermediate stiffeners, the design length is the tangent length pills the diameterl9.

Width and Thickness of Reinforcing Rings. Most external pressure stiffcning rings arc a simple flat plate cut (or rolled the hard way) into the shnpe of a donut. Calculate the momcnt of inertia of the flat plate by combining it with the shell section. You can also combine the moment of inertia of a beam section with the shell contribution using thc parallcl axis transfer theorem.

Example Problem

l~

A Simple Drum

4-29

w

...

~

-c;;-

~

>

I

;;

;;-

"

n

9-

z

~

cr.

g

""

-< ,..

~

>

"-

§

-0

<:-

;:;

"

C ~.

n

n

g

<

n

:~

~

n

-c;;-

§

~

rn

o


Stiffening


0-101 Shell and Heads - Extemal Pressure:

Using the

gCOIllCtlY

defined for internal pressurc, eheck the shell and heads for full vac-

UUIll.

Notes:

Questions: \Vhat is the rcquircd thickness? Whatthiekness will you use? What is the M.A. w.p for the Shell? What is the M.A. w.p. for the Head? Why are the M.A.W.p for the shell and the head so different?



4-31

Stiffening

Pressurc Vessel Design ,111(\ Analysis - Seminar Notcs

CQADE Engineering Software PVEI i te 4 . 00 Li censee: COADE Inc., Loca 1 \'/hi te Lock F'ileName : Seminar --------------------------------------P
Description: 0101 EXTERNAL

3.

Design External Pressure Temperature for External Pressure External Pressure Chart Name

PF.:XT

15.00 450.00 CS-2

NO

Include Hydrostatic lIead Components I·latedal Specification (Normalized) Allow
SA

SI\-516 70 20000.00 20000.00

E

.00

L

268.0000 244.0000 144.0000

in. in. in.

1.0000 o . 1250

in. in.

S

psi psi

D

Design I~ength of Section I,ength of Cylinder for Volume Cales. Inside Diameter of Cylindrical Shell

CYLLEN D

l-1inimum Thickness of Pipe or Plate Corrosion Allowance

CA

Type o[ Element:

Cylindrical Shell

WEIGHT Volume weight Inside

psig P

T

~Ieight

and VOLUME RESULTS, ORIGINAL THICKNESS: of Shell Component VOLI·1BT of Shell Component Wl·1ET Volume of Component VOLID of Water in Component \'111AT

111149.8 31455.4 3973'188.0 143497.9

in .• , 3 lb. in. "3 lb.

WEIGHT Volume Height Inside Weight

AND VOLUME RESULTS, CORRODED THICKNESS: of Shell Component, Corroded VOWETCA of Shell Component, Corroded Wl·1ETCA Volume of Component, Corroded VOLIDCA of l1ater in Component, Corroded WWATCA

97339.7 27547.1 3987598.0 143996.6

in." 3 lb. in. "3 lb.

EXTERNAL PRESSURE RESULTS, SHELL NUI·18BR 3, Desc.; 0101 EXTERNAL ASf.lB Code. Section VIII, Division 1, 1998, A-99 External Pressure Chart CS-2 Elastic "lodulus for Material

at

450.00 27500000.00

F psi

Results for t·lax. Allowable External Pl"essut·e (Emawp): Corroded Thickness of Shell TCA 0.8750 in. Outside Diameter of Shell 00 146.0000 in. Design Length of Cylinder or Cone SLEN 268.0000 in. Diameter / Thickness Ratio (0/1') 166.8571 Length / Diameter Ratio LD 1.8356 Geometl·y Factor, A f(DT,LD) l\ 0.0003286 f·laterials Factor, B, f(A, Chart) B 4518.0068 psi Maximum Allowable Working Pressure 36.10 psig Ef.lA\'IP .. (4'B)/(3'(D/T)1 """ (4 ' 4518.0068 1/( 3 ' 166.8571 ) .. 36.1028 Results fOl" Reqd Thickness for Ext.

4-32

Pressure (Tca)


Pressurc Vcssel Design and Analysis - SCml1l
COADE Engineering PVElite

4.00

Stiffcning

Soft\~are

Licensee; COADE Inc.,

Local I'lhite Lock

rileName Seminar ------------------------ -------------Page 7 Shell Analysis 0101 EXTERNAL Item; 3 9:51a Sep 21,2000 Corroded Thickness of Shell TCA 0,6158 Outside Diameter of Shell 00 146.0000 Design Length of cylinder or Cone SLEN 268.0000 Diameter / Thickness Ratio (0/1') 237.0924 Length / Diameter Ratio I,D 1.8356 Geometry F'actor, A f(DT,LD) A 0.0001940 r·laterials F'actor, B, f(A, Chart) B 2667.3970 r·laximum AllOl~able \'larking Pressure 15.00 EI1l\\'/P = (4*S)/D*(O/T)) = ( 'I * 2667.3970 )/( 3 * 237.0924 )

(

in. in. in.

psi psig = 15.0006

Results for J·laximum Length Bet"'Jeen Stiffeners (Slenl Corroded Thickness of Shell TeA 0.8750 in. Outside Diameter of Shell 00 146.0000 in. Design Length of Cylinder or Cone SLEN 6'1'1.9812 in. Diameter I Thickness Ratio (0/1') 166.8571 Length I Diameter Ratio 4.4177 LD Geometry F'actor, A f(DT,LD) 0.0001365 A Materials F'actor, B, f(A, Chart) B 1877.3041 psi r·laximuffi Allowable l'larking Pressure 15.00 psig EHA1'IP", (4*8)/(3*(0/1')) '" (4·1877.304) l/( 3 · 166.8571 ) = 15.0013 SUI·1I·1ARY of EXTERNAL PRESSURE RESULTS: Allowable Pressure at Corroded thickness Required Pressure as entered by User Required Thickness including Corrosion all. Actual Thickness as entered by User r·laximum Length for Thickness and Pressure Actual Length as entered by User The PV Elite Program,


(c)

36.10 15.00 0.7408 1.0000 644.981 268.00

psig psig in. in. in. in.

1989-2000 by COADE Engineering Software

4-33


Stilfenlng

e01\DE Engineering Sofl""'are PV~:1 i t.e 4. 00 Li censee: COAOE r nc ., Loca I \'/hi te Lock FileName: Seminar ------------------.--.--- •• -----------rage 8 Shell 1\rwlysis : 0101 EXT HE1\D Item; 4 9:51a Sep 21,2000 Input Echo, Component

4.

Design Ext.ernal Pressure Temperat.ure for f:o:xtel'nal Pressure External Pressure Chart. Name

Description: 0101 EXT HEAD PEXT

15.00 450.00 CS-2

Include Hydrostatic !lead Components !·latel'ial Specification (Normalized) 1\llowable Stress 1\t Temperature Allowable Stress At AMbient Curve Name [or Chart UCS 66 Joint efficiency [or Head Joint Inside

Diameter of Elliptical Head

r·linimum Thickness of Pipe or Plate Corrosion Allowance lI.spect Ratio Length o[ Straight Flal:ge Type of Element.:

psig

,.

NO

SA

SA-516 70 20000.00 20000.00

r.

1.00

D

144.0000

in.

T CA

1.0000 0.1250

in. in.

AR

2.0000 2.0000

in.

S

psi psi

0

STRTFLG

Elliptical Head

WEIGHT and VOLW1E RESULTS, ORIGINAL THICKNESS: Volume of Shell Component VOLt·1ET \'leight of Shell Component wt·IET Inside Volume of Component vaLID weight of I-Iater in Component WI-I1\T Inside Vol. of 2.00 in. Straight VOLSCA VOLTaT Total Volume for Ilcad ... Straight

26074.9 7379.2 390864.4 14114.5 32572.0 423436.4

WEIGHT AND VOLW~E RESULTS, CORRODED TIIICKNESS: VQUoIETCA Volume of Shell Component, Corroded Weight of Shell component, WI·IETCA Corroded VOLIDC1\ Inside Volume of Component, Corroded Weight of NateI' in Component, Corroded I'lWATCA Inside Vol. ot 2.00 in. Straight, Carr. VOLSCA Total Volume for Head * Straight Corroded VaLTCA

22815.5 6456.8 392903.7 14188.2 )2685.2 425588.9

in. H3

lb. in. * 03

lb. in. * 03 in.' 03

in. "3

lb. in.' 03

lb. in ... ) in.' 03

EXTERNAL PRESSURE RESULTS, SHELL NUt·mER 4, Desc.; 0101 EXT HEAD Code, Section VIII, Division 1, 1998, A-99

ASI~E

External Pressure Chart CS-2 Elastic 1·lodulu5 for 1·laterial

at

450.00 27500000.00

Results for 11ax. Allowable External Pressure (Emawp) TCA 0.8750 Corroded Thickness of S~ell Outside Diameter o[ Shell OD 146.0000 (O/T) 166.8571 Diameter I Thickness Ratio Geometry Factor, A f{DT,LD) A 0.0008324 r·laterials Factor, B, f (A, Chart) 10051.5:;n3 r·laximum 1\llo·....able I-Iorking Pressure 66.93

"

4-34

F

psi

in. in.

psi p5ig



eOADE Engineering Software PVF.lite <1.00 Licensee: eOADE fnc., Local l'lllite Lock PileName : Seminar - ----- --- -- -- - - - - - - - - - - -- - -- ------- -- - Page 9 Shell Analysis 0101 EXT HEAD Item: 4 9:51a Sep 21,2000 EHANP

B/(KO~

(O/T))

=

10051.5273/( 0.9000 • 166.8571 I

Results for Reqd Thickness for Ext. Corroded Thickness of Shell

Pressure TeA

=

66.9337

(Tca): 0.3882

in.

146.0000 in. Outside Diameter of Shell on (0/1') 376.0957 Diameter / Thickness Ratio Geometry Pactor, A [(DT ,I,D) 0.0003693 A '·laterials PactaI', B, f(A, Chart I B 5077.7563 psi I·laximum Allowable l'lorking Pressure 15.00 psig Ef·tM1P = B/ (KO· (0/1')) = 5077.7563/ ( 0.9000' 376.0957) = 15.0014

(

SI.Rolf·IJ\RY of EXTERNAl. PRESSURE RESULTS: Allowable Pressure at Corroded thickness Required Pressure as entered by User Reguired Thickness including Corrosion all. Actual Thickness as entered by User The PV Elite Program,

Examplc Problcm I-A Simple DnUll

66.93 15.00 0.5132 1.0000

psig psig in. in.

(cl 1989-2000 by COAOE Engineering Software

4-35

Nozzle Reinrorccmcnl 'Illd Failure P"th C:llculations


NOZZLE REINFORCEMENT AND FAILURE PATH CALCULATIONS The Code implements an "area replacement" concept: Near an opening there should be "extra" metal equal to the area of required metal missing due to the presence of the opening.

This area needs to be replaced because of stress intensification at the hole. A hole in a fairly large cylinder is roughly the same as a hole ill a flat plate loaded in plane. Stress intensification factors for this case arc Axial Load Only

3.00

Cylinder

2.50

Sphere

2.00

A second consideration in the design of the nozzle is making the connection strong enough so that pressure loads will not pop the nozzle off the vessel.

The Code rules for nozzle reinforcement and failure path calculations arc found in paragraphs UG-37 to UG-41. The Code takes into account the differing strengths of different

materials (i.e. the nozzle, the reinforcing plate, the shell) by using strength reduction factors on any area where the material used has a lesser strength thnn the basic shell material. All vessel and nozzle thickness calculations arc based on seamless components - E is always equal to I.

If the opening is ill the spherical portion of a torispherical head, the required thickness is based on a torispherical head with M= 1. This is a result of the lower stress in the central portion of a head, and the lower stress intensity at a hole in a sphere. If the opening is in an elliptical head and is inside a circle based on 80% of the diameter, the required thickness is based on a seamless sphere with a radius of90 percent of the vessel diameter.

For external pressure, the Code requires that only one half of the required area removed be replaced. The required area is based on the niles for external pressure. For nat heads (Paragraph UG-39), the Code requires that only one half of the required area

removed be replaced, because the head is loaded in bending rather than in tension. Limits of reinforcement (Paragraph UG-40) are intended to assure that the metal used to replace missing metal is close enough to the hole to be effective. They arc based on the characteristic length, or decay length of the vessel: 1.56 * SQRT(R *t)

This is an extremely eomlllon expression in pressure-vessel design; it describes the distance along a vessel shell in which the efTeet of any load will be felt. Beyond this distance, the effcct of the load has decayed to ncar zero.

The Code approxilnatcs this distance as 2x the corroded inside diameter orthe hole along the shell, and 2.5 times the thickness of the nozzle radial to the shell (based on rlt ~ 10). Small openings (Paragraph UG-36) arc exempt from reinforcement requirements: if t<=3/8, 3 in. and smaller are exempt if t> 3/8, 2 in. and smaller are exernpt

4-36

Example Problem I-A Simple Drulll

Pressure Vessel Design nile! Allnlysis

*

Seminar Notes

Nozzle Reinforcement nile! Fnilure Pnlh Calculntiolls

Large openings have special niles which are located in Appendix 1-7. These rulcs required that more oCthe reinCoreement be placed close to the opening. They apply when one oCthe Collowing statemenls is tme: The diameter oClhe nozzle is greater than 20 ill. or greater than D/2 when the vessel is 60 in. or less. The diameter of the nozzle is greater than 40 in. or greater than D/3 when the vesscl is

more than 60 in. Bending-end membrane stresses are computed Cor some "large" nozzles per 1-7.

Strength of reinforcement and failure path calculations arc found in paragraphs UG-41, UW-15, UW-16. The Code requires that the strength of any failure path through the nozzle and its reinCorcement be grcater than the strength oCa Cailure path through the vessel shell. There arc two basic placements oC nozzles Cor which the Cailure path becomes an issue: I)

inserted through the shell wall and 2) abUlting the shell wall. Area of reinforcement calculations arc diITerent Cor the two cases. Paragraph UW-I 5 (b) exempts many of Ihe commonly used nozzle attachment geometries Crom strength calculations.

Example Problem I-A Simple Drum

4-37

Pressure Vessel Design and Annlysis - Seminar Notes

Nozzle Reinforccment ma\ Fnilurc Pouh Calculnliulls

11.2

STRESSES AND LOADINGS AT OPENINGS

/-67t;

Nf'i

337

""'...

1,0 (axial onlY) (,)

" ~

Ullj

,

IOf£,

o.u;

o_~~

tnt, I

L

201 (cYlinder) (b)

~

r.

11 1 (sphere)

(el

Figure to-Stress Intensification at a Hole in a Flat Plate under Various Loadings

4-38

Example Problcm I-A Simple Drum

DIVISION I

....

, :

4-39

Nozzle Reinforcement :llld Failure Path Calculmions

1998 Sr:cnON VIll

r

Figure 2-- Typical Nozzle Configurations

Pressure Vessel Design ilnd Al1<1lysis - Seminar Notes

"g. UW·16.t


...

z

1-

o

o

~ ~

;,-

'"

~,

C'

ri

9

~

i

Co

:,;'

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."

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FIG. UW-lb.l SOME ACCEPTABLE TYPES OF WELDED NOZZLES AND OTHER CONNECTIONS TO SHELLS,

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4-41

Nozzlc Reinforccmcllt ;Jml F3ilufe P31b Calculations

1998 SECTION VIJI

,

E

E

", .}

,

. -":. o

,"

Figure 4-- Typical Nozzle Configurations (continued)

"

.}

d

./

Pressurc Vessel Dcsign ;Jnt! Analysis - Scminar

Fig. U'oV·16.1

--",,

"

Exal11ple Problem I-A Simple Dnllll

Nozzle Reinforccment and Failure Path Calculations

Pressun: Vessel Dcsign ;llld Analysis - Scminar Notes

PART UG -

Fig. UG-n.l

GENERAL REQUIREMENTS

U •• lOIOff v.l.... Ia< noul. WlllrUo1.d

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FIG. UG·)7.1 fWMENCLATURE AfW rORMULA$ FOR REINFORCED OPENltlG$ !This Flgurt IIlustrales a Common tlOlllt Confi9uraUon and Is Ilot Intended to PIohlblt Other Confi9uratlons Permitted by t~ Code.>

Figure 5··Area Replacement Calculations for Typical Nozzle Configurations

4·42

Exmnplc Problem I-A Simple Drul1l

Prc:-surc Vessel Design and Analysis - Seminar Noles

PART UO -

Nozzle Rcinforn:nu:n( and Failure Path Calculations

GENERAL REQUIREMENTS

Fig. UG41.l

.,

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III ArUI AI. A J• A). ~. rid A.,I" tJ'Iodifiotd by I nt riltlOlJ. lbl Nomtn<:IUurt Illh.llm.nln UG·)) ud fig. UG-H. I.

101 D.pl<:1. Tvpic.ll NOlll. 0'1.11 Wi1h Nick In.."teI Through 1M VI",,1 Will

FIG. UG-41.1 NOZZLE ATTACHMENT WElD LOADS APID WElD STRENGTH PATHS TO BE COIlSJOEREO

Figure 6--Failure Paths (or Typical Nozzle Configurations


4-43

Nozzle Rcinfon:emcnt and Failure Path C
Prc,:.;surc Vt:ssel Design ami Annlysis - Scminar Notes

1998 SECfION

Fig. UG·41.1

vm -

DIVISION I

....;:..

~

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Ll th

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.. (AI. A 41 1S..

GENERAL NOTES: (al Afeils AI' AI.II~. and AHilfO modified by ',xlaetofs. (hl Nomenclatufl is the same as in UG·37 ilnd Fig. UG·37.1.

'IM D.pict. TypiClI Nonl' 0.1111 With Nick Abutting lhe Vossel Well

FIG. UG-41.1 NOZZLE ATTACHMENT WELD LOADS AND WELD STRENGTH PATHS TO BE CONSIDERED lCDNI'D)

Figure T--Failure Paths for Typical Nozzle Configurations (continued)

4-44

Example Problem

I~A

Simple Drum

Pressure Vessel Design and Analysis - Scminar Notes

Hillside and Orr-Angle Nozzk Anglcs

HILLSIDE AND OFF-ANGLE NOZZLE ANGLES There are two main categories of ofT-angIe vessel nozzles: those which are off-angle in the longitudinal planc of the cylinder (Y-angle nozzles, Figure 23), and those which are orr angle in the circumferential plane ofa cylinder, or in a head (Hillside nozzles, Figure 24). In order to analyze these kinds of nozzles, CodcCalc asks the user for the angle between the nozzle centerline and a tangent to the vessel mean radius, as also showll in Figures 23 and 24. However, for hillside nozzles, the determination of this angle can be difficult. This section provides

a few simple equations that can help the user determine the nozzle angle.

The overall goal of this calculation is simple: to find an angle that will make the finished diameter of the hole fit perfectly within the nozzle. The finished diameter is the dimension

d in Figures 23 and 24, which is called DLII in the CodeCalc print-outs. Ifwe had the nozzle in front orus and could measure d, thcn we could calculate the input angle vcry simply lIsing the following equation:

where dn:=: inside diameter of nozzle

d

~

DLR

~

finished diameter of holc

sinn:=: sine of angle betweellllozzic

and vessel

Whcn we analyze Y-angle nozzles, this is really all the information we need: the angle is known and the result is exact:

d

~

DLR

~

!."SInn

However, when we analyze hillside nozzles, as shown in Figure 2, the angle is usually not

known. Instead, we may know the offset distance for the nozzle. This distance (L) is the distance between the centerline of the cylinder or head, and the centerline of the nozzle. A first approximation to the angle would take thc cosine of the angle as L I rm, where nll is the mean cylinder or head radius at the point of attachment. However, this approximation turns out to be too inaccurate for practical use.

The ASME Code has a sample problem (L-7.7) that shows what their preferred method appears to be. They do not explicitly address this off-anglc problcm in the body of the Code. Figure 3, taken from ASME (Section VIII, Division I, Addenda An, Page 512), shows their sample problem. The key to their approacli is the calculation of two angles, al and a2, and then the calculation of the finished diameter from the difference between these two angles. You can follow their calculation on Page 512 aml513 of the Code. For our purposes, we do not need to eany the calculation that far. The angle we are looking for, a

isjust the average of the two Code angles as calculated previously. The following equations show how to calculate this angle:

a

0:

I

r,,)

arccos ( L_- _

r

"'

L + ,. )

U

n

=

-J =

,.

arccos ( - - "

'"

al + a2 arccos--

Examplc Problcm I-A Simple DIlJlll

2

4-45

Hillside ami On'-Angle Nozzle Angles

Where

Pressure

Vl~55e1

Design and Annlysis - Semil1nr Noles

L ~ offsct distancc cylinder / head ccnterline

rn = inside nozzle radius

rill

=

mean vessel radius

These three equations can be used without any further information for any hillside nozzle in a cylinder. However, you need to apply thcm carefully to hillside nozzles in heads. When a hillside nozzle is in an elliptical or torispherical head, the nozzle may be located in thc spherical portion of the hcad, the toroidal portion of the hcad, or it may straddle the

two portions. This is shown in Figure 4. Each of these cases requires a slightly different L and I'm to be used in equations. When the nozzle lies entircly within the spherical portion of the head (Figure 4(a)), L is simply the offset from the head ccnterline, and I'm is the spherical radius of the head. For spherical or torispherical heads, this should be a known radius (Code dimension L in Figure 1-4 of Appendix I, for example).

Note

The Code uses inconsistent terms herc.

For elliptical heads, the spherical portion is taken to be a circle drawn on the head with a diametcr of 80 percent of the head diameter. The radius of the spherical portion is taken to be 0,90 times the head diameter. The nozzle offset from the vessel centerline should be

known fr0111 the vessel drawings. The nozzle can also lie entirely in the knnckle portion of the head (Figure 4(c)). The mean

radius (1'",) is the mean knuckle radius, and the offset (L) is distance from the origin of the knuckle radius to the centerline of the nozzle. Note that for an elliptical head, the knuckle is defined as anything outside a circle drawn on the head with a diamcter of 80 percent of the head diameter. The knuckle radius is 0.17 times the vessel diameter. Finally, the nozzle may be located so that parI of the nozzle is in the spherical portion, and part in the knuckle (Figure 4(b)). In this case, the angle at the part of the nozzle in the spherical portion should be caleulated as described for Figure 4(a), and the angle at the part in the knuckle portion should be caleulatcd as described for Figure 4(c). That is, calculate the inside angle using the spherical radius of the head and olTset from the ccnterline.

4-46


Ilillside and Off-Angle Nuzzle Angles

Calcultlte the outside angle using the mean radius of the knuckle and the offset from the knuckle origin.

Figure 8-- Y-angle Nozzle: Nozzle angled in the longitudinal plane 01 a cylinder


4-47

Ililiside ant! Olr·l\nglc Nozzle Angles

Prc:-;surc Vessel Design ,lilt! AlI:tlysis - Seminar NOles

L

=

Offset Dislonce

Figure 9--Hillside Nozzle

-....-"'--r---: ,7

_.Y

/

1/2 in. lillel_,

... /

! / /

' .... , ..:

r7'77"?/" [/ / '. -.-r.--.-t~-._. ;.:' R=15 in.

/

-In in.

Rn=2in. ./

./

/ ,,,,/ aJ/

j

/

I

a2

.,

" \

\i

\

\ ,

r

-L~12in.---

;

\, i

I

1·1/2 ill.

Figure 10·-ASME Code Figure L-7.7, Example of Reinforced Opening

4-48

Example Problem I-A Simple Drulll

Pressure Vessel Design and Analysis .. Scminar Noles

Ililiside anti Olr-Angle Nozzlc Angles

r-L-l A'

I

'

\

II Spherical 1'0l1ion

';-i< I

fI

{ '.L1

~

Tordial

Knuckle Radius

\ \

Spherical Radius

I--- Lo _ .

\29

1~-rll1--1

P0I1ioll

Examplc Problcm I-A Simple Drum

4-49

Hillside and

Orr-Angl~

Pressure Vcssel Design lind Analysis· Scminar Notcs

Nozzlc Angles

r-

L

___ I 'I' oruw "'1'

Portion

Figure l1--Hiflside nozzles in /leads

What do you need to know to perform nozzle reinforcement calculations?

Required Thickness of Head or Shell and Nozzle The thickness of an elliptical head is analyzed as "n equivalent spherical head, as speci fied in the Code, paragraph UG-37 (a), SimilarlY,thc thickness of the spherical portion ofa

torispherical head is analyzed using the same method and code reference. If your nozzle is outside 80% of the diameter of an elliptical head, or in the toroidal porlion of a torispherical head, you must calculate the required IhickJless using the rules of UG-32 rather than the I1Iles described above (but wilh E= 1.0 in all cases).

You must also calculate the required thickness for a flat head. However, you may reduce the required area of reinforcement by 50% for a Oat head.

For external pressure, you must calculate the required thickness using the guidelines in Paragraph UG-28 for cylinders and UG-33 for heads. Once again, however, you may

reduce the required area ofreinforeement by 50 percent for external pressure.

Geometry of Nozzle and Shell You must know the diameters, thicknesses, and corrosion allowances for both the shell component (cylinder or head) and the nozzle. Most nozzles are constnicted from pipe or long weld neck Oanges, bUI occasionally from plate (for large nozzles) or from integrally

reinforced inserts. The Code slates that metal usually associated with the pipe fabrication tolerance (mill undertolerance) may be included in lhe calculation of the available area in a nozzle. Sec paragraph UG-16(b),

The nozzle may be inserted through the vessel wall. Ifso, any metal in the insert, less corrosion, is available for reinforcement.

4-50

Example Problcm I-A Simple Drum


lIillside and Orr.Anglc Nozzle Angles

Diameter Limit, Thickness Limit Thc diamcter limit is the maximum distance, frolll the centerline of the nozzle along the vessel wall, that can be calculated to delermine available areas in the shell or a pad. The Code calculates this limit based 011 the corroded ID oflhe nozzle or the inside radius of the nozzle plus the shell thickness plus the pad thickness. YOlllllust be conscious orallY physical obstmctions (other nozzles, welds, the end of the vessel, ctc.) which would prevent the nozzle from making usc of reinforcement out to this diameter. The thickness limit is the distance, from the vessel surface along the nozzle axis, that can be taken credit for when calculating the areas available in the nozzle wall and the pad. The Code calculates Ihis limit based on the thickness of the shell and pad or the nozzle wall. You must be conscious of any physical obstnlctions (no insert, studding pad, welding outlei, etc.) that would prevent the nozzle from making use of reinforcement out to this thickness.

Is the Nozzle in a Seam? Ifso, the seam emcicncy is used in the "area available" calculations to reduce the area available in the shell. However, for the shell thickness calculations themselves, the seam efficiency is always 1.0.

Details of Nozzle Welds The welds that attach the nozzle and the reinforcing pad to the shell are important because thcy contribute to the arca available for reinforccment and because they frequently control the failure path calculations. Most nozzles arc attached with full penetration groove welds, plus required fillet welds. Some of these geometries arc exempt from failure path calculations. Specifically, UW-15(b) indicates that no strength calculations for nozzle allachment welds are required for figure UW-16.I, sketches (a), (b), (c), (d), (e), (f-I), (f-2), (f-3), (f-4), (g), (x-I), (y-I), and (z-I). Nozzle weld thicknesses are based on Figure UW-16. I. The outward nozzle weld is compared to the cover weld required by the Code. Note that the mii,imum dimension of a weld is 0.7 times its leg dimension. Note also thaI for cover welds the maximum weld the Code requires is 0.25 in. The pad weld requirement is typically at least one half of the element thickness. In addition to the cover welds, the total groove weld plus cover weld for inserted nozzles must bc at least 1.25 timcs thc minimum c1emcnt thickness.

Large Nozzle Considerations For large diameter nozzles, the rules of Appendix 1-7 rcquire that two-thirds of the reinforccmcnt be within 0.75 of the natural diameter limit for the nozzle. If the calculated value of the percent within this limit is greater than 66%, the nozzle is adequately reinforced for the large diameter rules.

Manway or Access Opening If this nozzle is a manway or access opening, the code docs not require the minimum noz· zle ncck thickness calculations per UG-45 be performed.

Example Problcm I-A Simplc Dnllll

4-51

Ilillside and Olr-Angle Nozzle

4-52

Allgle~

Pres~urc

Vessel Design
Examplc Problcm I-A Simple Drum

Ililiside mul On.. . 1\ngle Nozzle Angles

Pressure Vessel Design ;'Iud An;llysis - Seminar Notes

Pressure Vesscl Design and Analysis Scminar

Component Design Problem

D-IOI Nozzle A:

Location Size Material

Cylinder 12 in. (300 NS) SA 106, C

Thickness

Schedule 80

Flange Class Flange Material

300 1.1

The Nozzle is inserted into the vessel wall. There is a .687-in. (17.45 mm) partial penetration groove weld between the vessel wall and the nozzle neck. Nozzle Outside Projection is 8.0 in. (203 nun) No Internal Projections arc allowed in this vessel. There is a 22-in. (559 mm) diameter reinforcing pad The Pad material is the same as the Shell material. There is a full penetration weld between the pad and the nozzlc ncck. There is a 0.375-in.(9.55 mm) cover weld between the pad and the nozzle neck. There is a 0.625-in. (16 mm) weld between the OD of the pad and the vessel. Notes:

Questions: What pad thickness (ifany) will you use? What is the M.A.W.P. for the assembly? Docs this nozzle limit the M.A. w.P. for the vessel?

_

DO NOT TURN THE PAGE UNTIL YOU HAVE COMI'LETED YOUR ANALYSIS

EX,Hllpic Problem I-A Simple Dnull

4-53

Ilillside

~lld

Pressur~

Orr-Angle Nozzle Angles

Vessel Design

~lId

Anillysis - Seminar Noles

COMPLETED YOUR ANALYSIS PVElite 4.00 I~icensee; COADE Inc., Local \'lhite Lock FileName; Seminar ------------ ... -.---------------------- Page 14 Nozzle Analysis; 0101 Nozzle A Item; 9;Sla Sep 21,2000

Input Echo, Nozzle Item

Description: 0101 Nozzle A

Design Internal Pressure Case Temperature for Internal Pressure

P TEI·IP

230.00 450.00

psig F

SI\-516 70

Shell Haterial (Normalized) Shell Allowable Stress at Temperature Shell Allowable Stress At Ambient

20000.00 20000.00

psi psi

CAS

144.0000 1. 0000 0.1250

in. in. in.

Angle between Nozzle and Shell or Head

ANGLE

90.00

Haterial (Not Normalized or NI\) Allowable Stress at Temperature Nozzle Allowable Stress At Ambient

SN SNA

SA-I06 C 20000.00 20000.00

S SA

o

Inside Diameter of Cylindrical Shell Actual Thickness of Shell or Head Corrosion Allowance for Shell or Head

T

~ozzle

~ozzle

Diameter Basis for Nozzle Diameter of Nozzle Nozzle Size and Thickness Basis Thickness of Nozzle corrosion Allowance for Nozzle Joint Efficiency of Shell Seam at Nozzle Joint Efficiency of Nozzle Neck

BASISN

DIA

Degrees

psi psi

00 12 . 0000

in .

EN

Nominal SCH 80 0.1250 in. 1.00 1.00

Insert or Abutting Nozzle Type NTYP Outward Projection of Nozzle HD Weld leg size between Nozzle and Pad/Shell >1O Groove weld depth between Nozzle and Vessel WGNV

Insert 8.0000 in. 0.3750 in. 0.6870 in.

~ominal

DBN

THKNOi-l

Pad t·laterial (Not Normalized or NA) Pad Allowable Stress at Temperature Pad Allowable Stress At Ambient Diamet.er of Pad along vessel surface Thickness of Pad Weld leg size between Pad and Shell Groove weld depth between Pad and Nozzle

CAN

ES

SN SNA

DP TP >lP Io'IGPN

Is this is '·lanway/Access/Inspection Opening Skip Iterative Failure Thickness calculations

SA-516 70 20000.00 20000.00 22.0000 0.7500 0.6'-50 0.7500

psi psi in. in. in. in.

No

Yes

NOZZLE CALCULATION, NOZZLE NUI·lBER 1, Description: 0101 Nozzle A ASf·1E Code, Section VIII, Division 1, 1998, A-99, UG-37 to UG·45 Actual Nozzle Diameter used in calculation Actual Nozzle Thickness Used in Calculation

12.750 0.687

in. in.

Internal Pressure Results {or SHELL/HEAD

4-54


Ilressurc Vcsscl Design owl! Analysis - Seminar Noles

Ilillside and OlT.Angle Nozzlc Angles

COAOE Engineering SO[Lware PVElite 4.00 I.icensee: COAOE Inc., Local White Lock FileName: Seminar -. -_.- -- ---- --- - - -- - - - -. - - - - -- - Nozzle /\nalysis

0101 Nozzle A

Item:

Page IS

9:51a Sep 21,2000

Required thickness per UG-H(a) o[ Cylindrical Shell, Tr

CASE 1

Thickness Due to Internal Pressure (TR): lP·(D/2tCA»/(S'E-0.6'P) per UG-27 (c) (1) (230.00 t (144.0000/2tO.1250»/(20000.00 t l.00-0.6 t 230.00) 0.8352 in. InLernal Pressure Results for NOZZLE : Required thickness per UG-37(a) of Nozzle Wall, Tr

CASE 1

Thickness Due to Internal Pressure (TRI: (P'D/21/(S'EtO.4 t P) per Appendix 1-1 (a) 01 (230.00'12.7500/2)/(20000.00"1.0010.4*230.00) 0.0730 in.

,

UG-40, Thickness and Diameter Limit Results CASE Effective material diameter limit, DL Effective mat.erial t.hickness limit, no pad TLNP Effective material thickness limit, pad side TLI'/P RESULTS of NOZZLE REINFORCEr·IENT AREA CAl,CUl,ATIONS: AREA AVAILABLE, A1 t.o A5 Design External Area Required AR 9.710 NA Area in Shell A1 0.463 NA Area in Nozzle 1'laiI NA A2 2.108 A) Area in Inward Nozzle 0.000 NA Area in Helds A4 0.531 NA NA Area in Pad AS 6.938 TOTAL AREA AVAILABLE ATOT NA 10.039

2) . 2520 1 .4050

2. 1550

in . in. in.

r'lapnc

NA NA NA NA NA NA NA

in in in in in in in

Pressure Case 1 Governs the Analysis Nozzle Angle Used in Area Calculations

90.00

Oegs.

The area available without a pad is Insufficient. The area available with the given pad is Sufficient. SELECTION OF POSSIBLE REINFORCING PADS: Based on given Pad Thickness: Based on given Pad Diameter: Based on Shell or Nozzle Thickness:

Diameter 21.6250 22.0000 22.3750

Thickness 0.7500 in. 0.7500 in. 0.6875 in.

Reinforcement Area Required for Nozzle: AR (DLR'TR-t2·THK·TR'O-FFR1)l UG-37(cl or UG-39 AR (11.6260"0.8352.2*(0.6870-0.1250)·0.8352·(1.0-1.00) AR 9.710 in Areas per UG-37.1 but with DL = Diameter Limit, DLR = Corroded 10: Area Available in Shell (Al): Al (DL-DLRI· (r:S' (T-CAS) -TRI-2* (THK·CAN) * (ES· (T-CASI-TR)*l1-FFRl) 1\1 '" (23.252-11.626)' 0.00' (l .0000-0.125) -0.835) -2. (0.687-0.125) • n.oo' 0.0000-0.1250) -0.8352)' (l.0-1.00)

Example Problcm I-A Simple Dnull

4-55

llillside ;lnd

Orr.. Angle Nozzle Angles

Pressure Vessel Design ;lnd Analysis .. Seminar NOles

COADE Engineering Software PVElite 4.00 Licensee: COAOE Inc., Local White Lock FileName: Seminar ------Nozzle Analysis: DI01 Nozzle A Al

=

--

Item:

Pilge 16

9:51a Sep 21,2000

0.463 in

Area Available in Nozzle Hall, no Pad; ( 2 • tHN (TLNP, 110) ) • (THK CAN A2NP ( 2 • 1.4050 ) • ( 0.6870 .. 0.1250 fl2NP 1.3"/4 in A2NP

TRN ) " FFR2 0.0730 ) , 1.00

Area Available in Nozzle Hall, with Pad: A21-1P (2'llIN(TLWP,HO))' (THK"CAN .. TRN) 'PFR2 ( 2 • 2.1550 ) , ( 0.6870 .. 0.1250 - 0.07301 A21'IP A2\'IP 2.108 in

'

1.00 )

Area Available in Welds, no Pad: \'IO"'2'FF'R2 t ( 1'/I .. CAN/0.707 )"'2'FFR2 MNP MNP 0.3750"'2 • 1. 0000 t ( 0.0000 )"'2 • 1. 0000 A4NP 0.141 in Area Available in welds, with Pad: A4HP ~0 ... 2·FFR3t(WI .. CAN/0.707) ... 2·FFR2tWP ... 2·FFR4 A4HP 0.3750·2 • 1.00 • ( 0.0000 )·2 • 1.00 t 0.6250·2 • 1.00 A4\'JP 0.53} in Area Available in Pad: AS (lH N (DP, OL) .. (OIlH 2 'TIlK) ) " (Hi n (TP, TLI'IP, TE) ) 'FF'R4 AS (22.0000" 12.7500 ) • 0.7500 ' 1. 00 A5 6.938 in UG-45 Hinimum Nozzle Neck Thickness Requirement: '" t·lax (t.Jin (Hax (/·lax (UG4 S81, UG16B) ,Hax (UG4582, VG168) ) , VG4 584) , VG4 sA) f.lax(t-lin(Nax(f·lax( 0.9602, 0.1875),'·lax( 0.1250,0.1875),0.4531),0.1980) 0.4531 < Minimum Nozzle Thickness 0.6011 in. OK l-I.A.W.P. RESULTS FOR THIS NOZZLE GEOHETRY Approximate f
AHAP

Weight of Nozzle, with Pad, Vncorroded Weight of Nozzle, with Pad, Corroded /·1 IN If.\UH Hinimum rHnimurn /Hnimum

119.90 106.49

DESIGN t·lETAL TEI·lPERATURE RESULTS: Nozzle Temp. w/o impact per Fig. UCS-66 11 -129 Temp. at operating stress Temp. w/o impact per OO .. 20(f) -20

Nozzle HOHT Thickness Calc. per IHn. J.letal Temp. w/o impact per I·lin. "'etal Temp. at Req'd thk. l-lin. f·letal Temp. w/o impact per

Shell

-30 -35 -20

UCS-66 l(b), HIN(tn,t,te) Fig. UCS .. 66 11 (per UCS 66.1) ·129 UG .. 20(f) ·20

WELD SIZE CALCULATIONS, NOZZLE Nrn.JBER

1,

Desc.

l·linimum thickness for nozzle/shell welds Tmin I·linimum thickness for pad/shell welds TminPad

4-56

233.8

psig

lb. lb.

Pad 16 12 -20

,. F F

P F F

0101 Nozzle A 0.5620 0.7500

in. in.


Pressure Vessel Design ;md A1l3lysis. Seminar Noles

Ilillsidt' anti

Orr.. Anglc Nozzle Angles

COADE Engineering Software PVElite

4.00

I~ocal

Licensee: COADE Inc.,

i'lhite Lock

PileName Seminar .. --- .. ----- .. ---- .. - .. - .. - .... -----------Page 17 Nozzle Analysis 0101 Nozzle A Item: 9:51a Scp 21,2000 Results Per UH-16.I, Nozzle Held

Pad weld

Actual Thickness Required Thickness 0 .. 7 WO 0.2500 !-lin per Code 0.2625 0.3750 ::: 0.5'TNINPAD 0.4375 = 0.7 ' WP

HELD STRENGTH AND \'IELD LOADS PER UG-41.1,

SK~'TCH

(al OR

W

(AR-Al+2* ITHK-CAN) 'PPR1'(El(T-CASI .. TR)'S

1"1

19.7100 - 0.4627 t I 1.00 '

\1 \11

2

'

10.6870 - 0.1250 I

( 1.0000 - 0.1250)

.. 0.8352 I

I

'

ill.

in.

(bl

'

1.0000

20000

185841 .. lb. IA2tA5tA4-(WII-CAN/.707)A 2 *PPR2)"5

WI

(2.1077

m

190904.

t

6.9375

1

0.5312

.. 0.0313 '

W2

(A2tA3tA4t(2'(THK-CAN)'(T-CAS)'Prl)'S

~12

(2.1077 -t 0.0000 64636. lb.

1'/2

1.00 I

'

20000

lb. t

0.1406 .. 0.9835 )

• 20000

In

(A2IA3+A4tA5, (2' (THK-CAN)' (T-CASI'Prll) '5

1'13

(2.1077 t

1'/3

211199.

0.0000 .. 6.93"'5

I 0.5312 .. 0.9835 )

• 20000

lb.

STREN"GTll OF CONNECTION ELE1·1ENTS FOR PAl LURE PATH ANAl.YSIS SHEAR, OUTI'IARD NOZZLE \'IELD: SONI'I SON\'/

(PI/2) 'OLO'~,'O'0.49'SN\~ (].1416 / 2.0 ) , 12.7500 • 0.3750 '

SONW

13602. lb.

0.49'

20000

0.49 '

20000

SllEAR, PAD ELEI·IENT \'IELD: SPEH (PI/2)*OP'WP*0.49*SEW SPE\~

(3.1416

SPEW

211665.

SHEAR,

I

2.0 )

,

22.0000 '

0.6250 '

lb.

NOZZLE WALL:

SNW SN1~

IPI'(DLR-tDLO)/4.0) '(THK-CANI '0.7'SN 13.1416 * 6.0940) , ( 0.6870 - 0.1250 I

S~l

150632. lb.

TENSION, TPGW

0 .. 7 '

20000

PAD GROOVE WELD: (PI/2.0) 'DLO'WGPN'O. 74'5EG

TPG~1

( 3 . 1 4 1 6 / 2 . 0 ) ' 12.7500 '

TPGW

222307.

TENSION,

'

0.7500 '

0.74

* 20000

lb.

NOZZLE GROOVE WELD:

TNGW TNml

(PI/2)'DLO*(WGNVI·CASI*0.74'SNG (3.1416 I 2.0 I ' 12.7500 ' 10.6870 - 0.1250)

TNGW

166582.

,

0.74

'

20000

lb.

STRENGTH OF FAILURE PATHS: PATH 11 PATH22

(SPEI1. SNW 1 '" ( 211664 of 150631 I " 362296 lb. (SO~I -t TPGW of TNGW f SINN I

PATIO 3

( 7]601 -t 222306 t 166582 .. 0 I " 462490 lb. (SPEI'1 t TNGWt SINlrl I ( 211664

Examplc Problcm I-A Simple Dnllll

-t 166582

~

0 I

=

318246 lb.

4-57

Ilillside tlnd Orr.Angle Nozzle Angles


COADE Engineering soft.ware PVElite 4.00 Licensee: COADE Inc., Local \·/hit.e Lock FileName Seminar -------.--------------------------Page 18 Nozzle Analysis: 0101 Nozzle A Item: 9:51a Sep 21,2000 SUHt·1ARY OF Path 1 -1 Path 2-2 Path 3·3

FAILURE 362297. 462491. 318247.

PATH lb. , lb .• lb.


4-58

CALCULflTJONS: must exceed I·j must exceed I'j must exceed I'j

(e)

185841. lb. 0< 185841. lb. 0< 112 185841. lb. 0< 113

'"

190904. 64636. 211199.

lb. lb. lb.

1989-2000 by COAOE Engineering Software


Pressure Vessel Dcsign and Analysis - Scmilltlf Nolcs

Ilillside nnd Off-Angle Nozzlc Angles

Pressure Vessel Design .11)(1 Analysis Seminar Component Design Problem

0-101 Nozzle B:

Location

Cylinder

Size Material

8 in. (200 NS) SA 106, C

Thickness

Schedule 80


300 1.1

The Nozzle is inserted into the vessel wall. There is a .500-in. (12.7 mm) partial penetration groove weld between the vessel wall and the nozzle neck. Nozzle Outside Projection is 8.0 in. (203 mm) No Intemal Projections are allowed in this vessel. There is a 0.75-in. (19 mm) thick reinforcing pad The Pad material is the same as the Shell material. There is a full penetration weld belween the pad and the nozzle neck. There is a 0.375-in. (9.5 mm) cover weld between the pad and the nozzle neck. There is a 0.625-in.(l6 mm) weld between the 00 of the pad and the vessel. Notes:

Questions: What pad diameter (if any) will you use?

_

What is the M.A.W.P. for the assembly?

DO NOT TUlm THE PAGE UNTIL YOU HAVE COMPLETED YOUR ANALYSIS

EXl1ll1plc Problcm I-A Simple Drulll

4-59

Hillside

tllH..I

On·.Anglc Noule Angles

PVElite 4.00 Licensee: COADE Inc., Local \'lhite Lock FileName: Seminar _ ... -~-------------------------------- Page 19 Nozzle Analysis: 0101 Nozzle B Item: 2 9:S1a Scp 21,2000


2,

Description: 0101 Nozzle 13

Design Internal Pressure Case Temperature for Internal Pressure

P

TEf·1P

230.00 450.00

psig F

7.

Shell Haterial (Not Normalized or NAJ Shell Allowable Stress at Temperature Shell Allowable Stress At Ambient

$A-516 20000.00 20000.00

psi psi

CAS

144.0000 1. 0000 0.1250

in. in. in.

Angle between Nozzle and Shell or Head

ANGLE

90.00

Nozzle r·latedal (Not Normalized or NA) Nozzle Allowable Stress at Temperature Nozzle Allowable Stress At Ambient

SN SNA

SA-51G 20000.00 7.0000.00

BASISN DlA

8.0000

S SA

Inside Diameter of Cylindrical Shell Actual Thickness of Shell or Head Corrosion Allowance for Shell or !lead

Diameter Basis for Nozzle Diameter of Nozzle Nozzle Size and Thickness Basis Nominal Thickness of Nozzle Corrosion Allowance for Nozzle Joint Efficiency of Shell Seam at Nozzle Joint Efficiency of Nozzle Neck

D

T

Degrees

7.

DBN

THKNO:·1 CAN

ES EN

Insert or Abutting Nozzle Type NTYP Outward projection of Nozzle HO Neld leg size between Nozzle and Pad/Shell 110 Groove weld depth between Nozzle and vessel \'IGNV Pad Haterial (Not Normalized or NA) Pad Allowable Stress at Temperature Pad Allowable Stress At Ambient Diameter of Pad along vessel surface Thickness of Pad Weld leg size between Pad and Shell Groove weld depth between Pad and Nozzle AS1·IE Code Weld Type per UW~16.1

psi psi

OD in.

Nominal

SCH 8. 0.1250 1.00 1.00

in.

Insert 8.0000 in. 0.3750 in. 0.5000 in. 70 20000 _00 20000.00 15.5000 0.7500 0.6250 0.7500

SA~SI6

SN SNA

DP TP I1P

I'IGPN

Is this is Hanway/Access/Inspection Opening Skip Iterative Failure Thickness Calculations


A No

Yes

NOZZLE CALCULATION, NOZZLE Nrn~BER 2, Description: 0101 Nozzle B ASME Code, Section VIII, Division 1, 1998, A-99, UG-37 to UG-45 Actual Nozzle Diameter Used in Calculation Actual Nozzle Thickness Used in Calculation

8.625 0.500

in. in.

Internal Pressure Results for SHELL/HEAD

4-60



Ilil1sidc and OIT-Angle Nozzle Angles

COADE Engineering Software PVElite 4.00 Licensee: COhDE Inc., PileName Seminar --- .. --------Nozzle Analysis

Local \'Ihite Lock

D101 Nozzle B

Required thickness per UG-]7(a) Thickness Due to Internal

Item:

Page 20 9:51a Sep 21,2000

of Cylindrical Shell, 1'1'

Pressure

(P*(D/2+CA))/(S*E-0.G*P)

2

CASE 1

(TR);

per UG-??

(c) (1)

(2]0.00*(144.0000/2+0.1250))/(20000.00*1.00-0.6*2]0.00)

0.8]52 in. Internal Pressure Results for NOZ:t,I... E : Required thickness per VG-)7(a)

of Nozzle \'Iall, 1'1'

Thickness Due to Internal Pressure (P*D/2)/(S*E-I0.4*P)

CASE 1

(TR):

per Appendix 1-1

(a) (1)

(230.00'8.6250/2)/(20000.00*1.00+0.4*230.00) 0.0491 in. UG-40,

Thickness and Diameter Limit Results Effective material diameter limit,

,

CASE

DL TLNP no pad pad side TLI'IP

Effective material thickness limit, Effective material thickness limit,

RESULTS of NOZZI,E REINFORCEl-lENT AREA CALCULATIONS: AREA AVAILABLE, A1 to AS External Design Area Required AR 6.577 NA Area in Shell AI 0.313 NA Area in Nozzle \~a 11 NA A2 1.099 Area in Inward Nozzle 0.000 NA A3 Area in \~elds Area in Pad TOTAL AREA AVAILABLE

M

0.281

A5

5.156 6.850

ATOT

15 .7500

in.

0 .9375 1 .6875

in. in.

f·lapnc

NA NA NA NA NA NA NA

NA NA NA

in in

in in in in in

Pressure Case 1 Governs the Analysis Nozzle Angle Used in Area Calculations

90.00

Degs.

The area available without a pad is Insufficient. The area available with the given pad is Sufficient. SELECTION OF POSSIBLE REINFORCING PADS:

Diameter

Based on given Pad Thickness: Based on given Pad Diameter: Based on the estimated Diameter Limit:

15.2500 15.5000 15.6250

Thickness 0 .7500 0 .7500

in. ·in.

0 .7500

in.

Reinforcement Area Required for Nozzle: AR AR AR

(DLR'TR+2'THK*TR'(I-PPR1))

UG-37(c)

or UG-39

('l. 8'150'0.8352+2' (0.5000-0.1250) '0.8352* (1.0-1.00) )

6.577 in

Areas per UG-37.1 but with DL '" Diameter r.imit, DLR'" Corroded ID: Area Available in Shell (AI): A1 (DJ,- DLR) • (ES' (1'- CAS) -TR) .. 2' (THK- CAN) , (ES· (T-CAS) -TR) * (1 - FPRl ) Al '" (15.750-7.875) * (1.00' (1.0000-0.125) -0.835) .. 2* (0.500-0.125)

Example Problem

I~A

Simple Dnlln

4-61

I lil1sick alld On'-Angle Nozzle Angles


(OADE Engineering Software PVElite

4.00

Licensee: COADE

II~C.,

Local

~'/hite

Lock Page 7.1 9:51a Sep 21,2000

FileName Seminar -----------------------.----Nozzle Analysis 0101 Nozzle B Item: 7.

Al =

• 0.00' n.0000-0.1250) -0.8352) ' n . 0-1.00) 0.313 in

..

Area Available in Nozzle NaIl, no Pad: A2NP tuN (TLNP, HO) I I 2 I TIIK - CAN A2Np ( 2 ( 0.5000 - 0.1250 0.9375 I A2NP 0.611 in

.

•

TRN I

•

0.0494 I

Area Available in Nozzle NaIl, with Pad: A2\'/P (2'IUN(TLI'IP,HO»' (TIlK-CAN-TRN) 'FFR2 A2NP ( 2 • 1.6875 ) ' ( 0.5000 - 0.1250 - 0.0494 ) A2\'/P 1.099 in

FFR2

•

l. 00 I

, 1.00 )

Area Available in \'/elds, no Pad: MNP \':O"'2'FfR2 .. ( WI-CAN/O. 707 ) ""2'FFR2 0.3750"'2 ' 1.0000 + ( 0.0000 )"'2 • 1.0000 MNP MNP 0.141 in Area Available in Nelds, with Pad: A'1\'lP ~:O"'2' FFR3 .. (m -CANtO . 707) '"'2' F'FR2+Trap' F'F'R4 A'1WP 0.)750"'2 ' l.00 .. ( 0.0000 )"'2 ' 1.00 .. 0.1406 , 1.00 A4rlP 0.281 in Area Available in Pad: A5 (lUN (op, DL) - (OIA. 2 'THK) ) , (f·lin (TP, TL\'/P, TE) ) • FFR'1 A5 (15_5000 - 8.6250 I ' 0.7500 .. 1.00 AS 5.156 in UG-'15 IHnimum Nozzle Neck Thickness Requirement: .. !·lax (Min (t,lax (Hax (UG4 581, UGl68) ,!·lax (UG4SB2, UGI6B) ) ,UG4SBlax(Hax( 0.9602, 0.187S),t-1ax( 0.1250, 0.1875», 0,4068), 0.1744) 0.4068 < Hinimum Nozzle Thickness 0.4)75 in. OK .A.W,P. RESULTS FOR THIS NOZZLE GEOHETRY Approximate M.A.ILP. for given geometry

M

A/otAP

\'Ieight of Nozzle, with Pad, Uncorroded Height Nozzle, with Pad, Corroded

0'

NINIHUM Hinimum Ninimum Hinimum

-.

DESIGN HETAL TEt·1PERATURE RESULTS: Nozzle Temp. wlo impact per Fig. UCS-66 -146 Temp. at operating stress Temp. wlo impact per UG-20(f) -20

234.7

psig

60.16 51.13

lb. lb.

Shell 31

Pad

,.

F

2. -20

12 -20

F

-.

Nozzle ~lDf·lT Thickness Calc. per UCS·66 l(b), IUN(tn,t,te) l-lin. l1etal Temp. w/o impact per Fig. UCS-66 rHn. Netal Temp. at Req'd thk. (per ucs 66.1) -146 r,Hn. !·letal Temp. w/o impact per UG-20(f) -20 HELD SIZE CALCULATIONS, NOZZLE Nm1BER

2,

F F F

oesc.: 0101 Nozzle B

IHnimum thickness for nozzle/shell welds Tmin Ninimum thickness for pad/shell welds TminPad

4-62

F

0.3750 0.7500

in. in.

Example Problcm I-A Simplc Drum

Hillside and Orr-Angle Noah: Angles


COl\DE Engineel"ing Softwal-e PVElite

4.00

FileName

Licensee: COADE Inc.

Local

~Ihite

Lock

Seminar

Nozzle Analysis

DIOI Nozzle B

Results Per V\·1-16.1, Nozzle \'Ield Pad \'Ield The PV Elite Program,

Example Problem I-A Simple DnHn

Required Thickness 0.2500 0.3'/50 (c)

=

Item:

2

Page 22 9:51a Sep 21,2000

Actual Thickness

0.7 NO I·lin per Code 0.2625 0.4375 = 0.1 • NP O.S·TIUNPAD

in. in.


4-63

Ilillsidc tint! Orf-Angle Nozzle Anglcs

4-64

Pressure Vessel Design tint! Antilysis - Scmintlr Notes


lIillside and Oll.. Angle Nozzle Angles


Pressure Vessel Design alld Analysis Seminar Component Design I)roblem

D-I 0 I Nozzle C:

Size Material

Elliptical Head 20 in. (500 NS) SA 516, 70 rolled plate

Thickness

.75 (19 mm)


300 1.1

Location

The nozzle is inserted into the vessel wall. There is a .500-in. (12.7 mm )parlial penetration groove weld

betwecn thc vessel wall and the nozzle neck. Nozzle outside projection is 8.0 in. (203 mm) No Internal Projections arc allowed in this vessel. There is a 0.75-in. (19 mm) thick reinforcing pad. The pad material is the same as the shell material. There is a full penetration weld between the pad and the nozzle neck. There is a 0.375-in.(9.5 mm) cover weld between the pad and the nozzle neck. There is a 0.625-in. (16 mm) weld between the 00 of the pad and the vessel. Notes:

Questions: Why is the required shell thickness different than previously calculated?

What pad diameter (if any) will you use? What is the M.A.W.P. for the assembly? Docs this nozzle limitlhe M.A.W.P. for the vessel?

_


Example Problem I-A Simple DnJlll

4-65

I lillsidc and Off-Angle Nozzle Angles

Pressure Vcssd Design and Analysis - Seminar Noles

COADE Engineering Software PVElite 4.00 Licensee: COf\DE Inc., Local I"lhite Lock PileName : Seminar ------------.- .. -~------ ... ------ .. ---. Page 23 Nozzle Analysis: 0101 Nozzle C Item: 3 9:51a Sep 21,2000


3,

Description: 0101 Nozzle C

Design Internal Pl'essure Case Temperature for Internal Pressure Shell J.laterial (Not Normalized or Nf\) Shell Allowable Stress at Temperature Shell Allowable Stress At Ambient

5 5A

Inside Diameter of Elliptical Ilead Aspect Ratio of Elliptical Head Actual Thickness of Shell or Head Corrosion Allowance for Shell or Head Angle between Nozzle and Shell or Head

230.00 450.00

P TEr-1P

SA-516 70 20000.00 20000.00

CAS

144.0000 2.00 1. 0000 0.1250

ANGLE

90.00

D

AR T

Is the Nozzle Outside the 80\ diameter Limit

p

psi psi in. in. in. Degrees

NO

Nozzle Material (Not Normalized or NA) Nozzle Allowable Stress at Temperature Nozzle Allowable Stress At Ambient Diameter Basis for Nozzle Diameter of Nozzle

psig

70 20000.00 20000.00

psi psi

OD 7.0.0000

in,

SA~S16

5N 5NA

BASISN DIA

Nozzle Size and Thickness Basis Actual Thickness of Nozzle Corrosion Allowance for Nozzle Joint Efficiency of Shell Seam at Nozzle Joint Efficiency of Nozzle Neck

EN

Actual 0.7500 in, 0.1250 in. 1.00 1. 00

Insert or Abutting Nozzle Type NT'fP Outward Projection of Nozzle HO Weld leg size between Nozzle and Pad/Shell HO Groove weld depth between Nozzle and Vessel HGNV

Insert 8.0000 in. 0.3750 in. 0.5000 in.

Pad '·later-ial (Not Normalized or NA) Pad Allowable Stress at Temperature Pad Allowable Stress At Ambient Diameter of Pad along vessel surface Thickness of Pad Weld leg size between Pad and Shell Groove weld depth between Pad and Nozzle ASME Code Weld Type per UW-16.1

DBN THK

CAN E5

516 70 20000.00 psi 20000.00 psi 32.0000 in. 0.7500·in. 0.6250 in. 0.7500 in.

SA~

5N 5NA

DP TP HP WGPN

Is this is Hanway/Access/Inspection Opening Skip Iterative Failure Thickness Calculations

A No

Yes

NOZZLE CALCULATION, NOZZLE NUt·lBER 3, Description: 0101 Nozzle C Section VIII, Division 1, 1998, 11..-99, UG-37 to UG-
Asr·1E Code,

Aclllal No%%le Diameter used in Calculation

4-66

20.000

in.


Prel-iSlIrt; Vcsscl Design and Analysis ~ Seminar Notes

Ilillsidc and Orr-Angle Nozzle Angles

COADE Engineering So[tl"are PVElite 4.00 Licensee; COADE Inc .• Local I'/hite Lock FileName Seminar -----------------------) I tern; Nozzle Analysis 0101 Nozzle C

Page 24 9:51a Sep 21,2000

Actual Nozzle Thickness Used in Calculation

0.750

in.

Internal Pressure Results for SHELL/HEAD : Required thickness per UG-37 (a) of Elliptical Head, TR,

CASE 1

Thickness Due to Internal Pressure (TR); (P'(D+2'CAj'K)/(2'S*E-0.2'P) Appendix l-<1(c) (230.00'(14<1.0000t2'0.1250)'0.90)/(2'20000.00'1.00-0.2'230.00) 0.7474 in. Internal Pressure Results for NOZZLE Required thickness per UG-37(aj of Nozzle l'lall, Tr

CASE 1

Thickness Due to Internal Pressure (TR): (P'O/2)/(S'EtO.<1'P) per Appendix 1-1 (al (1) ~ (230.00'20.0000/2)/(20000.00'1.00tO.4'230.00) 0.1145 in. UG-<10, Thickness and Diameter Limit Results , CASE Effective material diameter limit, DL TI,NP Effective material thickness limit. no pad Effective material thickness I imi t. pad side TLI'IP

3".5000 1.5625 2.1875

RESULTS oE NOZZLE REINFORCEI·lENT AREA CALCULATIONS; AREA AVAILABLE, A1 to A5 Design External At:ea Required NA AR 14.013 Area in Shell 2.393 NA A1 Area in Nozzle Wall A2 2.23<1 NA Area in Inward Nozzle O. (100 NA A3 Area in l'Ields A4 0.531 NA 9. (100 Area in Pad AS NA TOTAL AREA AVAILABLE NA ATOT 14.158

/olapnc NA NA NA NA NA NA . NA

in in in in in in in

90.00

Degs.

ill.

in. in.

Pressure Case 1 Governs the Analysis Nozzle Angle Used in Area Calculations The area available without a pad is Insufficient. The area available with the given pad is Sufficient. SELECTION oe POSSIBLE REINF'ORCING PADS: Based on given Pad Thickness: Based on given Pad Diameter: Based on Shell or Nozzle Thickness:

Diameter 31. 8750 32.0000 31.8750

Thickness 0.7500 in. 0.7500 in. 0.7500 in.

Reinforcement Area Required for Nozzle: AR (OLR'TR+2'THK'TR' (1-FFRll) UG-37 (c) or UG-39 AR (18.7500*0.7<174 t2' (0.7500-0.1250) '0 . .,474' (l. 0 -1.00) j AR 1<1.013 in Areas per UG- 37.1 but wi th DL

Example Problem I-A Simple Dnnll

Diameter Limit, DLR

Corroded ID;

4-67

Ilillside lllld Orr-Angle Nozzle Angles


COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock ---~.-.----••• ~- ••. ~Page 25 FileName: Seminar ---------Nozzle Analysis 0101 Nozzle C Item: 9:51a Sep 21,2000 Area Available in Shell (All: I'll (DL¥DLR)·(ES·{T-CAS)-TR)-2·{THK-CAN)'(ES·(T-CAS)·TR)·(1-FFR1) I'll

(31.500·18.750)·(1.00'(1.0000-0.125)~0.741)·2·(0.750~0.125)

I'll

• (1.00· 11.0000-0.1250) -0.74'/4)" (l.0-1.00) 2.393 in

Area Available in Nozzle Wall, no Pad: A2NP ( 2 ' t-lIN{TLNP,HO) ) • ( THK - CAN A2NP ( 2 • 1.5625 ) • ( 0.1500 - 0.1250 A2NP 1. 595 in

TRN )

,. FFR2

0.1145 ) • 1.00 )

Area Available in Nozzle l'Iall, with Pad: A2~'IP (2·r·lIN(TU·IP,1I0»)· (THK·CAN-TRN) 'FFR2 A21'IP ( 2 • 2.1815 ) • ( 0.1500 - 0.1250 - 0.1145 ) • 1.00 ) A2~'/P 2.234 in Area Available in Welds, no Pad: MNP I'IO"2·FFR2 t ( I'II-CAN/0.70'! ' .... 2·FFR2 0.3750"2 • 1.0000 • ( 0.0000 )A 2 • 1.0000 MNP MNP 0.141 in Area Available in Welds, with Pad: MHP 1'I0 2' FFR3 -t (\'/I -CAN/a. 707' "2' FFR2-t\'/P"2' FFR4 MWP 0.3150"2 • 1.00 -t ( 0.0000 )A 2 ,. 1.00 -t 0.6250"2 • 1.00 A

A4~'IP

0.531 in

Area Available in Pad: AS
AHAP

Weight of Nozzle, with Pad, Uncorroded weight of Nozzle, with Pad, Corroded rHNIHUN Hinimum Hinimum Ninimum

231. 2 219.54 198.15

I.

DESIGN r-1ETAL TEr·1PERATURE RESULTS: Nozzle Temp. w/o impact per Fig. UCS-66 Temp. at operating stress -124 Temp. w/o impact per UG-20(f) -2.

4-68

3,

lb. lb.

II

Pad 1.

1.

2

,.

-2.

-2.

F

Shell

,.

Nozzle f1D/olT Thickness Calc. per UCS-66 1 (b) , mN(tn,t,te) rolin. f.letal Temp. w/o impact per Fig. UCS·66 I·tin. f.letal Temp. at Req'd thk. (per UCS 66.1) -124 -20 '·Iin. Netal Temp. w/o impact per UG-20(f) WELD S1 ZE CALCULATIONS, NOZ7,LE NUI·IAER

psig

P

p

P F

Desc.· 0101 Nozzle C


Pr~ssurc

Vessel Design and Analysis - Seminar Notes

Ililisidc
OnO_Angl~ Nozzl~

Angles

COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------.------------------------ Page 26 Nozzle Analysis; 0101 Nozzle C ILem: 3 9:51a Sep 21,2000

fOlinimum thickness for nozzle/shell welds Tmin f·linimum thickness fOl pad/shell welds TminPad o

0.6250 0.7500

in. in.

Results Pel· UN-16.1, Requil·ed Thickness Actual Thickness 0.7 \·/0 Nozzle \\'eld 0.2500 /·lin per Code 0.2625 Pad weld 0.3750 = O.5·T1UNPAD 0.4375 = 0.7 • WP The PV Elite Program,


in. in.

Ie) 1989-2000 by CQADE Engineering Software

4-69

Ilillside ilnd Orr-Angle Nozzle Anglcs

4-70

Prcssurc Vesscl Design and

An~lysis *

SCl1linar Notes

Example Problcm I-A Simplc Drum


Chapter 5: Cones and Conical Sections

Cones combine all of the analysis techniques we learned for internal and external pressure.

GUIDELINES FOR CONES The required thickness of a cone for internal pressure (Paragraph UG-32) is based on the same design formulas used for cylinders, taking into account the inclination of the cone: I =

PD 2cosa(SE - O.6P)

0;

P

=

2SEtcosa D + 1.2/cosa

The external pressure calculations for cones (Paragraph UG-33) arc also based on the cylinder analysis, but with adjustments to the thickness and effective length to account for the inclination of the cone. The effective length for toriconica! sections is adjusted to include a fraction of the knuckle in the design length. Cones are required to have reinforcement at the large and small ends under internal pressure (Appendix 1-5) because of the tendency of the cone/cylinder junction to buckle under the radial load developed in the cone. The Code calculates the maximum angle below which buckling will not occur as a function of the design pressure and allowable stress. This ratio is used because it is an accurate indication of the diameter thickness ratio for the cylinder, and takes into account the strength of the material. This approach has the odd effect that when you increase the allowable stress you decrease the allowable cone angle. However, you will nonnally find that for a given thickness this effect is offset by the increased area available in the cone for reinforcement. (

Given that reinforcement is required, the required area is a function of the radius divided by the allowable stress. Area available in the shell within one decay length may be included in the area available for stiffening. Cones are required to have reinforcement at the large and small ends under external pressure (Appendix 1-7) because of the tendency to buckle under axial external loads. At both the large and small ends there are requirements for area ofreinforcement and moment of inertia of the reinforcement. The area of reinforcement is based on considerations similar to those described for internal pressure. The required moment of inertia of the reinforcement is a function of the strain in the ring at the cone/shell junction, which is in turn calculated using the Code materials chart from the stress in the ring. See the comments on stiffening rings in the external pressure section for further insight.

Cones <:Ind Conical Sections

5-1

Typical Geometry for a Simple Cone

Pressure Vessel Design and Analysis - Semiuar NoLes

TYPICAL GEOMETRY FOR A SIMPLE CONE The following four figures arc provided as examples of simple cones. r-----

FLAT HEAD

HALF" APEX ANGLE

CYLINDER CORROSION ALLOVANCE

CRO\v'N RADIUS

ACTUAL THICKNESS TORISPHERICAL HEAD

Figure 1··Typical Conical Vessel

-------

DIAMETER OF

LENGTH OF

SHALL END

SMAll END

THICKNESS SMALL END

------HALF

LENGTH OF

I I I

APEX

or

r-

ANGLE

COOE

THICKNESS CONE

or

I

~

_ .:u..;=::T

REINfURCING RIIlG

THlCKNESS Of

LENGTH Of LARGE END

THICKNESS Of

LARGE END

(

'WIDTH OF REINFORCING RING

Figure 2--Typical Geometry for a Reinforced Cone

tana

=

Opposite Adjacent

Opposite =

(D,-D S) .2

Adjacellt a

5-2

=

:-=

L

(D L - D s )/2

a 1"C lanl--"---;L-"--

Cones and Conical Sections

Pressure Vessel Dcsign and An
Typical Geometry for a Simplc Cone

(

L

Os \

I \

I

\

ex \

(

'

\17 \

/ I

I

"

·1 / \I

Figure 3--Half Apex Angle Calculation

(

Cones and Conical Sections·

(

5-3

Typical Geometry [or a Simple COlle

Pressure Vessel Design and Analysis - Seminar NoLes

v

:---...t~V__.~

.----=t~

Figure 4--Free Body Diagram

5-4

Cones and Conical SecLions

Whal do you need to know to perform cOile calculations?

WHAT DO YOU NEED TO KNOW TO PERFORM CONE CALCULATIONS? Dimensions of the Cone and the Cylinders at Either End The Code calculates areas available not only in the cone itself, but in the cylinders attaehed to either end of the Cone. If you have no cylinders (i.e. a cone with flanges on either end) you mu·st make an approximation concerning how much area is available at either end.

Dimensions of Transition Knuckles (if any exist) If the cone has a knuckle at the large end and a flare at the small end, area ofreinforcement and moment ofinertia calculations are not required. However, the thickness ofthe knuckle and flare must meet the requirements for toriconical sections in Appendix 1-4.

Half Apex Angle of the Cone

(

For internal pressure calculations without a knuckle or a flare, the half apex angle should not be greater than 30 degrees. When there is a knuckle or a flare the calculations can be used up to 60 degrees. For external pressure calculations the half apex angle must not be greater than 60 degrees. Some vessel, such (IS reb oilers, have two different angles associated with the conc. In other words, the cone is eccentric rather than concentric. For Code analysis, we recommend that

the half apex angle chosen be the larger of the two angles.

Axial Forces on the Cone The axial force can be either positive or negative. Positive values cause additional compression at the cone/cylinder intersection. Negative values reduce the eompression at the intersection. Determine the sign for your loading by drawing a free body diagram.

Width and Thickness of Cone Reinforcement Cone reinforcement is usually a simple flat plate cut in the shape of a donut. The width and thickness of the flat plate are used to calculate a combined m9ment of inertia for the plate/shell section. The Code rules specifY how to include the shell in the calculations, since the conical part ofthe shell is inclined.


5-5

What do you need lo know 10 perfonn cone calculalions?

5-6




What do you need

10

know

10

pcrronn cone calculations?


D-I02 Transition Cone: The vessel shown on page 74 has the following dimensions: Lower cylinder ID = 42 in. (1067 mm), thickness ~ 0.375 in. (9.5 mm), length ~ 42 in. (1067 mm) Upper cylinder ID = 18 in. (457 mm), thickness ~ 0.375 in.(9.5 mm), length ~ 18 in. (457 mm) Cone thickness ~ 0.375 in. (9.5 mm), length = 30 in.(762 mm) Design pressure ~ 125 psi (.862 N/mm 2), Design temperature ~ 300°F (149°C) All materials arc SS316 with zero corrosion allowance. Use higher allowable stresses and 100% radiography

(

Do this analysis for internal pressure only using the SHELL program. Then do the full analysis including full vacuum at 300°F using the CONICAL program. For the full analysis, include reinforcing rings of316SS on the shells at both the large and small ends. Suggested dimensions for the rings are 2 in. wide and 1/2 in. thick. Notes:

Questions: What is the half apex angle? What is the M.A.w.P. for the Cone? Are the reinforcing rings adequate?



(

5-7

What do you need 10 know to perform cone calculations'!


COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------------------- Page 10 Shell Analysis: DI02 Cone Item: 5 9:51a Sep 21,2000 Input Echo, Component

5,

Design Internal Pressure Temperature for Internal Pressure

Description: DI02 Cone P

Include Hydrostatic Head Components Material Specification Allowable Stress At Temperature Allowable Stress At Ambient Joint efficiency for Head Joint Inside

Diameter of Conical Head

Minimum Thickness of Pipe or Plate Corrosion Allowance Cone 1/2 Apex Angle Type of Element:

125.00 300.00

psig F

NO

psi psi

E

SA-240 316H 20000.00 20000.00 1.00

D

'2 .0000

in.

T CA

0.3750 0.0000

in. in.

ALPHA

21. 8000

S

SA

degrees

Conical Head or Shell

INTERNAL PRESSURE RESULTS, SHELL NUMBER 5, Dese.: 0102 Cone ASME Code, Section VIII, Division I, 1998, A-99 Thickness Due to Internal Pressure (TR): (P·(D+2*CA»/(2*Coslalpha)*(S*E-0.6*P») per Appendix 1-4 (e) (125.00+(42.0000+2*0.0000)/(2*0.9285*(20000.00*1.00-0.6*125.00)) 0.1419 in. Max. All. Working Pressure at Given Thickness (MAWP): (2*S*E*(T-CA)*COSA)/«(D+2*CA)+1.2*(T-CA)*COSA) per App 1-4(e) (2*20000.00*1.00·10.3750)*0.9285)/(42.0000+2*0.0000)+1.2*(0.3750)*0.9285 328.34 psig Maximum Allowable Pressure, New and Cold (MAPNC): (2*SA*E*T*Cos(alpha»/(D+I.2*T*Cos(alpha») per App 1-4(e) (2*20000.00*1.00·0.3750*0.9285)/{42.0000+1.2*0.3750*0.9285) 328.34 psig Actual stress at given pressure and thickness (Sact): (P*(ID+2*CA)+1.2*(T-CA)*Cos(alpha»)/(2*E*{T-CA)*Cos(alpha» (125.00*({42.0000+2*0.0000)+1.2*(0.3750)*0.9285)/(2*1.00*(0.3750)*0.9285 7614.16 psi SUMMARY OF INTERNAL PRESSURE RESULTS: Required Thickness plus Corrosion Allowance, Trca Actual Thickness as Given in Input Maximum Allowable Working Pressure MAWP Design Pressure as Given in Input P

5-8

0.1419 0.3750 328.34 125.00

in. in. psig psig

HYDROSTATIC TEST PRESSURES ( Measured at High Point J: Hydro. per UG-99(b); 1.3 * MAWP * Sa/S 426.84

psig



What do

yOlI

need to know

10

perfonn

COlle

calculations?

COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar -~--------------------------------~---- Page 11 Shell Analysis: 0102 Cone Item: 5 9:51a Sep 21,2000 Hydro. per UG-99{c) i 1.3 .. MAPNC WEIGHT Volume Weight Inside Weight

and VOLUME RESULTS, NO C.A. of Shell Component of Shell Component Volume of Component of Water in Component


426.84

VOLMET WMET VOLID WWAT

1073.6 300.6 22337.8 806.6

psig

in.**3 lb. in ..... 3 lb.


(


5-9

What do you need fo kllow lo perronn cone calculatIOns?


COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName : Seminar ---~----------------------------------- Page 27 Conical Analysis; D102 Cone Item: 1 9:s1a Sep 21,2000

Input Echo, Cone Item

1,

Description: D102 Cone

Design Internal Pressure PINT Temperature for Internal Pressure TEMP IN Design External Pressure PEXT Temperature for External Pressure TEMPEX Take Cone as Line of Support for External Pressure; Cone Material Cone Allowable Stress at Temperature Cone Allowable Stress At Ambient Longitudinal Joint Efficiency of Cone Circumferential Joint Efficiency of Cone Actual Thickness of Cone Corrosion Allowance for Cone Diameter Basis for Cone and Cylinders Diameter of Small End of Cone Diameter of Large End of Cone Half Apex Angle for Cone Axial Length of Cone

125.00 )00.00 15.00 300.00

SA-2-40 316H 20000.00 20000.00 1.0000 1.0000 0.3750 0.0000

BASIS DS DL ANGLE LC

ID 18.0000 42.0000 21.80 30.0000

SA-240 316H Small End Cylinder Material Small Cylinder Allowable Stress at Operating SAS 20000.00 20000.00 Small Cylinder Allowable Stress At Ambient SOS 1. 0000 Joint Efficiency of Small Cylinder ES Actual Thickness of Small Cylinder 0.3750 TS Corrosion Allowance for Small Cylinder CAS 0.0000 18.0000 Axial Length of Small Cylinder LS Large End Cylinder Material SA-240 316H Large Cylinder Allowable Stress at Operating SAL 20000.00 Large Cylinder Allowable Stress At Ambient 20000.00 SOL Joint Efficiency of Large Cylinder EL 1.0000 Actual Thickness of Large Cylinder TL 0.3756 Corrosion Allowance for Large Cylinder 0.0000 CAL Axial Length of Large Cylinder LL 42.0000 Type of Reinforcement at Large End of Cone:

F

psi psi

in. in.

in. in. degrees in.

psi psi in. in. in.

psi psi in. in. in.

Bar

Large End Reinforcing/Knuckle Material SA-240 316H Large Reinforcing/Knuckle Allowable, Operating 20000.00 Large Reinforcing/Knuckle Allowable, Ambient 20000.00 Location of Reinforcement at Large End of Cone: Shell Radial width of Reinforcing Bar{Large End) RWLB 2.0000 Axial Thickness of Reinforcing Bar RTLB 0.5000 Type of Reinforcement at Small End of Cone:

5-10

F

psig

Yes

SAC SOC EC ECC TC CAC

Small End Reinforcing/Knuckle Material Small Reinforcing/Knuckle Allowable, Operating Small Reinforcing/Knuckle Allowable, Ambient

psig

psi psi in. in.

Bar 316H 20000.00 20000.00

SA~240

psi psi



What do you need lo know lo perfonn cone calculalions?

COADE Engineering Software PVElite 4.00 Licensee; COADE Inc., Local White Lock FileName: Seminar ---------------------------------- Page 28 Conical Analysis; DI02 Cone Item; 1 9:51a Sep 21,2000 Location of Reinforcement at Small End of Cone: Radial width of Reinforcing Bar(Small End) RWSB Axial Thickness of Reinforcing Bar RTSB

Shell 2.0'000 0.5000

in. in.

INTERNAL PRESSURE RESULTS, CONE NUMBER I, Description: DI02 Cone ASME Code, Section VIII, Division 1, 1998, A-99 INTERNAL PRESSURE CALCULATIONS for CONE: Thickness Due to Internal Pressure (TR): (P*(D+2*CA))/(2*Cos(alpha)*(S*E-0.6*P)) per Appendix 1-4 (e) (125.00+(42.0000+2*0.0000»)/(2*0.9285*(20000.00*1.00-0.6*125.00» 0.1419 in.

(

Max. All. Working Pressure at Given Thickness (MAWP): (2*S*E*(T-CA)*COSA)/((D+2*CA)+1.2*(T-CA)*COSA) per App 1~4(e) (2*20000.00*1.00*(0.3750)*0.9285)/{(42.0000+2*0.0000)+1.2*(0.3750)*0.9285 328.34 psig INTERNAL PRESSURE CALCULATIONS for SMALL CYLINDER: Thickness Due to Internal Pressure (TR): (P*(D/2+CA))/(S*E-0.6*P) per UG-27 (c) (1) (125.00*(18.0000/2+0.0000)/(20000.00*1.00-0.6*125.00) 0.0565 in. Max. All. Working Pressure at Given Thickness (MAWP): (S*E*(T-CA)/«(D/2+CA)+0.6*'(T-CA» per UG-27 (c) (1) (20000.00*1.00*(0.3750»)/«(18.0000/2+0.0000)+0.6*0.3750) 813.01 psig INTERNAL PRESSURE CALCULATIONS for LARGE CYLINDER: Thickness Due to Internal Pressure (TR): (p* (D/2+CA» / (S*'E-O. 6*P) per UG-27 (c) (1) (125.00*(42.0000/2+0.0000»)/(20000.00*'1.00-0.6*125.00) 0.1317 in.

( Max. All. Working Pressure at Given Thickness (MAWP): (S*'E*'(T-CA»/(D/2+CA)+0.6*(T-CA» per UG-27 (c)(I) (20000.00*1.00*'{0.3750»/«(42.0000/2+0.0000)+0.6*0.3750) 353.36 psig SUMMARY of INT. PRESSURE RESULTS: Required Thickness plus CA Actual Given Thickness Max. All. Working Pressure Design Pressure as Given

Small Cyl 0.0565 0.3750 813.01 125.00

Cone 0.1419 0.3750 328.34 125.00

Large Cyl 0.1317 in. 0.3750 in. 353.36 psig 125.00 psig

EXTERNAL PRESSURE RESULTS, CONE NUMBER 1, Description: DI02 Cone ASME Code, Section VIII, Division I, 1998, A-99


5-11

What do you need lo know lo perfonn cone calculalions?


COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock PileName : Seminar --------------------------------------- Page 29 Conical Analysis: DI02 Cone Item: 1 9:51a Sep 21,2000 EXTERNAL PRESSURE CALCULATIONS for External Pressure Chart HA-2 Elastic Modulus for Material

j

CONE~

at

300.00 26600000.00

F psi

Results for Maximum Allowable External Pressure: 0.3750 in. Corroded Thickness of Shell TCA 42.7500 in. Outside Diameter of Shell OD Design Length of Cylinder or Cone 21.58 in. SLEN Diameter I Thickness Ratio 122.7805 DT Length I Diameter Ratio LD 0.5048 Geometry Factor, A f{DT,LD) 0.0019588 A Materials Factor, B, f(A, Chart) 10835.5977 psi B Maximum Allowable Working Pressure 117.67 psig EMAWP = (4*Bl/{3*DT) '" { 4 * 10835.598 )/( 3 122.7805 } '" 117.6690 Results for Required Thickness for External Pressure: Corroded Thickness of Shell 0.1158 in. TCA Outside Diameter of Shell 42.7500 in. OD Design Length of Cylinder or Cone SLEN 21.58 in. 397.5126 Diameter I Thickness Ratio DT Length I Diameter Ratio 0.5048 LD Geometry Factor, A f(DT,LD) 0.0003363 A Materials Factor, B, f(A, Chart} B 4472.1846 psi Maximum Allowable Working Pressure 15.00 psig EMAWP", (4*B)/{3*DT) '" ( 4 * 4472.185 )/( 3 * 397.5126 ) '" 15.0006 EXTERNAL PRESSURE CALCULATIONS for SMALL CYLINDER: External Pressure Chart HA-2 Elastic Modulus for Material

at

300.00 26600000.00

F psi

Results for Maximum Allowable External Pressure: Corroded Thickness of Shell 0.3750 in. TCA Outside Diameter of Shell 18.7500 in. OD Design Length of Cylinder or Cone SLEN 18.00 in. 50.0000 Diameter I Thickness Ratio DT Length I Diameter Ratio LD 0.9600 Geometry Factor, A f(DT,LD) 0.0038380 A Materials Factor, B, f(A, Chart) B 12093.H26 psi Maximum Allowable Working Pressure 322.48 psig EMAWP = {4*Bl/{3*DTl '" ( 4 * 12093.143 )/( 3 * 50.0000 ) '" 322.4838 Results for Required Thickness for External Pressure: _0.0618 in. Corroded Thickness of Shell TCA 18.7500 in. Outside Diameter of Shell OD Design Length of Cylinder or Cone 18.00 in. SLEN 303.4573 Diameter I Thickness Ratio DT Length I Diameter Ratio 0.9600 LD Geometry Factor, A f(DT,LD) 0.0002567 A Materials Factor, B, f(A, Chart) 3414.0022 psi B Maximum Allowable Working Pressure 15.00 psig EMAWP ~ (4*Bl/{3*DT) '" ( 4 * 3414.002 )/( 3 ~ 303.4573) '" 15.0005

5-12


(


\VhaL do you need Lo know 10 perfonn cone calculaLions?

CDADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName : Seminar --------------------------------------- Page 30 Conical Analysis : D102 Cone Item: 1 9: 51a Sep 21,2000

EXTERNAL PRESSURE CALCULATIONS for LARGE CYLINDER: External Pressure Chart HA-2 Elastic Modulus for Material

at

300.00 26600000.00

F psi

Results for Maximum Allowable External Pressure: Corroded Thickness of Shell TCA 0.3750 in. Outside Diameter of Shell 42.7500 in. OD Design Length of Cylinder or Cone 42.00 in. SLEN Diameter I Thickness Ratio DT 114.0000 Length I Diameter Ratio LD 0.9825 Geometry Factor, A f(DT,LD) A 0.0010881 Materials Factor, B, f(A, Chart) 9549.2256 psi B Maximum Allowable working Pressure 111.69 psig EMAWP = (4*B)/(3*DT) = ( 4 * 9549.226 l/( 3 * 114.0000 ) = 111.6869 Results for Required Thickness for External Pressure: corroded Thickness of Shell 0.1423 in. TCA Outside Diameter of Shell 42.7500 in. OD Design Length of Cylinder or Cone SLEN 42.00 in. Diameter I Thickness Ratio DT 300.5217 Length I Diameter Ratio LD 0.9825 Geometry Factor, A f(DT,LD) 0.0002542 A Materials Factor, 8, f(A, Chart) B 3381.0564 psi Maximum Allowable Working Pressure 15.00 psig EMAWP = (4*B)/(3*DT) = ( 4 * 3381.056 lI( 3 * 300.5217 ) = 15.0008

(

External Pressure Chart HA-2 Elastic Modulus for Large End Reinforcement

at

External Pressure Chart HA-2 Elastic Modulus for Small End Reinforcement

at

SUMMARY of EXT. PRESSURE RESULTS: Reqd. Thickness + CA Actual Given Thickness Max. All. Working Pressure Design Pressure as Given

300.00 26600000.00

psi

300.00 26600000.00

psi

Small Cyl Cone 0.0618 0.1158 0.3750 0.3750 322.48 117.67 15.00 15.00

F

F

Large Cyl 0.1423 in. 0.3750 in. 111.69 psig 15.00 psig

REINFORCEMENT CALCULATIONS for CONE / LARGE CYLINDER: REQUIRED AREA of REINFORCEMENT for LARGE END UNDER INTERNAL PRESSURE Large end ratio of pressure to allowable stress 0.00625 Large end max. half apex angle w/o reinforcement 25.500 degrees Large end actual half apex angle 21.800 degrees REQUIRED AREA of REINFORCEMENT for LARGE END UNDER EXTERNAL PRESSURE Large end ratio of pressure to allowable stress 0.00075 Large end max. half apex angle w/o reinforcement 1.875 degrees Large end actual half apex angle 21.800 degrees Area of Reinforcement Required in Large End Shell:


5-13

What do you need 10 know to perform cone calculations?


COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock

FileName: Seminar --------------------------------------- Page 31 Conical Analysis: 0102 Cone Item: 1 9:51a Sep 21,2000

ARLE

(RKLE*QL*RCLQ*TAN(Alpha)*/(SOL*EL))* (1.0-0.25* ((PEXT*RCLO-QL)/QL)* (Delta/Angle)

(1.0000 * 160.3125 ,., 21.3750 ,., 0.400/( 20000 * 1.00 )) * ( 1.0 - 0.25 * «( 15.00 ,., 21.3750 - 160.3125 ) I 160.3125 ) * ( 1.8750 I 21.8000 ) ARLE = 0.0671 in ARLE

AREA of REINFORCEMENT AVAILABLE in LARGE END SHELL:

AeL AeL AeL

.55*( Dl*ts )~ ,., ( ts + tc/CoslAlphal ) .55 * ( 4.2.750 * 0.375 )" * ( 0.375 + 0.375/ 0.928 ) 1.7152 in

SUMMARY of REINFORCEMENT AREA, LARGE END, EXTERNAL PRESSURE: Area of reinforcement required per App. 1-8(1) 0.0671 Area of reinforcement in shell per App. 1-8(2) 1.7152 Area of reinforcement in stiffening ring 1. 0000 Additional Area needed to satisfy requirements 0.0000 REQUIRED MOMENT of INERTIA, LARGE END, EXTERNAL PRESSURE: Area Available in Cone, Shell, and Reinforcement 14.87 Force per Unit Length on Shell/Cone Junction 466.68 Actual Buckling Stress associated with this Force 1005.92 Material Strain associated with this stress 0.000076

in in in in

in lb./in. psi

REQUIRED MOMENT of INERTIA, LARGE END, EXTERNAL PRESSURE: ISL AL * DCLO * DCLO * ATL / 10.9 ISL 0.000076 * 42.7500 * 42.7500 * 14.87 / 10.9 ISL 0.19 in.**4 AVAILABLE MOMENT of INERTIA, LARGE END, EXTERNAL PRESSURE: Area Centroid Ar*Ce Dist I Shi 0.826 0.0000 0.000 0.2931 0.010 Can 0.889 -0.4'104 -0.392 0.1335 0.070 Sec 1.000 1.1875 1.188 -0.8944 0.333 TOT 2.715 0.796 0.413 Centroid of Section 0.2931 Moment of Inertia SUMMARY of LARGE END INERTIA CALCULATIONS Available Moment of Inertia ( Large End ) Required Moment of Inertia ( Large End )

A

Ar*Di 2 0.1 0.5

0.8 1.3 1. 76

1.762 0.189

(

REINFORCEMENT CALCULATIONS for CONE / SMALL CYLINDER: REQUIRED AREA of REINFORCEMENT for SMALL END under INTERNAL PRESSURE Small end ratio of pressure to allowable stress 0.00625 Small end max. half apex angle w/o reinforcement 6.750 degrees Small end actual half apex angle 21.800 degrees REQUIRED AREA of REINFORCEMENT, SMALL END, INTERNAL: ARS ( RKS * QS *" ReSI / ( SAS *" ES ) ) *" {I DELTA/ANGLE ) *" TanAlpha 9.0000/ ( 20000 * 1.00 ) ) *" ARS ( 1.00 *" 562 ( 1.0 - 6.75/21.80) * 0.4000 0.0699 in ARS

5-14


(


\Vha! do you need 10 know lo perfonn cone calculalions?

COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------------------- Page 32 Conical Analysis: 0102 Cone Item: 1 9:51a Sep 21.2000

AREA of REINFORCEMENT AVAILABLE in SMALL END SHELL; Aes .78. (Rs.Ts) .... «Ts-t) + (Tc-Tr) /Cos (alpha) ) Aes .78+{ 9.000* 0.375) .... «( 0.375- 0.056 )+( 0.375- 0.061 )/ 0.93 ) Aes 0.9413 in SUMMARY of REINFORMENT AREA, SMALL END, INTERNAL PRESSURE: 0.0699 Area of reinforcement required per App. 1-5(3) 0.9413 Area of reinforcement in shell per App. 1-5(4) 1.0000 Area of reinforcement in stiffening ring 0.0000 Additional Area needed to satisfy requirements

A

in in in in

REQUIRED AREA of REINFORCEMENT for SMALL END under EXTERNAL PRESSURE Area of Reinforcement Required in Small End Shell: ARSE (RKSE· QS • RCSI * Tan( Alpha) / (SOS*ES) ARSE (1.0000·70.3125*9.3750*0.4000/(20000+1.00») ARSE 0.0132 in AREA of REINFORCEMENT AVAILABLE in SMALL END SHELL: Aes .SS"'(Ds"'ts)A*[(ts-t)+(tc-tr)/Cos(angle))] Aes .55*( 18.750* 0.375)"'*[( 0.375- 0.062)+( 0.375- 0.070)/ 0.928 ») Aes 0.9361 in SUMMARY of REINFORCEMENT AREA, SMALL END, EXTERNAL PRESSURE: 0.0132 Area of reinforcement required per App. 1-8(1) 0.9361 Area of reinforcement in shell per App. 1-8(2) 1.0000 Area of reinforcement in stiffening ring 0.0000 Additional Area needed to satisfy requirements REQUIRED MOMENT of INERTIA , SMALL END, EXTERNAL PRESSURE: 10.37 Area Available in Cone, Shell, and Reinforcem~nt 409.14 Force per Unit Length on Shell/Cone Junction 554.56 Actual Buckling Stress associated with this Force 0.000042 Material Strain associated with this stress

(

REQUIRED ISS ISS ISS

SMALL END, EXTERNAL PRESSURE':' I Dist Ar*Ce 0.6364 0.006 0.000 0.025 .0.3447 0.172 0.333 -0.5511 1.188 0.364 1. 359 Moment of Inertia 0.6364

SUMMARY of SMALL END INERTIA CALCULATIONS Available Moment of Inertia ( Small End ) Required Moment of Inertia ( Small End )


in lb./in. psi

MOMENT of INERTIA , SMALL END, EXTERNAL PRESSURE: AS • DCSO * DCSO * ATS / 10.9 0.000042 * 18.7500 * 18.7500 + 10.37 / 10.9 0.01 in.·*4

AVAILABLE MOMENT of INERTIA, Area Centroid Shl 0.547 0.0000 Can 0.589 0.2916 1.1875 Sec 1. 000 2.136 TOT Centroid of Section

(

in in in in

0.960 0.014

Ar*Di "'2 0.221 0.070 0.304 0.595 0.96

in**4 in**4

5-15

What do you need to know 10 perfonn cone calculations?



FileName; Seminar --------------------------------------- Page 33 Conical Analysis: DI02 Cone Item: 1 9:51a Sep 21,2000 Results for Discontinuity Stresses per Bednar p. 236 2nd Edition Stress Type

Stress

Allowable

Location

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

5122.08 -1823.70 5210.00

80000.00 Small Cyl. Long. -80000.00 Small Cyl. Long. Small End Tang. 30000.00 80000.00 Cone Longitudinal -80000.00 Cone Longitudinal 30000.00 Cone Tangential

Tensile Compres. Membrane Tensile Compres. Tensile

15693.52 -8631.02 367.91 15965.51 -8359.04 911.88

80000.00 Large Cyl. Long. -80000.00 Large Cyl. Long. 30000.00 Large End Tang. 80000.00 Cone Longitudinal -80000.00 Cone Longitudinal Cone Tangential 30000.00

The

5-16

5004 _14

Tensile Stress Compres. Stress Membrane Stress Tensile Stress Compres. Stress Tensile Stress Stress Stress Stress Stress Stress Stress

PV Elite Program,

-1941.64

4974.11

(e)



(


Chapter 6: Welded Flat Heads

Welded flat heads are analyzed as flat plates under unifonn loading.

GUIDELINES FOR WELDED FLAT HEADS The Code fonnula for welded flat plates is t ~ dJzCPlSE

Z, a factor accounting for non-circularity of the plate, is based on the ratio of the large and small diameters. Lowercase d is the small diameter, and z ranges from 1.0 for circular plates to 3.4 for plates with very large ratios of Did: Z ~ 34- 2.4d

.

D

C is an attachment factor that ranges from 0.17 for configurations that closely approximate fixed-end conditions to 0.50 for plates that are simply supported. A typical value of Cis 0.3. For the special CaSe of a formed flat head with a straight flange, C can go as low as 0.100. In this case, the Code has taken into account the smaller effective diameter of the flat plate due to the bend radius at the edges. The C factor has a multiplier of 1.5 built into it because the stress in a flat head is bending, not membrane, and the allowable stress in bending is nonnally 1.5 times higher than the allowable stress in tension. Notice how the Code factors compare to those in theoretical solutions of the flat plate problem:

Welded Flal Heads

6-1


Guidelines for Welded FlaL Heads

SIMPLY SUPPORTED

FIXED EDGE

Il1c!m' " ~ O.309p(dlt)'

Code (Max

C~O.33)

" ~ O.188p(dlt)'

Code (Min

C~O.13)

~ d'CPlSE

" ~ d'CPISE

"

SE ~ CP(dllJ'

SE ~ CP(dll)'

SE ~ O.33P(dll)'

SE ~ O.13P(dlt)'

1.5SE ~ O.50P(dlt)'

I.5SE ~ O.20P(dlt)'

(conservative)

(conservative)

The thickness is linear with the diameter, but is a function of the square root of the pressure (and the attachment factor). This means that as you increase the thickness, you dramatically increase the M.A.W.P.

6-2

Welded Flal Heads

Guidelines Cor Welded Flat Ileads

Pressure Vessel Design alld Analysis - Seminar NoLes

'.,ttm .,-o,rnl.. . PART UG _

L

',m'... - 2'.

,·3, ...."".

-1--"':\iJ=t

GENERAL REQ1..IIR.HMENTS

,t.- c.n\.r of

,. - 3.,-

."

-

c"

d

min.

2.'.10'

., __

' · " " n

m'n.

C .. 0.30 C" 0.20 w"O. 13

...

",.,,3/41".

C ......... O.20 ~1'

."

'. > I.in In. bul'-nota,. _ _•

. ---- _ C-O.33Ih-

0.17

'-'/2 In"

'mIn." 0..

.

~~ft.r;:"~ T . ,-3,

- o.376In.

100- ' . 0<

.

I"",

Iyt-T~

1b-2)

'''In.

d

-

..

-

,

-

,.,

C· 0.'3

s..

c., e.)

Fig. UW·13.:I.k-u"-t (0 CIII. doI_n. of -'dMI1o.m

..... Flo- UW·13.2 ..... ~ I.) 10 C111.

Indu....... for

I...... t

(

,............ lor doIteJl 01 ouuldoo _ _ Iolon

. :rr:m'-,~

1.215',

--'----

:¥.~

--j

r

d

I

e-O.3 IU.. Eq. (2) 0' Jell

IU", Eq, !:II 0' (6))

C",

'" ""II

c.,

C-O.:JO

c_,

,-,

C- 0.30

C-O.30

-~ u...,. _~ e_O..,6

IpJ

•

C-O.3

C mlo>. - 0.20

R'h'nl"'il

O.2~

~ dJ .

.----

e-o.:»",

c",

"0

_-.l'

l~-j ~.,t1

·C"O.33

C ..

c.,

'"

NOTE'Wh... plpo!lh.....,._

T_ UG-43 '0>

~~=.;;;;:~ :~-"',,-,-. , t ...... .+.. O.Br.mln. mln.r'l".or...

.. ;:t'~= c" 0.33

'"

J .. .,-

r.

Coo 0.;J3 ,~

FIG. UG-34 SOME ACCEPTABLE TYPES OF UNSTAYED FLAT HEAPS AND COVERS The Above IlIllStraUons A~ Diagrammatic: Only. Other Designs Thai Meet. Ute kequlremenb of UG-34 Are AoeepLable.

Figure 1--Typical Flat Head Atlachment Delails and C Values

(

Welded Flal Heads

6-3

Whm do

yOll

need

10

know la anf1lyzc welded Oal heads?


WHAT DO YOU NEED TO KNOW TO ANALYZE WELDED FLAT HEADS? Attachment Details Figure UG-34 shows typical head attachment delails, characteristic diameters, and values of the attachment factor.

Required and Actual Thickness of the Shell C is frequently a function of 01, the ratio of required to aclual shell thickness. C = 0.33 but not less than 0.20

tr ts

m=Ir

~

Required thickness of seamless shell (E = 1.0)

ts

~

Actual thickness of shell, exclusive of corrosion allowance

Large and Small Dimensions for the Head If the head is non-circular yon need to know both the large and the small dimensions of the head in order to calculate Z.

(

6-4

Welded Ftat Heads


What do you need

10

know to analyze welded nat heads?


D-102 Flat Head: The vessel shown on Page 5-2 and analyzed in the previous examplc also has an inserted flat head in the upper cylinder. The head is welded with a fillet weld above and below the head. What is the required thickness of this head? Upper cylinder 1D ~ 18 in. (457 mm), thickness ~ 0.375 in. (9.5 mm), length = 18 in. (457 mm) Design pressure ~ 125 psi .862 (N/mm2), Design temperature ~ 300°F (149°C) All materials arc SS316 with zero corrosion allowance. Use higher allowable stresses and 100% radiography.

( Do this analysis for internal pressure only using the SHELL program.

Notes:

Questions: How do you calculate C

(~F)?

What is the required thickness?

( What is the MAWP for External Pressure?


(

Welded Flat Heads

6-5

What do you need 10 know to analyze welded nal heads?


COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lack FileName: Seminar --------------------------------------- Page 12 Shell Analysis: 0102 Flat Item: 6 9:51a Sep 21,2000 Input Echo, Component

6,

Design Internal Pressure Temperature for Internal Pressure

Description; 0102 Flat p

125.00 300.00

Include Hydrostatic Head Components Material Specification Allowable Stress At Temperature Allowable Stress At Ambient Joint efficiency for Head Joint Diameter of Flat Head

psig F

NO

psi psi

E

SA.-240 316H 20000.00 20000.00 1.00

D

18.0000

in.

T

S

SA

Minimum Thickness of Pipe or Plate Corrosion Allowance

CA

1.0000 0.0000

in. in.

Attachment Factor Large Diameter of Flat Head

CF DL

0.2000 0_0000

in.

Type of Element:

Flat Head

(

INTERNAL PRESSURE RESULTS, SHELL NUMBER 6, Desc.; 0102 Flat ASME Code, Section VIII, Division 1, 1998, A-99 Thickness Due to Internal Pressure (TR): (D+2ei-CA) *SQRT(Zei-CFei-p/ (S*E)) per UG-34 Ie) (3) (18.0000+2*0.0000)*SQRT(1.00*0.20*125.00/120000.00*1.00)) 0.6364 in. Max. All. Working Pressure at Given Thickness (MAWP): (T-CA) / (D+2:lCA) ) "'2* ((S-E) / ICP*Z» per UG-34 (c) (3) (1.0000)/(18.0000+2*O.0000))*ei-2*((20000.00*1.00)/(0.20*1.00)) 308.64 psig Maximum Allowable Pressure, New and Cold (MAPNC): IT/D) "'2* ((S*E) / (Cp*Z») per UG-34 Ie) (3) (1.0000/18.0000)772*((20000.00*1.00)/(0.20*1.00») 308.64 psig Actual stress at given pressure and thickness (Sact): (Z*CP-P) / ( ( ( (T-CA) / (D+2*CA) ) .... 2) *E) (1.00*0.20-125.00)/«((1.0000)/(18.0000+2*0.0000)**2)*1.00) 8100.00 psi SUMMARY OF INTERNAL PRESSURE RESULTS: Required Thickness plus Corrosion Allowance, Trca Actual Thickness as Given in Input Maximum Allowable Working Pressure MAWP Design Pressure as Given in Input P

0.6364 1.0000 308.64 125.00

in. in. psig psig

HYDROSTATIC TEST PRESSURES ( Measured at High Point );

6-6

Welded Fla' Heads


\Vilal do you need 10 know Lo analyze welded Oal heads?

COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------------------- Page 13 Shell Analysis; 0102 Flat Item: 6 9:51a Sep 21,2000

Hydro. per Hydro. per WEIGHT

UG-99 (bl : 1.3 UG- 99 (c) i 1.3

*

•

MAWP It MAPNC

Sa/S

401.23 401.23

psig psig

and VOLUME RESULTS, NO C.A.

Volume of Shell Component

VOLMET

254.5

Weight of Shell Component

WMET

71.3


in.**3

lb.


(

(

Welded Flal Heads

(

6-7

What do yOll need to know to analyze welded flal heads?


(

6-8

Welded Flat Heads


Chapter 7: Half-Pipe Jackets

Half-pipe jackets are frequently used on vessels as heating or cooling coils around the outside of a cylindrical shel1, as shown in Figurc I.

GUIDELINES FOR HALF-PIPE JACKETS Appendix EE of the Code gives rules for the required thickness of the jacket and the required thickness of the shel1 under the combination of internal pressure and shel1 pressure.

R

Figure 1--Half-Pipe Jacket

The required thickness of the shen is first calculated using the normal rules ofUG-27(b). A half-pipe jacket introduces bending stresses in the shen in the longitudinal direction. Therefore, the anowable stress for the additional pressure in the jacket is F~1.5S-S',

where S' is the actual longitudinal stress due to internal pressure: S' ~ Pr/2t The Code does not givc an explicit formula for the actual bending stress in the shen due to internal pressure. Instead, it includes three graphs (EE-l, EE-2, and EE-3) that plot the stress factor (stress/unit pressure) as a function of the shen diameter. Thus the anowable pressure in the jacket as limited by shen bending is the anowable stress (F) divided by the stress factor (K). The required thickness of the half-pipe jacket is calculated from the standard internal pressure formula for a cylinder, but taking E = 0.85. The finet weld attaching the half-pipe jacket to the vessel shan have a throat thickness not less than the smaner of thc jacket or shell thickness. When a vessel is in cyclic service, a penetration weld plus a fil1et weld should be used to attach the jacket to the vessel.

Half-Pipe laekets

7-1

Guidelines for Half· Pipe Jackcls

7-2


Half-Pipe lackets

\


Guidelines for Half-Pipe Jackels

Pressure Vessel Design and Anal)'sis Seminar Componenl Design Problem

Half Pipe Jacket: What is the required thickness of a cylindrical shell subjected to an inside pressure of 190 psi (UI N/mm2)and a half-pipe jacket pressure of300 psi (2.06 N/mm2 )? The jacket is in non-cyclic service. J.D. of shell = 40 in. (1016 mm) Allowable stress of shell = 16,000 psi (110.316 N/mm 2) Joint efficiency of shell ~ 1.0 Half-pipe jacket is NPS 3 Allowable stress ofjacket material ~ 12,000 psi (83 N/mm 2) Corrosion allowance = 0.0 Using the HALFPIPE program, do this analysis for internal pressure only. Notes:

Questions: What is the required thickness of the shell? What is the required thickness of the jacket?

(


Half-Pipe Jackels

7-3


Guidelines for Half-Pipe Jackets

COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------------------- Page 103 Half Pipe Analysis; JACKETED Item: 1 9:51a Sep 21,2000 Input Echo, Half-Pipe Item

I,

Inside Diameter of Shell Thickness of Shell Internal Pressure in Shell Temperature for Internal Pressure Shell Section Material Shell Allowable Stress, Design Temp Shell Allowable Stress, Ambient Corrosion Allowance of Shell Joint efficiency for Shell Joint Nominal Pipe Size of Half-Pipe Jacket Minimum Thickness of Half-Pipe Jacket Design Pressure in Jacket Design Temperature for Jacket Jacket Material Name Jacket Allowable Stress, Design Temp Jacket Allowable Stress, Ambient Corrosion Allowance of Jacket

Desc~iption;

DIN TS P

S

SA CA E

NPS TJCK PI

Sl

SIA CM

JACKETED

40.0000 0.3125 190.00 300.00 SA-516 65 18600.00 18600.00 0.0000 1.00

in. in. psig

3.0000 1. 0000 300.00 300.00 SA-53 S/A 13'700.00 13700.00 0.0000

in. in. psig

F

psi psi in.

F

psi psi in.

Half-Pipe Jacket Results per ASME App. EE, 1998, A-99 SHELL THICKNESS CALCULATIONS: Required Thickness of Shell per UG-27 Eqn(l) (Includes CA); Tr (P '" R ) / ( S '" E - 0.6 ... P ) + ( CA + CAJ ) Tr (190.00'" 20.000 )/( 18600.00 '" 1.00 - 0.6 '" 190.00 ) + 0.000 Tr 0.2056 in. Required Thickness of Shell to Withstand Jacket Pressure: Trj = 0.2500 in.

PRESSURE CALCULATIONS FOR INPUT SHELL THICKNESS: Input Value of Shell Thickness: Ts = 0.3125 in. Chart Used to Find the K-Factor: FIG. EE 2 K-Factor Read from Chart: K = 46.5000 Longitudinal Stress in Shell due to Internal Pressure (Includes CAl: SPrime (P '" R ) / ( 2 '" Ts ) Sprime (190.0000 * 20.0000 ) / ( 2 ... 0.3125 ) Sprime 6080.0000 psi Permissible Jacket Pressure per Appendix EE-l, Equation {Il: Pprime 1.5 * S - Sprime ). I K 0=

7-4

(

Half-Pipe Jackels


Pressure Vessel Design and Analysis - Scminar Nolcs

COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName; Seminar --------------------------------------- Page 104 Half Pipe Analysis : J~CKETED Item: 1 9:51a Sep 21,2000 Pprime Pprime

( 1.5 ~ 18600.00 469.2473 psig

6080.0000 ) / 46.50

HALF-PIPE JACKET THICKNESS CALCULATIONS: Input Half-Pipe Jacket Thickness: Tj '" 1.0000 in. Req'd Half-Pipe Jacket Thickness per App. EE-l, Eqn. (2) {Includes CAl: T (PI '" R ) / ( . 85 ~ SI .6 '" PI 1 + CAJ T {300. 0000 * 0.7500 1 / ( .85 '" 13700.00 - .6 ,. 300.0000 ) + 0.0000 T 0.0196 in.

MINIMUM FILLET WELD SIZE CALCULATIONS: Minimum Fillet Weld Size (Based on Shell Thickness) : Fillet Ts· 1.414 Fillet 0.3125 * 1.414 Fillet 0.4419 in.

SUMMARY OF RESULTS; Input Thickness of Shell Req.d Thickness of Shell due to Internal P. Req.d Thickness of Shell due to Jacket P. Pressure Used for Jacket Design M.A.W.P. of Jacket for Input Thickness M.A.W.P. of Jacket for Required Thickness

0.3125 0.2056 0.2500 300.0000 469.2473 338.3333

in. in. in. psig psig psig

Input Thickness of Half-Pipe Jacket Required Thickness of Half-Pipe Jacket

1.0000 0.0196

in. in.

Minimum Acceptable Fillet Weld Size

0.4419

in.


(e) 1989-2000 by COADE Engineering Software

(

Half-Pipe Jackets

7-5



---- ----

r

----

-- - - - - - - - - -

--

CHANNEL COVER

:

I

CHANNEL _ TIllCKNESS

TIlB ESIDE CORROSION ALLOWANCE

I

I

II

/

"r-I

I

MEAN GASKET DIA.

I-J r--

I

I I I

:

,

I

-

I

I IJJ~l~~ I_~~_II

\,

SHELLSIDE CO RROSION _ ALLOWANCE

II I STATIONARY TUBEFRONT

-.<

I

II

SHELL GASKET DIA.

SHELL TIDCKNESS--

C HANNEL

I

-

l' I

CHANNEL ill -OR-

,,

I I I I I

\,l--W-J X /\ L

f--

DOLT CIRCLE DIAMETER

•

I f-- f--

SHELL

I

NOTRA FIXED TUBESHEET EXCIIANGERHASTWO STATIONARY ENDS (NOFLOATINO TIlBESIIEEl)

,

,

,

I

! ! ! I

,-

:

I

I

I

\

SHELL FLANGE

C~,

..J l

'~

\~

/1"/

.

FLOATING TUBESHEET FLOATINGHEAD

SHELL COVER

Figure 2--Typical Geometry for a TEMA Heat Exchanger

7-6

Half-Pipe Jackets

Pressure Vessel Design Clnd Analysis - Seminar Noles

Fl,mgc Design and Analysis

FLANGE DESIGN AND ANALYSIS Flange design rules were first published by Tay lor Forge in 1937. These nt les were included in Ihe Code in 1942 and are essenlially unchanged. They are found in Appendix 2. The Taylor Forge bulletin is also still available, and is still one oflhe most uscfullools for flange analysis. The flange design sheels in Ihis section arc modeled after Ihe Taylor Forge analysis. .

Gaskets A key eomponenl in Ihe design and practical applicalion of flanges is Ihe gasket. There are many differenllypes of gaskel malerials, useful in differenl services and al differenllemperalures. The Code defines Iwo gasket factors: y m

=

Gasket seating stress, minimum stress 10 seat the gasket Multiple of pressure needed as a slress on the gasket surface to keep it from leaking

Unfortunately, these factors are not reliable as measures of gasket behavior. First of all, they arc not constant; they vary with such Ihings as applied load, temperature, leakage rate, and size. Second, they do not by themselves adequately predict leakage rate for a flanged joint. The Pressure Vessel Research Council is working on improving these design factors, and has published preliminary findings for the past two years at the annual piping

(

and pressure vessel conference.

Gasket types and their associated m and y factors are found in Table 2-5.1, shown on the following two pages. Some typical dimensions for gaskets are also shown on Ihe following two pages. In some cases additional gaskel material is used 10 seal partilions. These partition gaskets are used for channel flanges and also for channel covers, which seal a channel partition. Include Ihe area of Ihe gasketed partitions in caleulalions for seating forces and overall loads.

(

Half-Pipe Jackets

7-7

Flange Design

Analysis


Gasket Materials and Gasket Factors 1998 SECTION VIn

T.ble 2-5.1

OlVlSION

TABLE 2·5.1 GASKET MATERIALS AND CONTACT FACINGSI Gasket Factors m ror Operaling Condilions and Minimum Design Seating Stress y

MIn. Dellgn SeaUng

Fa<:lng Skelth ;lAc! Column

,,,

SIIHST,

Gasket Malerlal SeIf-eMrlllllng typei fO rinlls, metallic, elaslo~, other gul(el lypes tOMJdtrtd n stlf'Malll19J

Elastomer$ wilhoYl fabric or high perce"l of aweslos fiber: B~IO¥t 75A SOOlt Dwomet.er 751>. or higher Sl'>ore Durometer

o

0.50 1.00

Skekhes

in Table 2·5.2

o

o

lllll,(lbl,llel,(ldl,

1-41,C5l; Colu,.", II

200

Asbul.OS with roltablr birder for- operalillg cOllditlOfJ':

V.ln. Wcic

2.00

Vl6 In. thick V» In.. Ihlck

2.75

11:>00 )100

3.50

6500

l.25

'00

Elastomer$ WfUl canon fabrk IMtrt.lon

( la),(lbl,llel,Cld),

141,151;

CoIUIIVI II

(lill,() bl,nel,( ldl,

141,(51; Celu,.", II

flastonms with .uOeslos labrlc Instr1ron (willi or without wire relnfon:emtn(J:

~

2.25

2200

2."

2.'100

~

2.75

)700

~

1.75

aoo

Carbon

2.50

10,000

Stalo1JtsS, MOt'leI, and nlckel·bast all0)'5

'.00

10,00<1

2-ply

Ve9tlable fiber

~

(la),Ubl,lld,CldJ, (4),(51; Column II

(laJ,flb),(leI,nd), f4mli Colllll1l'1 II.

Sprral-wolHld metal, asbestos filled:

Conug.altd rnffiJJ, a:sbestos Inserted, or torrugaled metal, Jacketed asbe110$ filltd: Soft ah,rnll'l\.ltrl Soft topper 01' bran

2.50

2900

2.75 '.00 3.25

3100

Iron or ~ Jlrer Monel or 4%-0% ,hr(llM Stalnltss steels and nrckel-base aJJOI'S

3.50

<500

nlll,lIbl; Column II

na),( IbJi CollII'M II

5500 60500

Figure 3--Gasket Materials and Conlact Facings

7-8

Half-Pipe Jackets

\


Flange Design and Analysis

Gasket Materials and Gasket Factors (Continued) APPENDIX 2 -

MANDATORY

TABlE 2·5.1ICONT'D} GASKET MATERIAlS AND CONTACT FACINGSl Gasket Factors n1 for Operallng Condition' and Minimum Design Seating Stress y MI". Des19" Gasht Factor Gas\;et MaterIal

•

Sealing

,.,

Fae!"; SktlCh and ColllM

Stress .r.-

In Table 2·52

Skelchu

Corh/96kd mm.l:

)700 4500 5500 6500 7600

~

5500 6500 7600

4%-{)% chrome Stainless sleels III>d nlchl-base alloys

3.25 3.50 3.75 3.50 3.75

==
3.75

9000

GroO'o'to mel..l: Soft alurnlm,m Soft copptor' or brass Iron or soft metal Monel or 4%-6% chrome Slalnlns sleels and "lckel-base alloys

3.25 3.50 3.15 3.15 4.25

5500 6500

Soft ahllllinurn Soft e~r or btan lroo or 50ft stffl MON!I or 4%-6% chrome Stainless fleels and I\kkel-ba!oe 11110)"5

Flill metal, Jacketed asbestos filled: Soft Ill,I1linl.m SolI copper or btau tron or soft Sleel

"-,

2.75 ).00 3.25

).50 ).75

BOOO 90(>0

noo

'000 10,100

Ual,Ubl,llcl,lldl; Colt.m11\ II

~

~

llal,f1b),Ue),l Ild)'!;12J 1;

Column II

(U),() bJ,()cl,()d), m,U); Column II

Solid Rat mnal;

Soft alurnlt'lUm Soft copper or tnss Iron Of'" 50ft $letl MOM! or 4%-6% worne $liI.1n1m JI.~ls aIId nickel-base alloys

'.00 '.00

8800 n,OOO 18,000 21,800

6.50

26,000

5.50 6.00 6.50

18,000 21,800 26,000

4.75

5."

(laJ,()bl,llc),11 !fl, f21,(3),14l,{5J; COrllllVl1

RIII9 ./OInt: 1rOll or soft steel Monet or 4%-6% chrome Slalnless s1~rs and nlckel-base allO)'S'

(61; Colum" I

HOTES: H) ThIs T~e gives a list of manr convnonlyllsed 9l»:et matfrlills -.nd conLact liKings V
(

Figure 4--Gasket Materials and Contact Facings (Continued)

Half-Pipe Jackels

(

7-9


Fhmgc Design and Analysis

USEFUL TECHNICAL DATA BOLTING DATA FOR ASME/ANSI B 16.5 & BS 1560 FLANGES

Class 150

N... p~,

RAN" NO

BOLT B.r

RAN" NO

BOLT B.r

Sil:e

llA

0' BOLll

llA

llA

llA

0' BOLll

llA

IIA

llA

\\ y, % 1 IV.

3% 3Y, 3Y. 4V. 4% 5 6 7 7V, 8Y, 9 10

4 4 4 4 4 4 4 4 4 8 8 8 8 8 12 12 12 16 16 20 20

y, V, V, V, V, V,

2\\ 2% 2% 3V. 3V, 3Y, 4% 5V, 6 7 7% 8% 9% 11% 14\\ 17 18% 21V. 22% 25 29V,

3% 3% 4% 4% 5\\ 6V. 6'(, 7% 8\\ 9 10 II 12% 15 17V, 20% 23 25V, 28 3OY, 36

4 4 4 4 4 4 8 8 8 8 8 8 12 12 16 16 20 20 24 24 24

Yi V, % % % %

2\\ 2% 3\\ 3V, 3Y. 4V, 5 5Y, 6% 7V. 7Y. 9\\ 10% 13 15\\ 17%

3% 3% 4% 4% 5V. 6V. 6% 7V, 8V. 9 10 II 12Yi 15 17% 2OY, 23 25V, 28 3OY, 36

W,

2 2% 3 3Y, 4 5 6 8 10 12 14 16 18 20 24

11

nv,

16 19 21 23V, 25 27V, 32

% % % % % % % % Y, % I I IV. 1% 1%

Clm 600

Class 400

ClassJOO

RAN" NO OF

BlICY B.r llA llA

RAN" NO BIA

BOLT B.r OF llA llA BOLT<

3% 3% 4% 4% 5\\ 6V. 6% 7\\ 8\\ 9 10% 13 14 16V, 20 22 23% 27 29V. 32 J7

4 4 4 4 4 4 8 8 8 8 8 8 12 12 16 20 20 20 20 24 24

BOW

% % % % % % % %

1 1% 1% 1\\ 1\\ IV.

20v.

22V, 24% 27 W, 32

4 4 4 4 4 4 8 8 8 8 8 8 12 12 16 16 20 20 24 24 24

Yi Y, % % % % % % % % Y, % % I IV. 1\\ 1\\ Pis 1%

IY, IV.

2\\ 2% 3\\ 3V, 3% 4% 5 5% 6% 7V. 7% 9V. 10'/, 13 15\\ 17% 20\\ 22% 24% 27 32

y, V, % % % % % % % Y. Y, I 1 IV. 1\\ IV. 1% W,

1% 1% 1%

2\\ 2% 3\\ 3V, 3Y. 4% 5 5% 6% 7\\ 8V, 10% llY, 13% 17 19V. 20% 23% 25% 28V, 33

.

. Figure 5--Bolling Data

7-10

Half-Pipe Jackels

\

Flange Design ,llld Analysis

Pressure Vessel Design rlllrJ Analysis - Seminar Notes

USEFUL TECHNICAL DATA BOLTING DATA FOR ASME/ANSI B 16.5 & BS 1560 FLANGES -

'~~'._-

NollliAal Pipt Siu

v, 'I. Ph

1 lYi

2 2Y, 3 4 5 6 8

-- - -

---"'-~.~

n",,,,

'"

4% 5Va Sl'a 6Y4 7 81'2

9% 9Y2 IlVl 13¥.& 15 181'2

21Y2 24

10 12 14

2jy.,

16

18 20 24

NO

0'

."'" 4 4 4 4 4 8 8 8 8 8 12 12

.ou H .u. .u. * 'I. Y. V. 1 V. 1 V. JY. llh lY. 1%

27%

16 20 20 20

1%

1%

31 33% 41

20 20 20

1% 2 2V,

m Iv.i

.

..

""'''' ou

NO

3v.

4%

" """ 4

3Y,

5Yo 5V! 6lh 7 8Y, 9% 10\12 12Y4 14% 15\12 19

4 4 4 4 8 8 8 8 8 12 12

23 261'2 29112 321'2

12 16 16 16

36

16 16 16

4 4% 4% W, 7Y, 7Y, 9!1.l 11 12 Y, 15 Y, 18 Y, 21 22 24

"',

.u. * * Y. V.

__'N"' ___

•.e.

ou. 31' 31'2

n ...",

.u

5!h 5Y, 6Y.i

NO

0'

""" 4

12

l!1.i

8 9Y2

jy,

llY2

16Yl

1% 1%

121'2 15lh

19 21%

4 4 4 4 8 8 8 8 8 8 12

1% 2 2y.,

19 22Y2 25 27%

261'2 30

-

I

Y. 1

tVa

21'2

-

-

-~

_._'-~

Class 2500

Class 1500

Class 900

4 4% 41's 61'2 71'2

7Y4 8 9'h lOY:!

14

."', .u. 'I. 'I. Y.

-

H

,u. 3Y2 3% 4y"

I

51's 5* 6'1.

I

tva

tVa

7%

lY.i

9

lY1 1% 2 2

lQJ'A 120/4 141'1 17Y.&

12 12

2Y, 2%

21Y4 24%

-

-

-

-

-

I'

27 29 Y, 35 Y,

38% 46

2%

301'2

3 314

32Jh

-

39

-

-

Figure 6-- Bolling Data Continued

(

Half-Pipe Jackels

7-11



Other Gasket Types The following gasket types arc not pictured in the Code tables, but are also acceptable gasket materials and geometries: ~nprel8ullzed ~res~rized

"

I

'\

.

\'enled Venl7

I1IJ ~ 1)1IJ rrorJto 1/· ; ~. Y 1/ JI

JJ

I

'I

J1

C-Rlng

I

J

~ ,i

lens

0'11.

Doubl, Cone

Brldg,"'n

Figure 7-Olher Gasket Types

Facing Sketches There are several types of flange facing types. The flange facing is simply the metal surface against which the gasket seats. The Code shows several sketches of typical facing types in Table 2-5.2. This table also shows how to calculate the basic and effective gasket seating widths and the diameter of the gasket load reaction. Two common mistakes in gasket design (and even ill some computer programs we have seen) are, (I) to use the basic gasket seating width instead of the effective gasket seating width in the flange calculations, and, (2) 10 use the mean diameter of the gaskel as the diameter of the gasket load reaction in all cases. A special type of gaskel geometry, which is not included in the Code sketches, nor even in the Code design rules, is the flange with a flat face and a gasket that extend from the ID of the flange to the OD, beyond the bolt circle. The gaskets used with this type of flange are

7-12

Half-Pipe Jackets

(



usually quite soft. These flanges can be analyzed using the Taylor Forge calculation sheets. 1998 SECTlON

Table 2-5.2

vm -

DMSION 1

TABLE 2-5.2

EFFECTIVE GASKET WIDTHl BAsic Cuket Seating WkIlh bQ Facing Skelch IEu99tr.. ted) "«'UUe cul'

{lal

=

~

lIb)

'(

fu

¢N'<4

.,"

'7t2k~'

Ste Note III

1/... III. nubbin

'"

II... In. IIIJbbln

t~

"W""

¢',:::[t.., ~

,<) See Note (II

w+ r (w+ N --,--: -.- m")

. . T ( -.-m" .. " ) -,-;

W<" ~

•

~

wSN/l

wS Nn

•"

'",

!."-

7"

...

N

!1'"

,"

=

•'"'5'& => r" T

,2l

"-,

J>'tN:J';; r

:'f:g0i';' '" t'ii jI"

lId)

Co/umn II

'«UtUC" ,

(I)

Uel

Column I "'"

~

= 7""~~

Set! Hote

(

vct/ u

f

.

cry

=

•

•

' tN'J

16

"\

'" See NtJtt UI

= ~ZI"'"

•

~

•

-j

'M

'",

N

1--

•

Effecti~ GMktt

...

Sealing Wi6lh. b

0.5 ~ when b. > 1/. In. LoutJon. of Gasll:tlload Reaction

b - b" when b,.s: ".t. in.; b

=

G

~ -I~ hG "

(

-1

G..,kCl

F~

NOTE: (11 Whefl!

~lTaUom do not e~utd V"" In. c1tpthand '/jl In. wldlh 5paCln9, sketches UbI ~ Cld) ~I be used. (2) Tile: gaSh! factlm listed lmly apply 10 flanged )
Figure B--Facing Sketches and Gasket Seating Calculations

Flange Types There are essentially only two categories of flanges for purposes of analysis. These are integral type flanges, where the flange and the vessel to which it is attached behave as a unit, and loose types, where the flange and the vessel do not behave as a unit. Within these categories, however, there are several additional subdivisions.

Half-Pipe Jackels

7-13



Weld Neck FJanges-These have a hub that is butt-welded to the vesseL Slip-on Flanges-These have hubs, and are normally analyzed as loose type flanges. To qualify as integral type flanges they required a penetration weld between the flange and the vesseL Ring Flanges-These do not have a hub, though they frequently have a weld at the back of the flange. They are normally analyzed as loose, but may be analyzed as integral if a penelration weld is used between the flange and the vessel. Lap Joint Flanges-These flanges mayor may not have a hub, but they are completely disconnected from the vessel, bearing only on a vessel "lap." They are always analyzed as loose. Reverse geometry flange-Here the gasket seat is on the inside of the shell diameter. These use integral flange rules, which arc suitably modified for the reversal of the bending moments. Sec Appendix 2-13. Loose-type flanges, especially Jap joints, may be split. A split is used when it is required to have the flange completely removable from the vessel. If the flange is split into two pieces by a single split, the design moment for the flange is mnltiplied by 2.0. If the flange consists of two separate split rings, each ring shall be designed as ifit were a solid flange (without splits) using 0.75 times the design moment. The pair of rings shall be assembled so that the splits in one ring shall be 90 deg. from the splits in the other. 2-5

1998 SECTION

vm -

1-5

DIVISION l

Guk~t

Slope

1:3 ima>l.l

Whet'& hUb lIop!! adiillC81t (0 f111'1ge ex«Bdl I :J,

151

Slope 1;3 (max.)

L.....ILL:;:r I.S go (min.)

UI. tJ
181 0.2590 but nOt Ius tharI 114 In .. 1hl minimum l6bl ...... - --1--,i----''+I'-' for lither leg. This ,"';!Id mil,!, be mee/'llned til 8 cOUM!r ,..Jius os petmilled in

sketch (51 in which ease 91 ~ 9 0 GENERAL NOTES ILoow lind Integ,.l Type Fr.ngoell: fll FHle\ flIdiU$ r to be II lean 0.25.91 bUI mit I~$ than 3/16 in.

~

"-j;""::'-t~::!~~t:J~

~T G

i . - . HD

gQ

,'.

8

.,12

lbiFiKin91hlcknttWl or llfOQYll'dlePttll9'"eall!r lIlan 1116 In. ,hall be In uons of Ihl required minimum 1l9nte thlckn.eu I; IhoI'lI ~uIIIO or Ins lhan 1116 in. m.lIV be intluded In Ihl overall fiar>Qt1hlckOffi.

Figure 9--ASME Code Flange Types

7-14

Half-Pipe Jackets


2-5

Fl
APPENDIX 2 -

MANDATORY Full pene1f81lon weld,

r:+:~

A

, D

lingl.. Of" double. The full penc'UlItion _ld may biI throVOh the hI>

lIJI« through thl

C

t

wall (tn ),

I.

11~~=*_"

~~=-==-~·HD T

"

,

To be taken IIl1 midpoint or COotlel bl!tweln flatlge end IIlP lndepeodlnt 01 gallr.cl location

min,"0.7r;

This Mk:I may be meehined to. com".. radius

ScrlM'Qd F.~ Wilfl Hub

to luit It5'ldmIlap jolnl "eng..

III

u Selrwtd

FIa~

WiI'hout HUb

(

121

Jt~ m ....

r;"

1211 (Not. Ill]

• 0_7 t:

~

1/4 in

131lNoie (2)1

~~ _0" (_

t:

min. ~

.....j

Q,7r;

I-

\

1/2'

...

(tneX.I

+ 1f4 In.

11I11N01.121)

14J

INol1

1211

{4111 (Noli 12)1

NOTES (LOO$8 Type FI.~I: III For hubtl~ 6 de9- or lnl, Ulego • "I' 12) Loedinfj end dlmtn,lOllllor Skttctlel {2,1. (31,13.1. {41, and (4a) tlO1lho.... n Ir. lIle SIlITle U 101 sketch (21. Lo~T.,.~FI ..~

Figure 10-·ASME Code Flange Types (Continued)

(

Half-Pipe Jackels

(

7-15



APPENDIX 2 -

ng.

MANDATORY

24

,~~~,;;~ t}. .~ 101

FilII PIIl'It1l1liOn and backchlp [lte Fig. UW·13.2 ,ketUt.. Iml .ocI 1111 and 00.93 Idl UJ I

110)

(11)

GENERAL NOTES (Optional Typa Flenges), fal Optiona. type flanges m.y b. calellialed u ,ither loole or integ.allype. S" 2-4 Ibllo.dlngs .nd dimenSIO", nol1hown In sk,tchlll UI).18.).19j.1ge).1101, .nd 110011 are Ihe "me u ,hown III Iketch (21 wtlen lhe nang. II clllcllial.d 81 8 loose type lI.nga and as ehown ill ,k.leh (7) when lhe lI,nge Is nlellialed II en Integrellyp. lIeng'_ (el The groove and lill.1 weld, behyeen Ih. lIanga blck lacI 'nd the shell given in sketch Ie) also apply 10 ,kelches (8.). (91.190).110). lind (1081_ Opllof\ll Typ. FI."V'<

~gl""

,-%In.

FOI illttglllllY

4-llllide di!lmeler )

!-31'''" Subtype 1.1

r.infOl'm 1l0UI".

~In. -nut

he19ht .. 1f4 In:. SllblyPt lei

Sllbl""':lbJ

t

(,-, I+SulMvpe fdl .../"lJ I.~ Nul f10pdllmetH"

!1I16In.

All Oillet" dll.il. 'I shOWn In shIeh 1121

0

1121 GENERAL NOTE If...,..Wilh NIlI SI0PtJ: For wbl:yP8I fd Ind fbl"o II the Ihicltn.u ollhe hub.1 the ...... 11 .nd. For Jubtypet 1(11nd Idl. 90 .91' FlingM \'tlln Hut Stops

FIG. 2-4 TVPES OF flANGES (CONT'D)

Figure 11--ASME Code Flange Types (Continued)

Flange Behavior The flange design rules incorporated in the Code were based on a paper written in 1937 by Waters, Westrom, Rossheim, and Williams. This paper assumes linear behavior of all the components-flanges, bolts, and gaskets. The Code rules do not take into account external piping loads or Ihennal stresses due to thermal gradients. In most cases, the flange rules result in a good flange, except Ihat the bolts must be stressed higher than the Code allowable to get a leak-free seat.

7-16

Half-Pipe Jackets



The flange model assumes that the flange can be modeled as stiff clements (the flange and hub) and springs (the bolls and gaskets). The initial boll load compresses Ihe gasket. This load needs to be high enough to seat (deform) the gasket, and high enough to seal even when pressure is applied. The pressure load adds to the boll load and unloads the gasket. When pressure is applied, Ihe flange rolates around the gasket, thus decreasing the bolt load. However, the pressure end load tends to unload the gasket and load the bolts. The degree of decrease of the gasket load is a function of the gasket stiffness. Real configurations can go either way. If the gasket is stiff relative to the bolt, the gasket unloads preferentially, and the bolt load stays relatively constant. If the bolt is stiff relative to the gasket, the boll loads preferentially, and the gasket load slays relatively constant. The Code assumes that if excess balling is provided a propensity exists for prcloading the bolls in excess of the calculated required bolt load, and it makes an approximation to take this into consideration so as not to overstress the flanges. The Code also recognizes (Appendix S) that it is usually advantageous to load the bolt to a higher value than the allowable stress. Typical values range from 40,000 to 50,000 psi, or up to 90 percent of yield. There are several methods of achieving uniform and high boll loads. From simplest to most complicated these include: Cross tightening Torque wrenching Pneumatic Tensioning Pneumatic Tensioning with ultrasonics Pneumatic Tensioning with strain gaging

HalfRPipe Jackets

7-17

Flange Design (lnd Analysis

The following table shows typical torque values for various bolt sizes.

I TORQUE REQUIRED TO PRODUCE BOLT STRESS

I

The torque or turning effort required to produce a certain stress in bolting is'dependent upon a number of condilions. some of which are: I. Diameter of Bolt. 2. Type and number of threads on bolt. 3. Materialofbo!t. 4. Condition OfllUt bearing surfaces. 5. Lubrication of bolt threads and nut bearing surfaces. Generally. standard Flexitallic spiral-wound gaskets will require that standard AS ME flange bolting be stressed to 30,000 psi for proper gasket seating. However, it is the users responsibility to follow ASME Code calculations, and to ensure thaI sufficient pre-load is applied to withstand infernal pressure. properly seat the gaskel, and compensate for the effects of bolt relaxation. The table below reflects the results of many tests to determine the relation between torque and bolt stress. Values are based on sleel bolting welliubricaled with a graphite and oil mixture.

(

TORQUE DATA FOR USE WITH ALLOY STEEL STUD BOLTS Load in Pounds on SLud Bolts when Torque Loads are applied NO.IIN.\L

.....ffi

,,~

ornou:.ws

OFBOLT (bo<1I<.»

Ow.
(porLod)

(bI<1I<.)

-

ATROOr 0'

n ...... St·""\

'"

5/16

318 7/16 1/2

,n6

'" 314

7" 1

1· 118 1·1/4 1·3/8 1·111 1· 518 1· 3/4 1·718 2 2·114 2·1/1 2-314 3

2.

" l' 16

13

II 11 10

, • ••• •• ,•

,• ,,,

.185 .240 .294

.34'

.027 .<>1,

.06'

.093

.400

.126

..154 .507

.162

.202

.620

.302

.731

.'119

.."

.963 1.088 1.213 1.338 1.463

30.000 PSI C..." .....

T,,,.. ,~.

,

•

12 2. 30

"60 100 160

.551

24'

.728

3"

92' 1.155 1.405 L.6W 1.980 2.304 2.652

...

..,

To"...

.1. 1350 2<>1. 2790 3780

6

'860 6060 9060 12570 115530

"""

45POOPSI

"

12

" "

30

C. . .

""Io.

u.

1215 2025 3060 4185 5670

r~u.

•

16 2'

,.

60

1020 1200 1650

'3680 55740 69300 84300 100800

3000 '000 4400 6360

118SOO 138240 ]59120 205380 257520 315540 379440

000

1100

'Q400

"'"'"

69120 79560 102690

225. 3000 3300 4770

89100 103680 110340 154035

128760 157770 189720

6600 "00 11580

193140

'800

2366"

11840 15440

600

1.58S 1.713 1.838 2.088

3An

1500 2000 2200 3180

2.338 2.588 2.108

4.292 5.259 6.324

4400 '920 7720

ono

32760 41'805 51975 63225 75600

moo 18SS5 24795

21840 27870 34650 42150

"'"

..,

1620 2700 '.80 '''0 7560

90 120 200 320 '90 710 1000 1360 1600 2200

6' 90 150 2'0 36' 533

7290 9090

llO)lOOUI C........Io.

To"...

75.

284580

12120 18120 25L40 33060

Figure 12--Torque Required to Produce Bolt Stress

7-18

Half-Pipe Jackels

(


Pressure Vessel Design Clnd Analysis - Seminar Noles

BOLnNG DATA-R.ECOM,M£NDED MINIMUM (All Dimensions In Inche'S unleS5 (llhei"fl'ise stated)

IYt 8 0.128 IIVIl 2.0)2 i m I--"'-J----'---t-==--+---'C.II---cc,.,-+ ~HI 8 i1929 1 nl9; 2'~i

Hi

Iii

m

m

1'14

m

~--'-+--'-+':=+-c-+-~_cc-'--'-+--'-+---'----1-c-lY.. Hi

8 8

'-155 1.405

m

Hi

8

1.680:

ll~.

(

8

1.9&:1

8

2-J:l4:

3%

1m

~%

2\i

]~

H.62

~y,

2Y.

]~

4.215

511

rYi,

5259!

4%

4-688

SY.

Wi

I.4-31!

5

5.515

6~

J% 4y!

8

4.m

a

2Yo

I i

lYo

8

...::'Yo~,+.J~_O~":_I-.:~

ni

I

H23

8

3\:\

m

2.652

_._.-

I

~

8

3

2.828

3.m

8

2M

[

l%

1'14

._- -- - f -

.'.'.16 __ ~ _ J~l"'+_'~;"'--j_~I~"--i_~"~'-l--=J~"--i,-'~":...j tEn i jl'o l Ph I 1\13Y. m

2"Ai

lY.

L __ .

2V"

H.

, IY.

f---C~'---'~-:=':+-"'-.I--='-6.324 4~ 5.102 f,y, j%

8.749

l

SY.

5.923

7Y.

IO.lM

I

5y'

6..341

/11.

1I.S66

]'/,

."

7

,

/'t..

3'h

.

'"

'"

t~·~

D.C.

I (

Nut dimensions are based on American Nalional Standard 818.2.2· (1972)

Figure 13--80Iting Data-Recommended Minimum

Half-Pipe Jackels

7-19

Flange Stresses


FLANGE STRESSES Stresses on the flange are determined entirely by the bending moment on the flange. All the loads on the flange produce bending in the same direction (i.e., counterclockwise) and this bending is resisted by the ring behavior of the flange, and in integral flanges by the reaction of the pipe. The flange design moments differ from the norm for external pressure, reverse flanges, and flat flanges. Under external pressure only the end load and flange pressure are included in the design, and their sense is reversed. For reverse flanges all the moments are present, but the moment arm h o is negative, making M u negative. The load H T is negative, and the moment arm h, may be either positive or negative. The absolute value of the moment is used in the calculations. For flat faced flanges an alternate value ofhg (h"g) is used to calculate a reverse moment at the bolt circle. No calculations for seating conditions for full faced flanges arc required.

(

)

Figure 14--Calculating the Reverse Moment at the Boll Circle

Stress Analysis The stresses produce by the bending moment were calculated by Waters Rossheim, Wesslrom, and Williams using a complicated combination of elasticity and discontinuity equations.

These equations have been simplified for Code use into curves, formulas, and tables that contain constants depending on the geometry of the flange assembly. The form of the stress equations is a

=

K(geOmelry)~ I

That is, a constant (dependant on the flange geometry) times the bending moment, divided by some thickness squared, either the thickness of the flange or the thickness of the hub. The Code provides formulas so that computer programs can consistently arrive at the answers that are normally select from charts in the appendix.

7-20

Half-Pipe Jackels

(


flange Slrcsscs

Below and on the following pages are the Taylor Forge calculation sheets widely used for flange analysis calculations. Flanges are frequently subject to external forces and moments, in addition to internal pressure. Using the following fonnula, you may calculate a roughly approximate equivalent pressure for flanges loaded axially and/or in bending: P eq

=

P des

t6M -,+-, nG nG

+ 4F

Where: Equivalent Pressure Design Pressure Axial foree, lb Bending moment, in.-Ib Diameter of gasket load reaction, in. There are two commonly used tables of bolt area, which also show spacing required for the bolts. These spacings frequently govern the geometric design of the flange. The TEMA table is based on Table D-S in the TEMA Standards. The UNC table is the Unified National Course thread series, found in many places, stich as Mark's handbook. Sometimes the mating flange strongly influences the design. For instance, ifyou are designing two flanges bolted to one another in such a way as to sandwich a tubesheet in between them, you would have to analyze both flanges and detennine which one had the maximum bolt loads, then examine the mating flange under those loads.

(

Half-Pipe lackets

7-21

Pressure Vessel Design rl1ld Analysis - Seminar Noles

Flange Stresses

TYPE 1: WELD NECK FLANGE DESIGN (INTEGRAL)

1

DESIGN CONDITIONS

AlIo".ble SII.....

prllsaure. P Dllalgn lempenllule

[)e,lgn lomp.. s", "«Ill. temp.,

flallQII1N1ot18I ~log

De&lgl'l I(Jll1lP••

s..

malGl'ial

AIm. 'emp.,

~

s.

I

Conoelotl ;1II1owanol!I

2

OASKET AHD

4

TABLES 2-3 AND 2-4

I ~ING

"do"

Oaskel

3

.......

F_

W~

DETAILS LOAD AND 80U CAlCULATIONS

.. b.oy

b

H,.

G

H.

Y

Wm,-Hp+H

A", - gl&llillf of W....¥S.OfW.. t~

c. ~.omP

.P/.

",

W•

.6(A", .. ,ys.

m

5

MOMENT CAlCULATIONS

,

Coo.

lww m OplA!lng

Ho .. ,,9 f>/4

ho~R+.5g..

Ha .. Wm,-H

110 ...6{C - G) hI ... 6(fl .. III .. hnl

HT~H-Ho

",.W

6

ha -

.5{C

G}

--

Mo·_

""""", .-

~-Ho~

:;".

... ...

u,. .. H1h-r

S...""

-

K AND HUB fACTORS

K_MI

11-~"

"'"

T 1 Y

F

V

t

I

u s.,.,

II_FIh" d ..

ho",..reg;

7

,

~F7",~1F:': ~+-h,~

~tJ..g.,~

STRESS FORMULA FACTORS

0'_18 .. 1

" ....1319+1 , . ofT , .. I'/d ~~., .. 6 m" -

I"

H,

I

MJB

mc-M,;IB If bon tpacll"lO txcuds :2a .. t. rnukiply In" 8nd rna in abeNe ~Iinn by:

~~ 2a+1 lBo/IS

C

t::+f HG

G"

FIgure 2·10. Dimensional dala and torces fOt a weld neck nange (integral),

B

STRESS CALCULATIONS "lloWI~ Str...

1.5 St,

...

...

.

Operellng

1"ong_IJdiFlII hUb.

SH ,"Mg,·

Ratialllatlgs•

SA .. o1mJ).f2

~~~J~St.

Grealer or .5CSt! + ~ or .6($" -+ 8,

,. ...

Allowable 81r...

... ...

...

Se.t''''II Longitudinal hub, SH =-lrncI).g,~ Radia.lI\ar1~.

~-~t2

~~~~

GreSler 01 .~S,I + ~ ~.

5,,+5, .

.

Adilpled from Tayloi' FOIOO Inll9lnab~:H1a1. Irx:. by pelml$$icm

Figure 15--Type 1: Weld Neck Flange Design (Inlegral)

7-22

Half-Pipe Jackels

Pressure Vessel Design Clnd Anlllysis - Seminar Noles

Fl
TYPE 2: SLIP·ON FLANGE DESIGN (LOOSE)

1

....,.

DESIGN CONDITIONS

Do$lgrl P'OMUle, P

AJlo'QbIo SlnlUH

Deeigl'llemperatore Flange BoI\ng meterilll COHosion llJoMInee

"",,,,1&1

D9a1gl temp.,

FJong.

s..

Deaigtll&mp.6t. Alrn. lamp., S.

AIm. 1811'9.. Sso

2

QASI
Gu...

Facing

3

4

TABlES 24 ,\NO 2-4

LOAD AND DOLT CALCULATIONS

..

W"", .. b Y H,. .. 2brllmP

N b 0

H.G .,P,4

Y

Wml",Hp+H

"- .. glllalK of

WrdS.or W.A W ... 5
m

.

5 lood ~

Ho. ~ W...

H

'"

T

...

z

v,

Momon'

..... ""'" .... -""'" ...

hal

Mr .. HI'"

Stlldng

u,;

to· -'
K AND HUll FACTORS

K=NB

F,

~A.

V

u

0"

"".... ..mg; ,7

5.

i

'"

d=~Mol

t-

STReSS FORMULA FACTORS

a.l, + I

-.

j.-E. -+f4-Ro

ofT

~

~g,"

C.

I

),"'Y.' ... -M,IB

~~

M;.1B

H boll. spadng elCO&O(Is 23. + I, rnlltiply rno and me n above equlillion by:

28 ... 1 ~"""'" ~

!lo-

1--8.

,.

...... I!IG"

1 w

h-

~"'413I.+-t

T

-

"".

hI"= .6(f1 ... ; ....

W

'"6

CALCUlATIONS

Levltr Ann

ho .. R+g, ho .. .5(C-GI

....B 14

HJ= H

M~ENT

:

Bolls

1.

G·

Ho

Figure 2-11. Dimensional data and lorees for a slip.on lIanga (loose).

8

STRESS CALCULATIONS

Allowable Sireu 1.5s",

Openllng loIlWfucf1f1ll1 hUb.

~

=mJ).g,'

1.6

s..

s"

Retial Aanw,

s"

s"

........ "'''~

s"

s"

Gt'88191' 01 _~s.. + SlIl oc. 5H+S,.)

S~ ~ tJrnJl..l~ Sr~

m,Y/I'-

.

S..llng

!1ow"I)IeStreM

.

lOOQlludlnll Ixb, 8H _ "'<'Ag,z

s,,-_

Radial RaIl",

T........ ~

Sr

maY/l2

GUlaiet 01 .5(SH + S~ or·~+St

Adapted Irom laylor Forgo Inlernationill, loc., by perrrus.slon.

Figure 16--Type 2: Slip-On Flange Design (Loose)

Half-Pipe Jackels

7-23

flange Stresses


TYPE 3: RING FLANGE DESIGN 1

DESIGN CONDITIONS

AIJoW/lbl. Strlll.lI11

ptll9SUlli. P

[)e.slgo lempBfal"ne RatI{Ill malOfie} 8oI1i1lg rna\Orial C«roslQn a11o'wance

DeaIgn 'Imp., s.. Aim. t~ .• St.

2

""""

R"",

Desigtllomp.•

s"

Aim. lernp_. S.

OASKET AND FACJNQ DETAILS

rw",

Ge$l
3

4

TABLES 2-3 AND 2-4

LOAD AND BOLT CALCULATIONS

N

Wrrd--hG)'

b

H,.-=2brllmP H .. G'rP,", W.., .H,+H

,

0

.

A... .. 9'••ler of

W.dSIOfWlfJ~

w .. M..... -+AtJS.

m

5

....,

MOMENT tALC!JL.ATIONS

,

.

Lever Arm

""""',

~r.llng

Hg .. ,BPf4

Ito" .5{C - B)

Mo - Hoho

Ho-wm,-H

hG- .~C -0) tit ~ ,5(1ID+ ho)

Ma~liGha

H,-H

"-

M, = Hrhl

M" Slitting

",.W

M"

ha '" .S(C - 0)

6

SHAPE CONSTANTS

K. AlB

+ [II r+'~ W

IA'r.-C~

Y

~~~r.Pa:.v~u2a~~1:Ulliply

7

J8oJ1

spIocing

20 .. I

FLANGE THICKNESS REQUIREO

~h

I '" gf8lIler 01

""""'"

, Bolls

ScKlliog

...

,.~

1-

I

h

g,

'

~;

I I· •

B

H(jl

G.

HT

1

Agure 2-12. Dlmensk>nal data and forces ror a ring nange.

~ ~

.25g. 1T'"~90 > 1/4 In.~ h

B~

ho

~=90

,

-IA ~ ~<+"'i"

~.

1,.,"'90

hl-l;

g,

I

B

.7cmin

T

-g,

"'" I. •

B

Figure 2·13. Various allachmenls of ring flanges. (All olher dimensions and loadings per Figure 2·11.)

8

NmES

j

la. <1-5t.lI'ldh<9•• de$Ic)'lllo! IrI!aglIll-Kg. > 1.$1. &rdh>il"•. ~sq, ... ~ I g, sM h. BIg. s3CKl, p~ p!!illt>dClWgtl *'1:>. < TCIO". o..q, Il$ mllu.

e"'''"",,<:(I~"'I-=:~.

bulro1 loss Ihan

I/~ In.

Figure 17--Type 3: Ring Flange Design

7-24

Hair-Pipe Jackels

\

Flange Stresses


TYPE 5: SLIP-ON FLANGE, FLAT FACE, FULL GASKET I

,-

DESIGN CONomONS lIign pressln, P

DeaIgn lempereh"e fIange rnalorial BolIng material

()e$ign temp.. Aim. lOmP-.

s..

s..

Facing

3

4

TABLES 2~ AND 2-4

G C b

-""

tomp.• St.

GASKET AND FACJHO DETAILS

GB$lI:;eI

,

De~"

Alm.lelnj)., S.

T

Cotros;on aIowlll'C9

2

Ioweble Str.....

""

~{C-B""

•I H. • P/4

.

W.,_H+H,.+Ht\

H6,v .. (halh6)brGy

H,.

m

5

LOAD AND BOLT CAlCULATIONS

..,..

~

b~Gr

+ HQy 21numP

W...:I. -

A", .. grealCO" or W,dS. 01 W",,1SD w-.5(A.+~.

MOMENT CALCULATIONS

•

.

"",,,Am> Operallftg

Ho" TB'PJ4

ho .. A +111

Hr-H-Ho

1lr-.6(AtQ,+11G)

(

.

Mo" Hollo

IMr _H,h,

levlr Arrrl.

hir._l.....

"o",(C-9X2BtC) 6(9 + C)

Momenl

CK2A+C) 6(C +

A'J

Rev&l'loIl Moment

Ha .. W-H

hO .. hohQ

'-"G = He.1IG

ha+hG

6

K AND HUB FACTORS

, • AlII

z

F,

•

U

g,,.,

d .. U h"g.,a

v,

". ,mg; 0

~A.

V,

y

,7

Ho

""""

T

STRESS fORMULA FACTORS

•l,--,t' Id

"+. + I

m,. M,III

jl .. .(f31e

,-ofT " bo-. ~ IlCeed:s 2a ... I. nVtIpty m" In 1\

equalion by:

B

4Bo1l2a$p&Cino of t

STRESS CALCuLATIONS

Allowable SIre..

.. s" s"

OperlUng longitvcJ:nalllub.

Sw

EO

~;'~Il~I~s..

s"

Grtlll8r ol.5
"",.

...

I-- A~

t-

6Mo

"he - nd,)

B. 4

W

I-g.~

4

I~

L

}-9O~

~E~ .~ho

h·

\

--,H'G

rnJ),g,Z

~~~a,r.

s"

s"

i

1

--

h'o

Co

!

- -

~h,

---

HtH ., G

f4-- h

G

r G~

4 Bolls

Figure 2~15. Dimensional data and rorces lor a sJip-on nange. flat face. full gasket.

apt od from Ta y\Ol Forge 1~18mahof'lal. Int.• by pefmlulon

Figure IB--Type 5: Slip-On Flange, Flal Face, Full Gasket

Half-Pipe Jackels

7-25


Flange Stresses

TYPE 4: REVERSE FLANGE DESIGN

1

DESIGN CONOmONS

"-

Design pressu-e. P

DesIgo tempetAlu,.

Flange mel,riaJ

Design tomp., So,

Bolting maloriill

Aim. temp.,

Ss.

Corroalon alCM'anoe

2

Allowable Str"," Design

temp_.

"""""

s..

AIm.lemp., S.

OASKET AND FACING DETAILS

Ga~E'1

Facing

3

4

TABLES 2-3 AND 2·"

N

lOAD AND BOLT CALCULATIONS

..

W,.z .. bl'6), H,. "" 2/nGmP H .. OaT P/4

b

,

G

A", .. grealef of

WrdSa 01 W.,,!St.

W.,.H... +H

W-.l)(A,.,'tA.)S.

m

5

MOMENT CALCUl...ATtON$

•

"""

Ho- .8 14

Illy -

I

.5(C

(8 i-O}f2) ,

'.... '

....,ng

.... W

ho"

6

1M;

.5(C - 0)

K AND HUB FACTORS

K .. AlB'

H,

"'"

T l V

•.

Hrllr

M,- ""

moment. Ilgabralcalty. "'on UN Ihe absolute voJue 1M,1 in ..1 subsequent caleU'aliorls.

F V

,

u

I"

8 .. FIh"

",

h, • v~

01

Hob,

.....

I

0)

-

....

.5(C+g,-~-8J

ha ...5(C -

Hr~H-IlD

Add

OperllJng

hI)"

.., H

Ho~W

-

Lever Afm

_111 W

'I

d .. ~~1

.n;g;

.. ORY

UR - "AU

+

L --.

TA_l!...!:.!)...",T ~-'I

7

, BoIlS

STRESS fORMULA FACTORS a_ 11d

,

Q- _10+ 1 ,8.41'310 + 1

). ... 7+6 11\0" "VB'

't- afT"

""-MOm'

.... ,us,.. 8

... ..

G.

't-"':l.

~J". .,.

bI

~-+lK1-·)1 rV

...

W

.

, ~B'"

c.

'-oJ+ H.

+--8.

~A.

FIgure 2·14. DimensIonal dala and forces for a (everse Nange. STRESS CALCULATIONS

~

•

u_•

u...

Allow.bI. Str... 1.5s..,

longiWdinll hb, SIi" 1mJ).9,t

~-~'

.

iWlQ8ntillll fI{l.nge•

S, .. m.,y"p! -

zs..

(0.(1710" 1)'.8

a...IBf 01 ~8tl + t;i Of

SH+S,

~'lIialt~

SI! rB')

Radial ftlflg!, ~tr

s... ...

~"al~, maY~ -

Sr ..

zs..

(uno + 1lr'.8

GI'lNll." oI.6(Stt + ~ 01"

"'H"lj (kt

I~

·5(SH +

iangenl\ell ~ngo

Sr{".e') ..

-il' Iv

... ....

, ...Ung

~:~;~\AI.

~

~_ "'H"l! (l<.Z-I)).

Figure 19--Type 4: Reverse Flange Design

7-26

Half-Pipe Jackets


Flange Stresses

Pressure Vessel Design and Analysis Seminar Componenl Design Problem

Body flange for a heat exchanger - Weld neck type. A body flange (SA350,LF2) has been designed with the following dimensions: Flange inside diameter ~ 40.000 in. (1016 mm) Flange outside diameter = 47.650 in. (1210 mm) Flexitallic gasket, 40.5-in. (1029 mm)!D, 41.5-in. (1054 mm) OD, .5-in. (12.7 mm) ring 56 I-in. (25.4 mm) bolts on 45.5-in. (1156 mm) bolt circle Bolt material is SA-193, B7 Shell thickness 0.750 in. (19 mm) Hub thickness 1.250 in.(32 mm) Hub length 2.5000 in. (63.5 mm) Flange thickness 3.500 in. (89 mm)

(

No corrosion allowance

Is this flange acceptable for conditions of 500 psi (3.45 N/mm 2) at 300°F (149°C). Notes:

Questions: What arc the m and y factors for the giveu gasket?

_

What is the required thickness for the given conditions?


Half-Pipe Jackets

7-27

flange Slresses


COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------------------- Page 39 Flange Analysis: EX-2 Channel F1 Item: 1 9:51a Sep 21,2000

Input Echo, Flange Item

I,

Description: EX-2 Channel F1

Description of Flange Geometry (Type) Description of Flange Analysis Design Pressure Design Temperature Corrosion Allowance

Integral Weld Neck Analysis Only P 500.00 300.00 FCOR 0.0000

Flange Inside Diameter B Flange Outside Diameter A Flange Thickness T Thickness of Hub at Small End GO Thickness of Hub at Large End G1 Length of Hub HL Perform thickness cales. based on rigidity

40.0000 47.6500 3.5000 0.7500 1. 2500 2.5000

psig F

in. in. in. in. in. in. in.

No

Flange Material (Not Normalized) Flange Allowable Stress At Temperature Flange Allowable Stress At Ambient

SFO SFA

SA-350 LF2 20000.00 20000.00

psi psi

Bolt Material Bolt Allowable Stress At Temperature Bolt Allowable Stress At Ambient

ssa SBA

SA-l93 B7 25000.00 25000.00

psi psi

Diameter of Bolt Circle Nominal Bolt Diameter Type of Threads Number of Bolts

45.5000 1.0000 TEMA Thread Series 56 C

DB

FOD

Flange Face Outside Diameter Flange Face Inside Diameter Flange Facing Sketch Gasket Gasket Gasket Gasket

42.5000 40.0000 I, Code sketch 1.

FlO

GOD GID

Outside Diameter Inside Diameter Factor, m, Design Seating Stress

M

Y

Column for Gasket Seating

41.5000 40.5000 3.0000 10000.00

in. in.

in. in.

in. in. psi

2, Code Column I I

FLANGE ANALYSIS, FLANGE NUMBER I, Description: EX-2 Channel Fl ASME Code, Section VIII, Division 1, 1998, A-99 Corroded Flange lO, Corroded Large Hub, Corroded Small Hub, Code R Dimension,

BCOR G1COR GOCOR

Gasket Contact Width, Basic Gasket Width, Effective Gasket Width, Gasket Reaction Diameter,

7-28

R N

BO

BE G

B+2.0*FCOR G1-FeaR GO-FCOR ({C-BCOR)/2.0)-G1COR

40.000 1.250 0.750 1.500

in. in. in. in.

(GOD-GID) / 2 2.0 BO (GOD+GID) / 2.0

0.500 0.250 0.250 41.000

in. in. in. in.

N /

Half-Pipe Jackels

(

Flange Stresses

Pressure Vessel Design and Analysis. Seminar NoLes

COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------------------- Page 40 Flange Analysis: EX-2 Channel F1 Item: 1 9:51a Sep 21,2000

BASIC FLANGE AND BOLT LOADS: Hydrostatic End Load due to Pressure: H 0.785 * G • G • PEQ H; 0.7854 * 41.0000 * 41.0000 • 500.0000 H ;; 660127. lb. Contact Load on Gasket Surfaces: HP 2 * BE * PI • G • M PEQ + HP ;; 2 • 0.2500 • 3.1416 * 41.0000 • 3.0000 * 500.00 HP;; 96604. lb. Hydrostatic End Load at Flange ID: HD 0.785 * Bear * Bear * PEQ HD;; 0.785 * 40.0000 * 40.0000 * 500.0000 HD ;; 628319. lb. Pressure Force on Flange Face: HT H - HD HT ;; 660127 - 628318 HT ;; 31809. lb. Operating Bolt Load: WM1 H + HP + HPP WM1 ;; ( 660127 + 96603 + 0 ) WM1 ;; 756731. lb. Gasket Seating Bolt Load: WM2 y * (( BE * PI * G ) + (BEPG * GLPG)) + HPGY WM2 10000.00*«(0.2500*3.141*41.000)+(0.00*0.0000))+0.00 WM2 322013. lb. Required Bolt Area: AM Maximum of WM1!ABSTR, WM2!ABASTR AM Maximum of 756731 ! 25000 , 322013 ! 25000 AM 30.2692 in Bolting Information for Total Area of Bolts, AD Minimum radial distance Minimum radial distance Minimum circumferential Actual circumferential Maximum circumferential

(

TBMA Thread Series: between between spacing spac;:ing spacing

hub and bolts

bolts and edge between bolts between bolts between bolts

30.856 1.375 1.063 2.250' 2.551 8.000

in in. in. in. in. in.

Min. Gasket Contact Width (Brownell Young): Nmin AS * ABASTR!(GDSTR • PI • (GOD+GID) 30.856 * 25000.00!( 10000.00 * 3.14 ,. ( 41.50 + 40.50) ) Nmin 0.299 in.

Flange Design Bolt Load, Gasket Seating: W ABASTR * ( AM + AB ) ! 2.0 W ; 25000.00 * ( 30.2692 + 30.8560 ) W ;; 764065.56 lb. Gasket Seating Force: HG WM1 - H HG 756731 - 660127 HG 96604.00 lb.

Half-Pipe Jackels

! 2.0

7-29

Flange Stresses


COADE Engineering Software PVElite 4.00 Licensee; COADE Inc .. Local White Lock FileName: Seminar --------------------------------------- Page 41 Flange Analysis: EX-2 Channel F1 Item: 1 9:51a Sep 21,2000

MOMENT ARM CALCULATIONS; Distance to Gasket Load Reaction: (C-G)/2.0 DHG DRG '" ( 45.5000 - 41.0000 ) /2.0 DHG = 2.2500 in. Distance to Face Pressure Reaction: DHT ( R + G1COR + DHG ) / 2.0 DHT = ( 1.5000 + 1.2500 + 2.2500 I I 2.0 DHT '" 2.5000 in. Distance to End Pressure Reaction: DRO R + ( G1COR / 2.0 ) DRO 1.5000 + ( 1.2500 / 2.0 ) DHD 2.1250 in. SUMMARY OF MOMENTS FOR INTERNAL PRESSURE: LOADING Force Distance End Pressure, MD 628319. 2.1250 Face Pressure, MT 2.5000 31809. Gasket Load. MG 2.2500 96604. Gasket Seating, MA 764066. 2.2500

Bolt Corr 1.0000 1.0000 1. 0000 1.0000

TOTAL MOMENT FOR OPERATION, RMO TOTAL MOMENT FOR GASKET SEATING, RMA Effective Hub Length, HO Hub Ratio, HRAT Thickness Ratio, GRAT

SQRT(BCOR*GOCOR) HL / HO (GICOR/GOCOR)

Flange Factors for Integral Flange: Factor F per 2-7.2 Factor V per 2-7.3 Factor f per 2-7.6 Factors from Figure 2-7.1 K T 1.842 U y 11.188 2 d 131.819 in. "3 e Stress Factors ALPHA BETA -L721 GAMMA DELTA 0.325 LAMBDA

Moment 111265. 6627. 18113. 143262.

( ft.lb. ft. lb. ft. lb. fLIb.

136005. ft.lb. 143262. ft.lb. 5.477 0.456 1. 667

in.

0.846 0.287 1.000 1.191 12.295 5.772

0.154 in. "'-1 1.541 0.836 1.161

Longitudinal Hub Stress, Operating~ SHO ( f * RMO / BOOR ) / ( Rlambda * GlOOR"'2 ) SHO = { 1.0000 * 0.163E+07 / 40.0000 / ( 1.1615 * 1.2500"'2 ) SHO = 22482. psi Longitudinal Hub Stress, Seating: SHA ( f * RMA / BCOR ) / ( Rlambda * G1COR"'2 ) SHA = ( 1.0000 * 0.172E+07 /40.0000 ) / ( 1.1615 * 1.2500"'2 ) SRA = 23682. psi Radial Flange Stress, Operating: SRO (BETA * RMO / BeaR ) / ( Rlambda * TH"'2 ) SRO (1.7210 * a .163E+07 / 40.0000 ) / ( 1.1615 * 3.5000"'2 ) SRO 4935. psi

7-30

Half-Pipe lackets


Flange Stresses

COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName : Seminar -~------------------------------------- Page 42 Flange Analysis: EX-2 Channel FI Item: I 9;51a Sep 21,2000 Radial Flange Stress, Seating: SRA (BETA*RMA/BCOR) / ( Rlambda*TH~2 SRA "'- ( 1.7210 * 0.172E+07/ 40.0000 l / SRA = 5199. psi Tangential Flange Stress, Operating: STO (Y·RMO I (TH*TH*BCOR) ) - Z*SRO STO",- ( 11.1883 * 0.163E+07 / ( 3.5000'"'2 STO "'- 8777. psi Tangential Flange Stress, Seating: STA (Y*RMA / (TH'"'2*SCORl) Z·SRA STA = { 11.1883 " 0.I72E+07 / ( 3.5000'"'2 STA "'- 9246. psi Average Flange Stress, Operating: SAO (SHO + MAX( SRO, STO ) ) / 2 SAO = ( 22482 + MAX ( 4935, 8777 ) / 2 SAO "'- 15630. psi Average Flange Stress, Seating; SAA (SHA + MAX( SRA, STA ) ) / 2 SM = ( 23681 + MAX ( 5198, 9245 ) l/ 2 SAA = 16464. psi Bolt Stress, Operating: SBO (WM1 / AS ) SSO "'- ( 756731 / 30.8560 SBO = 24525. psi Bolt Stress, Seating: SBA (WM2 / AB ) SBA {322013 / 30.8560 SSA 10436. psi Stress Computation Results: OPERATING Actual Allowed Longitudinal Hub 22482. 30000. Radial Flange 4935. 20000. Tangential Flange 8777. 20000. Maximum Average 15630. 20000. Bolting 24525. 25000.

) ( 1.1615 * 3.5000'"'2 )

• 40.0000) ) - 5.7724 * 4935

* 40.0000) ) - 5.7724 * 5198

GASKET SEATING Actual Allowed 23682. 30000. 5199. 20000. 9246. 20000. 16464. 20000. 10436. 25000.

Estimated M.A.W.P. ( Operating) Estimated M.A.W.P. {Gasket Seating Estimated Finished weight of Flange Estimated Unfinished weight of Forging APP. S Flange Rigidity Index for Seating Case APP. S Flange Rigidity Index for Operating Case The PV Elite Program,

H.lf-PipeJ.ckels

509.7 716.9 612.8 894 c.2 0._827 0.785

psi psi psi psi psi

psig psig lb. lb.


7-31

Flange Stresses


, J

7-32

Half-Pipe Jackets


Flange Slrcsses


Body flange for a heat exchanger - Ring Flange Redesign the flange in pmblem I as a ring type flange.

Notes:

Questions: Which flange is lighter? Which flange is easier to fabricate?

(


(

Half-Pipe Jackets

7-33

Flange Stresses


COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------------------Page 43 Flange Analysis; EX-2 Channel Fl Item; 2 9:51a Sep 21,2000

Input Echo, Flange Item

Description: EX-2 Channel Fl

2,


Loose Ring Design, Geometry P 500.00 300.00 FCOR 0.0000

Flange Inside Diameter B Perform thickness calcs. based on rigidity

41.5000

psig F

in. in.

No


SFO SFA

SA-350 LF2 20000.00 20000.00

psi psi


SBO SBA

SA-I93 87 25000.00 25000.00

psi psi

Length of Weld Leg at Back of Ring Number of splits in Ring Flange Type of Threads

WLEG 0.3750 NSPLT o TEMA Thread Series

FOD

Flange Face Outside Diameter Flange Face Inside Diameter Flange Facing Sketch Gasket Gasket Gasket ·Gasket

44.0000 41.5000 I, Code Sketch la

FID

GOD GID

Outside Diameter Inside Diameter Factor, ro, Design Seating Stress

M

Y

Column for Gasket Seating

43.0000 42.0000 3.0000 10000.00

in.

in. in.

in. in. psi

2, Code Column II

Geometry Selection for Flange Design Option: Number of Bolts Selected for Flange Diameter of Bolts Selected for Flange Bolt Circle Diameter Selected for Flange Outside Diameter Selected for Flange Initial Thickness Given for Flange

36

1.2500 45.7500 4B.2500 5.0000

in. in. in. in.

FLANGE ANALYSIS, FLANGE NUMBER 2, Description: EX-2 Channel Fl ASME Code, Section VIII, Division 1, 199B, A-99 Corroded Flange ID, Corroded Large Hub, Corroded Small Hub, Code R Dimension, Gasket Contact Width, Basic Gasket Width, Effective Gasket Width,

7-34

BCOR GlCOR GOCOR R N

BO BE

B+2.0~FCOR

Gl-FCOR GO-FCOR ((C-BCOR)/2.0)-GICOR (GOD-GID) 2.0 BO

N /

/ 2

41.500 0.000 0.000 2.125

in.

0.500 0.250 0.250

in. in. in.

in. in. in.

Half-Pipe Jackets


Flange Siresses

COADE_Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------------------- Page 44 Flange Analysis: EX-2 Channel Fl Item: 2 9:51a Sep 21,2000 Gasket Reaction Diameter,

G

=

(GOD~GID)

/ 2.0

42.500

in.

BASIC FLANGE AND BOLT LOADS: Hydrostatic End Load due to Pressure: H 0.785" G ,., G ,., PRQ H = 0.7854 ., 42.5000 ., 42.5000 * 500.0000 H = 709313. lb. Contact Load on Gasket Surfaces: HP 2" BE * PI * G * M * PRQ ~ HP = 2 ,., 0.2500 * 3.1416 * 42.5000 ., 3.0000 * 500.00 HP = 100138. lb. Hydrostatic End Load at Flange ID: HD 0.785 * Bear * Bear * PEQ HD = 0.785 ., 41.5000 ., 41.5000 * 500.0000 HD = 676326. lb. Pressure Force on Flange Face: HT H - HD HT = 709312 - 676326 HT = 32987. lb. Operating Bolt Load: WM1 H ~ HP + HPP WM1 = ( 709312 ~ 100138 + 0 ) WM1 = 809451. lb. Gasket Seating Bolt Load: WM2 y * «( BE * PI * G ) + (BEPG ., GLPG») ~ HPGY WM2 10000.00*(0.2500*3.141o,42.500)+(0.00*0.0000)~0.00 WM2 333794. lb. Required Bolt Area: AM Maximum of WM1/ABSTR, WM2/ABASTR AM Maximum of 809451 / 25000 , 333794 / 25000 AM 32.3780 in

\

Bolting Information for TEMA Thread Series: Total Area of Bolts, AS Minimum radial distance between hub and bolts Minimum radial distance between bolts and edge Minimum circumferential spacing between bolts Actual circumferential spacing between bolts Maximum circumferential spacing between bolts Distance Across Corners for Nuts Circular Wrench End Diameter

a

Min. Gasket Contact Width (Brownell Young) : AS * ASASTR/(GDSTR • PI * (GOD~GID) Nmin 33.444 * 25000.00/( 10000.00 * 3.14 * ( 43.00 0.313 in. Nmin

(

33.444 1. 750 1.250 2.813 3.987 11. 229 2.209 3.250

~

in in. in.

in. in. in. in. in.

42.00)

)

Flange Design Bolt Load, Gasket Seating: W ABASTR * ( AM + AB ) / 2.0 W = 25000.00 ., ( 32:3780 + 33.4440 ) / 2.0 W = 822775.50 lb. Gasket Seating Force:

Half-Pipe Jackels

7-35

Flange Stresses



FileName: Seminar --------------------------------------- Page 45 Flange Analysis : EX-2 Channel FI Item: 2 9:51a Sep 21,2000 HG HG HG

WMl - H 809451 - 709312 100138.25 lb.

MOMENT ARM CALCULATIONS; Distance to Gasket Load Reaction: DBG (C - G ) I 2.0 ORG ~ ( 45.7500 - 42.5000 ) I 2.0 OHG ; 1.6250 in. Distance to Face Pressure Reaction: DHT (DRD + DHG ) I 2.0 DRT ~ ( 2.1250 + 1.6250 l I 2.0 DHT = 1.8750 in. Distance to End Pressure Reaction: DRD DRO

(C - BCOR )/ 2.0 (45.7500 - 41.5000 )

DRO

2.1250 in.

/2.0

SUMMARY OF MOMENTS FOR INTERNAL PRESSURE:

LOADING End Pressure,

MD Face Pressure, MT Gasket Load, MG Gasket Seating, MA

Force

676326. 32987. 100138. 822775.

Distance

2.1250 1. 8750 1.6250 1.6250

TOTAL MOMENT FOR OPERATION, RNO TOTAL MOMENT FOR GASKET SEATING, RMA

Bolt Corr 1.0000 1.0000 1.0000 1.0000

Moment

119766. 5154. 13560. 111418.

ft. lb.

it .lb. .lb. it.lb.

ft

138481. it.lb. 111418. ft.lb.

Effective Hub Length, HO 0.000 in. Hub Ratio, HRAT Defined as 0.0 0.000 Thickness Ratio, GRAT Defined as 0.0 0.000 Factors from Figure 2-7.1 K 1.163 T := 1.854 U 14.231 Y := 12.950 Z 6.686 Tangential Flange Stress, Operating; STO (Y*RMO / (TH*TH+BCOR) ) - Z+SRO STO = ( 12.9500 * 0.166E+07 / ( 5.0919~2 + 41.5000) ) - 6.6B58 + 0 STO = 20000. psi Tangential Flange Stress, Seating: STA (Y+RMA / (TH~2+BCOR») Z+SRA STA:= (12.9500 + 0.134E+07 / (5.0919"-2 + 41.5000) ) - 6.6858'" 0 STA = 16091. psi Bolt Stress, Operating: SBO (WM1 / AB ) SRO = ( 809451 / 33.4440 SBO := 24203. psi Bolt Stress, Seating~ SBA (WM2 / AB ) SBA (333794 / 33.4440 SEA 9981. ps~ Stress Computation Results: OPERATING Actual Allowed

7-36

GASKET SEATING Actual Allowed

Half-Pipe Jackels

\

Flange Stresses



FileName: Seminar Flange Analysis:

--------------------------------------EX-2 Channel FI

Tangential Flange Bolting

20000. 24203.

Item:

20000. 25000.

16091. 9981.

2

Page 46

9:51a Sep 21,2000 20000. psi

25000. psi

Minimum Required Flange Thickness

5.092

in.

Estimated M.A.W.P. ( Operating) Estimated M.A.W.P. ( Gasket Seating Estimated Finished Weight of Flange Estimated Unfinished Weight of Forging APP. S Flange Rigidity Index for Seating Case APP. S Flange Rigidity Index for Operating Case

500.0 7-46.9 685.6

psig psig lb. lb.


1022.3 1. 268 1.575

(c) 1989-2000 by CQADE Engineering Software

(

(

Half-Pipe Jackets

7-37

Blind Flanges and Channel Covers


BLIND FLANGES AND CHANNEL COVERS Blind flange design is a mixture of the flange analysis we have just been studying and the flat head design we reviewed previously. The ASME Code formula for a circular blind flange is I

~

"JCP/SE+ 1.9Wh/SEd'

The first term in this formula is the bending of a flat plate under pressure. The second term is the bending of the plate due to an edge moment. The stress is limited to 1.5 times the allowable stress, but the 1.5 factor is already built into the equation, as with welded flat plates. For bolt-up conditions the first term is zero-the thickness of the flange depends only on the edge bending. For non-circular blind flanges the formula is modified as follows:

(~

"JZCP/SE+6Wh/SELd'

Where Zis the samc nou-circularity factor used in welded flat plates, and the factor 1.9/d has been replaced by 6/L. L is the perimeter of the non-circular head, measured through the centers of the bolt holes. Notice, therefore, that for a circular head, d is equal to LI 3.14159, as you would expect. It is interesting to note that the Code coverS non-circular blind flanges, but no other type of non-circular flange (not even in the rectangular vessel appendix).

Chaunel Covers designed to TEMA must meet at least the minimum thickness requirements of the Code. In addition, if there is a pass partition groove, the cover deflection is limited.

The formula for flange deflection limitation is found in paragraph 9.21 ofTEMA: y

~ E~(O.0435G3p+0.5SeABhg)

The deflection is, of course, a function of (J and GJ. Thus, a very small increase in flange thickness will decrease the deflection significantly. One implication of using the TEMA formula to check a channel cover is that when the program selects thickness based on TEMA, the stresses in the channel cover may be Jess than the Code allowable. The following graph shows how the required thickness of a channel cover varies with pressure for both the TEMA and ASME requirements. Notice

7-38

Half-Pipe Jackets

(


Blind Flanges
thai the required thickness for TEMA can be higher than AS ME for some pressures, and lower for others (because ASME is a function of d', and TEMA a function of
u ~

H

j ~

J

" ", ,. ,. ,..

" ,".. ,.. " ,~

,~

D<>lt.,

'00

P,......

Figure 20--Channel Cover Thickness Note

(

Half-Pipe Jackels

The Seventh Edition ofTEMA also gives recommended deflection as a function of flange size. The previous editions hid the actual deflection you were working toward in a thickness equation.


Pressure Vessel Design and An
(

(

7-40

Half-Pipe Jackels


Blind Flanges and Chmlllci Covers


Channel Cover Design Analyze a channel cover 10 mate wilh Ihe flange designed in Ihe previous examples. Geomelry for Ihe channel cover should be as follows; Thickness ~ 5.0 in. (127 mm) Outside Diameter ~ 47.375 in. (1203.3 mm) Diameter of Bolt Circle ~ 45.125 (1146.17) 48 1.125-in. (28.5750) diameter bolts Flange Face 10 ~ 41.5 in. (1054.1 mm) Flange Face OD ~ 44.0 in. 1117.6 mm) Gasket 10 = 42.0 in. (1066.8) Gaskel OD ~ 43.0 in. (1092.2) Flexitalic Gasket (m~3.000, y=IOOOO.O) (68.94 N/mm2 )

(

Notes;

Questions; Whal is the required thickness of the flange? Why arc the slresses less than the allowable slresses at the required thickness?


Half-Pipe lackets

(

7-41



COADE Engineering Software PVElite 4.00 -Licensee: COADE Inc., Local White Lock

FileName: Seminar Flange Analysis:

Input Echo,

--------------------------------------EX-2 ChannelCV

Flange Item

Item:

3

Page 47

9:51a Sep 21,2000

Description: EX-2 Channel cv

3,


TEMA Channel Cover Analysis Only p 500.00 300.00 0.0000 FeOR

Flange Inside Diameter B Flange Outside Diameter A Flange Thickness T Perform thickness cales. based on rigidity

psig p

in.

0.0000 47.3750

in. in.

5.0000

in.

No


SFO SFA


SBO SBA

25000.00 25000.00

psi psi

Diameter of the Load Reaction, Long Span DL Diameter of the Load Reaction, short Span OS Perimeter along the Center of the Bolts L Youngs Modulus for Blind Flange MatI. YMOD Allowed Channel Cover Deflection CVDLT

47.3750 47.3750 141.7644 28500000.00 0.0300

in.


SA-350 LF2

psi psi

SA-193 B7

45.1250 1.2500 TEMA Thread Series C

DB

in. in. psi in. in. in.

••

Flange Face Outside Diameter Flange Face Inside Diameter Flange Facing Sketch

FOD FID

Outside Diameter Inside Diameter Factor, m, Design Seating Stress

GOD GID

Gasket Gasket Gasket Gasket

20000.00 20000.00

44.0000 41.5000 1, Code Sketch la

M Y

Colunm for Gasket Seating

43.0000 42.0000 3.0000 10000.00

in. in.

in. in. psi

2. Code Colunm I I

FLANGE ANALYSIS, FLANGE NUMBER 3, Description: EX-2 Channel CV ASME Code, Section VIII, Division 1, 1998, A-99

Gasket Contact Width, Basic Gasket Width, Effective Gasket Width, Gasket Reaction Diameter,

7-42

N

BO BE G

(GOD-GID) / 2 2.0

N /

BO (GOD+GIDl

/ 2.0

0.500 0.250 0.250 42.500

in. in. in. in.

Half-Pipe Jackels

Prcssure Vessel Design and Analysis - Seminar Noles

Blind Flanges and Channcl Covers

COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName : Seminar --------------------------------------- Page 48 Flange Analysis: EX-2 Channel CV Item; 3 9:51a Sep :n,2000 BASIC FLANGE AND BOLT LOADS: Hydrostatic End Load due to Pressure: H 0.785 * G ... G ... PEQ H ~ 0.7854 * 42.5000 ... 42.5000 * 500.0000 H = 709313. lb. Contact Load on Gasket Surfaces: HP 2 * BE * PI ... G ... M * PEQ + HP = 2 * 0.2500 ... 3.1416 * 42.5000 ... 3.0000 * 500.00 HP ~ 100138. lb. Hydrostatic End Load at Flange ID: liD 0.3*G*G*PEQ liD ~ 0.3 • 42.5000 * 42.5000 • 500.0000 liD ~ 270937. lb. Pressure Force on Flange Face: HT 0.0 For Blind Flanges HT = 0.0 For Blind Flanges HT=O.lb. Operating Bolt Load: WM1 H + HP + HPP WM1 ~ ( 709312 + 100138 + 0 ) WM1 = 809451. lb. Gasket Seating Bolt Load: WM2 Y'" «( BE ... PI * G ) + (BEPG * GLPG)) + HPGY WM2 10000.00*(0.2500*3.141*42.500)+(0.00*0.0000))+0.00 WM2 333794. lb. Required Bolt Area: AM Maximum of WM1/ABSTR, WM2/ABASTR AM Maximum of 809451 / 25000 , 333794 / 25000 AM 32.3780 in

(

Bolting Information for Total Area of Bolts, AS Minimum radial distance Minimum radial distance Minimum circumferential Actual circumferential Maximum circumferential

TEMA Thread Series: between between spacing spacing spacing

hub and bolts bolts and edge between bolts between bolts between bolts

Min. Gasket Contact Width (Brownell Young): Nmin A8 * ABASTR/fGDSTR * PI * (GOD+GID) 44.592 * 25000.00/( 10000.00 * 3.14 Nmin 0.417 in.

44.592 1.750 1.250 2.813 2.951 11: 071

in in. in. in. in. in.

( 43.00 + 42.00) )

( Flange Design Bolt Load, Gasket Seating: W ABASTR * (AM. + AS ) / '.0 W = 25000.00 ... ( 32.3780 + 44. 5920 ) / 2.0 W = 962125.56 lb. Gasket Seating Force: HG WMl HG 809451 HG 809451. 00 lb.

Half-Pipe Jackels (

7-43



COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------------------- Page 49 Flange Analysis: EX-2 Channel CV Item: 3 9:51a Sep 21,2000

MOMENT ARM CALCULATIONS; Distance to Gasket Load Reaction: DHG

(C - G )

I

2.0

DHG ~ ( 45.1250 - 42.5000 ) / 2.0 DHG = 1.3125 in. Distance to Face Pressure Reaction:

DHT

0.0 for blind flange

DHT = 0.0 for blind flange DHT = 0.0000 in. Distance to End Pressure Reaction: DHD

G/4.0

DHD DHD

42.5000 / 4.0 10.6250 in.

SUMMARY OF MOMENTS FOR INTERNAL PRESSURE: Force Distance MD 270937. 10.6250 Face Pressure, MT 0.0000 0. Gasket Load, MG 1. 3125 809451. Gasket Seating, MA 1. 3125 962126. LOADING End Pressure,

Bolt Carr

TOTAL MOMENT FOR OPERATION, RMO TOTAL MOMENT FOR GASKET SEATING, RMA

1.0000 1.0000 1. 0000 1.0000

Moment 239893. 0. 88534. 105232.

ft.lb. ft.lb. ft.lb. ft.lb.

(

328426. ft.lb. 105232. ft .lb.

Tangential Flange Stress, Flat Head, Operating: A .o. .o. STO 1.9 to WMl"'DHGtoBCORR/(TH 2 to G) + CtozRAT*PEQtoG 2/(TH 2) STO 1.9 to 809451 * 1.3125 * 1.0000/( 5.0000.0. 2 to 42.5000 ) + 0.30 * 1.0000 to 500.00 * 42.5000"'2 )/( 5.0000.0. 2 ) 12737. psi STO Tangential Flange Stress, Flat Head, Seating: .o. STA 1.9*W to DHG*BCORR/(TH 2 to G» + 0 A STA .. 1.9 * 962125 * 1.3125 * 1.0000 /( 5.0000 2 * 42.5000 ) + 0 STA : 2258. psi Bolt Stress, Operating: sao (WM1/AB) SBO = ( 809451 / 44.5920 SBO = 18152. psi Bolt Stress, Seating: SSA (WM2/AB) SSA (33)794 / 44.5920 SSA 7486. psi Stress Computation Results: OPERATING Actual Allowed Tangential Flange 12737. 20000. Bolting 18152. 25000. TEMA Channel Cover Deflection Calculations: Actual deflection at center per TEMA formula Allowed deflection at center of Cover

Estimated M.A.W.P.

7-44

( Operating)

GASKET SEATING Actual Allowed 2258. 20000. psi 7486. 25000. psi

0.0286 0.0300

534.0

in. in.

psig

Half-Pipe Jaekels

l



COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName : Seminar -----------------------------~--------- Page 50 Flange Analysis: EX-2 Channel CV Item: 3 9:51a Sep 21,2000

Estimated M.A.W.P. (Gasket Seating Estimated Finished weight of Flange Estimated Unfinished Weight of Forging The PV Elite Program,

3281.1 2494 .3 2494.3

psig lb. lb.

(el 1989-2000 by COADE Engineering Software

(

Half-Pipe Jackets

7-45

Large Cenlral Openings


LARGE CENTRAL OPENINGS

.'

Some flat heads have large nozzle openings. A special modification ofthc flange rules is providcd in Appendix 14 in order to analyze the stresscs in thesc hcads. The opening in the head may have a nozzle or may be a simple hole, as shown: 14-3

1998 SECTION VIII -

Tl'li' haU depiell opMlng wilh noule.

9 0 hheUI

DMSION I

"'9'0 loon leI

14-'

"_If

Thll dvpielS op!oingwilhoul noule.

1 . - - - - - - - - . , -1-----

--1

1+--1------ A----jf--------++J

Figure 21--lntegral Flat Head with Large Cenlral Opening Notice that these rules apply only to an integral head. A flat head with a large nozzle in it may be analyzed using the normal rules for flat heads and nozzles. Only when the flat head is welded or otherwise integral to the shell do these rules apply. The design procedure begins with the calculation of the normal bending moments for a flanged connection, taking the point of bending to be not the gasket, but the inside diameter of the vessel wall. The next step is to calculate the slresses at this point (the OD of the head). These are used to calculate an equivalent moment and geometry for the head with the large opening. (The equivalent moment is M JI" the equivalent geometry is expressed by (Eg)' andXj ). The equivalent geometry and loading are used to calculate the stresses at the Head/Shell Juncture and at the Opening/Head Juncture. In each case longitudinal, radial, and tangential stresses are calculated. These stresses are then compared to the standard ASME Code allowables from Appendix 2-8.

7-46

Half-Pipe JackeLs

Large Central Openings


Pressure Vessel Design and Analysis Seminar Component Desigll Problem Large Central Circular Opening. A flat head designed for 75 psi. (5 I 7N/mm 2) internal pressure at 500°F (260°C) has a 12in. (305 mOl) centrally located nozzle. Determine the stresses for the head and the nozzle openillg. The following additional data is available: Corrosion Allowance ~ 0.0625 in. (1.6 mm) Flat head thickness = 1.0 ill. (25.4 mm) Flat head outside diameter ~ 20.0 in. (508 mm) SheH thickness ~ 0.25 in. (6.35 mOl)

(

Shell hub thickness, large end ~ 0.5 in. (12.7 mOl) Shell hub length = 3.0 ill. (76.2 mOl) Flange and nozzle are both SAI06 B. Nozzle hub thickness (wall thickness) at small end

~

0.25 in. (6.35 mm)

Nozzle hub thickness at large end ~ 0.6 in. (15.24 mm) Nozzle hub length

=

2.8 in. (71.12 mm)

Notes:

Questions:

(

What is the longitudinal stress in the head at the shell? What is the longitudinal stress in the head at the nozzle?


Half-Pipe Jackels

(

7-47



COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName ; Seminar --------------------------------------- Page 105 Lgcenter Analysis: HAND CHECK Item: 1 9:51a Sep 21,2000

Input Echo, Large Opening Item

1,

Design Temperature Design Pressure Integral Flat Head Thickness Opening Inside Diameter Flat Head Outside Diameter

Description: HAND CHECK TEMP

500.00 75.00 1.0000

P T

F

psig in.

12.0000 20.0000

in. in.

FCOR

0.2500 0.5000 3.0000 0.0625

in. in. in. in.

Flange and Nozzle Material Allowable Stress at Design Temperature Allowable stress at Ambient Temperature

SFO SPA

SA-I06 B 17100.00 17100.00

psi psi

Nozzle Side Hub Thickness, Small End Nozzle Side Hub Thickness, Large End Nozzle Side Hub Length

NGO NGl NHL

0.2500 0.6000 2.8000

in. in. in.

Shell Side Hub Thickness, Small End Shell Side Hub Thickness, Large, End Shell Side Hub Length Corrosion Allowance

ASME VIII App. 1. A-99 FLANGE RESULTS

B

A

SGO SGl

1, Desc.; HAND CHECK

Diameter of the Load Reaction G ~ C per App. 14 G FLGOD - 2.0 * SGI + 2.0 * FCOR G ~ 20.0000 - 2.0 * 0.2500 + 2.0 * 0.0625 G ~ 19.6250 in. The Flange K Factor K Flange 00 / Flange ID K = 20.0000 ! 12.1250 K ~ 1.6495 The Flange Radial Distance Dimension R 0.5 * ( G - FLGCID ) - ( SG1 - FCOR ) R R ~ 0.5 * ( 19.6250 - 12.1250 ) - ( 0.5000 - 0.0625 ) R ~ 3.3125 in. The Hydrostatic End Force Hd Hd PI/4. Flange 10 A 2 ,., P Hd = PI/4 * 12.1250 A 2 * 75.0000 Hd = 8659.9355 lb. The Moment Md Md HD * ( R + 0.5 * ( SGI - FCOR ) ) Md '" 8659.94 * ( 3.3125 + 0.5 * ( 0.5000 - 0.0625 ) ) Md '" 2548.3665 ft.lb. The Total Hydrostatic End Force H A H PI!4 * G 2 * P H", PI/4 * 19.6250 A 2 * 75.0000 H '" 22686.6562 lb. The Differential End Force Ht Ht H - Hd Ht 22686.66 - 8659.94 Ht 14026.7207 lb.

7-48

Half-Pipe Jackels

\


Pressure Vessel Design tlod Analysis - Seminar Noles

COADE Engineering Software

PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar ~-------------------------------------- Page 106 Lgcenter Analysis; HAND CHECK

Item;

1

9:51a Sep 21,2000

The Moment Mt

Mt 0.5 ~ Ht * ( R + SGI ) Mt = 0.5 * 14026.72 * ( 3.3125 + ( 0.5000 - 0.0625 ) ) Mt = 2191.6750 it.lb. The Total Moment Mo Mo

Md -t Mt

Mo Mo

2548.37 + 2191.68 4740.0415 it.lb.

Flange Factors

( Opening

T

1.647

U

1.649 4.455

Y

4.054

Z

2.162

K

Factors from Figure 2-7.1

(

110 Effective Hub Length, BRAT Hub Ratio, Thickness Ratio, GRAT

HBLNG / HO

1.508 in. 1.857

(GONE/GZERO)

2.867

SQRT(S*GZERO)

Flange Factors for Integral Flange; Factor F per 2-7.2 Factor V per 2-7.3 Factor f per 2-7.6 d

4.067 in. 3 A

Stress Factors BETA DELTA

1. 501 0.203

e ALPHA GAMMA LAMBDA

0.604 0.058 1.000 0.400 in. _l 1.375 0.835 1.038 A

Longitudinal Hub Stress, Operating: A SH~ ( F . RMO / B ) / ( RLAMBDA • GONE 2 ) A SHy ~ ( 1.0000 ~ 56880 / 12.1250 ) / ( 1.0379 • 0.4375 2 ) SHy = 23613. psi Radial Flange Stress, operating: A SR. (BETA * RMO / B ) / ( RLAMBDA • TH 2 ) A SR. ~ ( 1.5006 * 56880 / 12.1250 ) / 1.0379· 0.9375 2 ') SR~ = 7716. psi Tangential Flange Stress, Operating: ST* (Y.RMO / TH*TH*B ) - Z·SRO (4.0541.56880/ 0.9375 A 2 ~·12.1250 ) - 2.1623 • 7716 STy ST~ 4954. psi

(

The Value for E Theta Star ( Integral Nozzle ) Et* 0.91. (NGl/NGOI A 2·Bl*V*Sh*/{f*ho) Et* 0.91*{ 0.538 / 0.188 )A2 ~ 12.312 * 0.058 * 23613/( 1.000 * 1.508) Et* 83723.1953 psi Flange Factors ( Shell ) : Effective Hub Length, 110 IlRAT Hub Ratio, GRAT Thickness Ratio,

(

Half-Pipe Jackels

SQRT(B*GZERO) HBLNG / HO (GONE/GZERO)

1.936 in. 1.549 2.333

7-49



COADE Engineering Software PVElite 4.00 Licensee; COADE Inc .. Local White Lock FileName : Seminar --------------------------------------- Page 107 Lgcenter Analysis HAND CHECK Item; 1 9:51a Sep 21,2000 Flange Factor Factor Factor

Factors for Integral Flange: F per 2-7.2 V per 2-7.3 f per 2-7.6 e •

0.663 0.095 1.000 0.343 in .... -1

Moment Acting at the Shell to Plat Head Juncture Mh Et*/C(1.74*ho*V/(SGO"'3*B1»+Et*/Mo*(1+Pt/ho») Mh 83723 /[{1.74 * 1.9365 * 0.0947 /( 0.1875"'3 * 19.9375 »+ 83723 / 56880 * ( 1 + 0.6633 * 0.9375 / 1.9365 ) Mh 1595.8574 ft.lb. The Xl Factor Xl (Mo - Mh ( 1 + Ft/ho ) ) / Mo Xl (56880 - 19150 ( 1 + 0.6633 * 0.9375 I 1.9365 ) ) / 56880 Xl 0.5552

(

Stress Results for the Head/Shell Juncture Longitudinal Hub Stress in the Shell : Shs (Xl) (E*) (l.lO*ho*fj / ((gl/g0) "'2*Bs*V) Shs (0.56){ 83723) (1.10 * 1.936 * 1.0001/(( 5.444)"'2 * 19.62) 0.095) shs 9788.3799 in. Radial Stress at the Outside Diameter ; A srs 1.91*Mh(1+F*t/ho)/(Bs*t 2) + 0.64*F*Mh/{Bs*ho*t) A Srs 1.91 * 1595{ 1 + 0.663 * 0.938 I 1.936 1/( 19.625 * 0.879 2 ) + 0.64 * 0.663* 1595/( 19.625 * 1.936 * 0.938 ) Srs ~ 3029.7085 psi Tangential Stress at the Outside Diameter : A Sts X1*Etheta**t/Bs - .57 {1+P*t/ho)Mhl (Bs*t 2) + .64*P*Z*Mh/(Bs*ho*t) Sts = 0.555 * 83723.195* 0.938 / 19.625 .57( 1 + 0.663 * 0.938 / 1.936) 1595 /( 19.625 * 0.8789"'2 ) + .64 * 0.663 * 2.162 * 1595 I( 19.625 * 1.936 * 0.938 ) Sts = 1877.8865 psi where Z

=

(K"'2+1)/(K A 2-1)

=

(

2.721 + 1 )/( 2.721 - 1 )

2.162

Stress Results for the Opening Head Juncture Longitudinal Hub Stress in Central Opening : sho = Xl * Sh* = ( 0.555 * 23613.205 ) = 13110.363 psi Radial Stress at Central Opening Sro = Xl * Sr* = ( 0.555 * 7716.479 )

=

4284.292 psi

Tangential Stress at Diameter of Central Opening Sto X1*St* + .64*P*Zl*Mh/(Bs*ho*tl Sto 0.555 * 4953.979 + .64 * 0.6633 * 3.162 * 1595 I ( 19.625 * 1.936 * 0.938 ) Sto 3472.0491 psi

7-50

Half-Pipe lackets



CQADE Engineering Software

PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------------------Lgcenter Analysis: HAND

where Zl =

2*K~2/(KA2-1)

~HECK

=

Item:

(2* 2.721 +

1

Page 108

9:51a Sep 21,2000

1/12.721 - 1

=

3.162

Flange Stress Results per Appendix 14

Allowed

Head/Shell

Allowed

Opening

25650 17100

13110

25650

Radial

9788 3029

4284

17100

Tangential

1877

17100

3472

17100

Long.

Hub

M.A.W.P. for the given Geometry Estimated Finished Weight of Forging Approximate Minimum Flange Thickness The PV Elite Program,

146.735 69.3 0.5444

psi psi psi

psig

lb. in.

(c) 1989-2000 by CQADE Engineering Software

(

Half-Pipe Jackets

7-51


7-52

Pressure Vessel Design ilnd Analysis - Seminar Notes

Half-Pipe Jackels

\

Chapter 8: Floating Heads and Spherically Dished Covers Floating hcads and spherically dished covers are discussed in Appendix 1-6.

TYPES OF SPHERICALLY DISHED COVERS There arc four types of spherically dishcd covers: Type A covers are evaluated using nomlal head rules. Only the other three rcquire special analysis. The Type B head has a thin spherical section extended across the flange face, with a flange ring behind the plate. Type C heads arc made from a single piccc of steel the thicknesses of the flange with the centcr portion dished. Type D heads, very commonly used as floating heads for heat exchangers, consist of a flange ring with the spherical cap welded to the inside of the ring. Type D heads are the most intcresting, because of the inward reaction of the head to internal pressure (and outward to external pressure, sincc these are usually found inside heat exchangers). An additional analysis techniquc, called Soehren's calculation, takes into account the rotation of the flange and head at thc attachment point. This calculation may be used as a design basis.

DESCRIPTION Thcse heads always have a flange, which must be designed using modified flange rules, and frequently a backing ring which is designed as a ring flange . .Backing rings usually have slightly larger diameters than the flange ring. The backing ring may be a split ring. If the ring has one split, then it has been split along a diameter, into two pieces. The bending moment on the ring is multiplied by 2.0 for this case. A ring with two splits has been sliced in half like a bagel, and then each half has been split along a diameter. The ring is assembled with the diametral splits offset by 90 degrees. For this case, enter the thickncss of one half of the original ring, since cach half is required to support 75 percent of the original design momcnt. The attachment point of the head to the flange is measured by two diffcrent values.

Q is the distance from the bolting face of the flange to the intcrsection of the head inside diamctcr and the flange. This dimension is uscd in Soehren's calculation. HR is lhe distance from the flangc ccntroid to the intcrscction of the head centerline and the flange. HR is positive if it is above the flange centroid, and negative if it is below thc flange centroid. HR is used in the Code calculation.

Floating Heads and Spherically Dished Covers

8-t

Example


A slolled flange hns bolt holes that extend radinlly to the onter edge of the flange. Slotted flanges are allowed by the Code for floating heads. A full face gasket extends from the 10 of the flange 10 the OD, enclosing the bolt holes. These gaskets are nsually soft malerials such as rubber or nn elastomer, so that the bolt stresses do not go too high during gasket seating. The Code takes into account the use of full face gaskets for floating heads. The tubeside pressure is internal to the flange, creating a typical flange bending moment, to which is added the moment caused by the pressure on the head acting to rotate the flange. If the head is allached above the centerline of the flange, this will result in a moment that is opposite to the olher flange bending moments. Ifthe head is allached below the centerline, this moment will have the same sign. The shellside pressure is external to the flange. The rules in Appendix 2 of the Code for flanges under external pressure are used by the program to calculate the bending moment on the flange. Note that because the sign of the pressure on the head is also reversed, the head component of the flange bending will always have the same effect that it does for internal pressure.

EXAMPLE The following figure shows typical geometry for floating heads and spherically dished heads. ,oJ

Al'PENDD.: I -

MANDATORY

1..

NotL.qn.....

'tend 1/1 No CtM IMt: T'htn Il2itl.

..-~"9

It- ~~.

L\

a....ft· IIl~F""T'r'fIt

1,1

Figure 1--Typical Geometry for Floating Heads and Spherical Caps

8-2


Example

Pressure Vessel Design and AIl
Pressure Vessel Design and Analysis Seminar Component Design Problem Special Design Problem-Floating Head 275 psi shell side

. 1.89 N/sqmm

215 psi tube side

1.48 N/sqmm

Design temperature

300°F

149°C

Corrosion allowance

0.125 in.

3.175 mm

Head Material Flange and backing ring Bolting Gasket ID Gasket OD Spiral Wound SS/Asbestos Gasket Number of Bolts Bolt size Backing ring ID Backing ring thickness

SA5l6,70 SAlO5 SA193, Grade B7 45.125 in. 45.688 in.

II 46mm 1160 mm

52 314-in. 45.6875 in. 3.75 in.

19mm 1160 mm 95mm

Design Pressure

0.675 in_ \hick

rC\ 1

Hr

Centroid

T

7500 in thick

'5.00

I

(

I

'5.125

---j

'6.750

I '9.375

Questions: What is the required head thickness? What are the required flange thicknesses?

_

What are the required ring thicknesses? DO NOT TURN THE PAGE UNTIL YOU HAVE COMPLETED YOUR ANALYSIS


(

8-3

Example


COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName; Seminar --------------------------------------- Page 34 Plohead Analysis; EX-2 PLOHEAD Item: 1 9:51a Sep 21,2000

Input Echo, Floating Head Item

1,

Description: EX-2 FLOHEAD

Floating Head Type Appendix 1-6 type (d) Tube side ( Internal ) Design Pressure PTS 215.00 Shell Side ( External ) Design Pressure 275.00 PSS Design Temperature for Spherical Head TEMP 300.00 Head Material Head Allowable Stress at Temperature Head Allowable Stress at Ambient

psig psig F

SAH

SA-516 70 20000.00 20000.00

psi psi

Crown Radius for Spherical Head CR Head Thickness TH Tube Side ( Internal ) Corrosion Allowance CATS Shell Side ( External ) Corrosion Allowance CA.sS

45.0000 1.1250 0.1250 0.1250

in. in. in. in.

SOH

Flange Material Flange Allowable Stress at Temperature Flange Allowable Stress at Ambient

SOC SAC

SA-lOS 20000.00 20000. 00

psi psi

Flange Outside Diameter Flange Inside Diameter Flange Thickness

FOD FlO TC

49.3750 45.1250 7.7500

in. in. in.


SBO SBA

SA-I93 B7 25000.00 25000.00

psi psi


DB 46.7500 DBOLT 0.7500 TEMA Thread Series 52

Full Face Gasket ( Yes or No Flange Flange Gasket Gasket

Face OUtside Diameter Face Inside Diameter Outside Diameter Inside Diameter

Gasket Gasket Flange Column Gasket Flange Length Width

Factor, m, Design Seating Stress Facing Sketch for Gasket Seating Thickness Face Nubbin Width of partition Gasket of Partition Gasket

in. in.

No

FOD FlO GOD GIO

-45.7500 45.1250 45.6880 45.125~

3.0000 10000.00 I, Code Sket.ch la 2, Code Column I I 0.0000 0.0000 0.0000 0.0000 M Y

Distance from Head Centerline to Flange centroid

1.5000

in. in. in. in.

psi

in. in. in. in. in.

The Flange is not Slotted.

8-4


)


Example

COADE Engineering Software PVElite 4.00 I.icensee: COADE Inc .• Local White Lock FileName: Seminar --------------------------------------- Page 35 Flohead Analysis; EX-2 FLOHEAD Item: 1 9:51a Sep 21,2000

Backing Ring Material Specification Backing Ring Allowable Stress, Temperature SOTS Backing Ring Allowable Stress, Ambient SATS Backing Ring Inside Diameter DR Backing Ring Thickness TR Number of Splits in Backing Ring NSPLIT

SA-IDS 20000.00 20000.00 45.6875 3.7500

psi psi in.

in.

1

INTERNAL PRESSURE RESULTS FOR SPHERICAL HEADS ASME Code, Section VIII, Div. 1, 1998, A-99 Appendix 1-6 Thickness Due to Internal Pressure: t 5PL / 6S per Appendix 1-6 t (5 * 215.00 * 45.1250 ) / t 0.4042 in.

( 6

*

20000 )

Maximum Allowable Working Pressure at Given Thickness: Pa 6S(T-Cass-CatsJ / 5L per Appendix 1-6 Pa (6 * 20000 * 0.8750 ) / ( 5 * 45.1250 ) Pa 465.37 psig Maximum Allowable Working Pressure, New and Cold: Pnc 6ST / 5L per Appendix 1-6 Pnc (6 * 20000 * 1.1250 ) / ( 5 * 45.0000 ) Pnc 600.00 psig Actual stress at given pressure and thickness: Sact 5PL / 6(T-Cass-Cats) per Appendix 1-6 Sact ( 5 * 215.00 * 45.1250 ) / ( 6 * 0.8750 Sact 9240. psi EXTERNAL PRESSURE RESULTS, SPHERICAL HEAD ASME Code, Section VIII, Division 1, 1998, A-99 External Pressure Chart CS-2 Elastic Modulus for Material

(

at

300.00 29000000.00

Results for Maximum Allowable External Pressure: Corroded Thickness of Shell TCA 0.875Q in. Outside Crown Radius of Shell 00/2 46.0000 in. Crown Radius / Thickness Ratio (00/2) /T 52.5714 0.0023777 Geometry Factor, A f(D/T,LD) A Materials Factor, B, f(A, Chart) 15586.9805 psi B Maximum Allowable Working Pressure 296.49 psig EMAWP = B/( (OD/T)! 2 ) = 15586.9805 /( 105.1429/2.0) = 296.4915 Results for Reqd Thickness for Ext. Pressure (Tea) : Corroded Thickness of shell TCA 0.B217 Outside Diameter of Shell OD 92.0000 Diameter / Thickness Ratio (O/T) 111.9651 Geometry Factor, A f(DT.LD) A 0.002232B Materials Factor, B, f(A, Chart) 15397.0B9B B

(

F psi


in. in.

psi

8-5

Example

Pressure Vessel Design and Anslysis - Seminar Notes

COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName; Seminar --------------------------------------- Page 36 Flohead Analysis: EX-2 FWHEAD Item: 1 9:51a Sep 21,2000 Maximum Allowable Working Pressure EMAWP = B!((D!T)!2) = 15397.0898!< 111.9651 ! 2

275.03 275.0337

psig

INTERMEDIATE CALCULATIONS FOR FLANGED PORTION: ASME Code, Section VIII, Div. 1, 1998, A-99 App. 2 Gasket Contact Width, Basic Gasket Width, Effective Gasket Width, Gasket Reaction Diameter,

N BO BE G

(GOD-GID) ! 2 ! 2.0 BO (GOD-tGID) I 2.0

0.281 0.141 0.141 45.407

N

Bolting Information for TEMA Thread Series : Total Area of Bolts Minimum radial distance between hub and bolts Minimum radial distance between bolts and edge Minimum circumferential spacing between bolts Actual circumferential spacing between bolts Maximum circumferential spacing between bolts

in. in. in. in. in.

in.

in

BASIC FLANGE AND BOLT LOADS: Hydrostatic End Load due to Pressure Contact Load on Gasket Surfaces Hydrostatic End Load at Flange ID Pressure Force on Flange Face Radial Component of Head Membrane Force Operating Bolt Load: Gasket Seating Bolt Load Required Bolt Area

H HP HD HT HH WM1 WM2 AM

348148.4 25900.4 347665.5 482.9 609473.4 374048.8 200778.0 14.962

lb. lb. lb. lb. lb. lb. lb. in

Flange Design Bolt Load Gasket Seating Force Distance to Gasket Load Reaction Distance to Face Pressure Reaction Distance to End Pressure Reaction

W HG DHG DHT DHD

383324.4 25900.4 0.6717 0.6796 0.6875

lb. lb.

SUMMARY OF MOMENTS LOADING End Pressure, Face Pressure, Gasket Load, Floating Hd. Load, Gasket Seating,

in. in. in.

FOR INTERNAL PRESSURE: Md Mt Mg Mh Ma

Force 347666. 483. 25900. 609473. 383324.

Distance 0.6875 0.6796 0.6'717 1.5000 0.6717

Bolt Corr 1. 0000 1.0000 1. 0000 1.0000 1.0000

TOTAL MOMENT FOR OPERATION ( Internal Pressure ) TOTAL MOMENT FOR GASKET SEATING ( Int. Pressure) BASIC FLANGE AND BOLT LOADS; Hydrostatic End Load due to Pressure Contact Load on Gasket Surfaces Hydrostatic End Load at Flange ID Pressure Force on Flange Face Radial Component of Head Membrane Force Operating Bolt Load:

8-6

15.704 1.125 0.813 1. 750 2.823 14 _786

in. in. in.

H HP HD HT HH >lMl

Moment 19918. ft. lb. 27. ft.lb. 1450. ft. lb. -76184. ft.lb. 21458. ft.lb. 54789. ft. lb. 21458. ft.lb.

445306.1 33128.4 444688.5 617.7 779559.0 374048.8

lb. lb. lb. lb. lb. lb.



Example

COADF. Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------------------- Page 37 Flohead Analysis: EX-2 FLOHEAD Item: 1 9:51a Sep 21,2000

Gasket Seating Bolt Load

WM2

Required Bolt Area Flange Design Bolt Load Gasket Seating Force Distance to Gasket Load Reaction Distance to Face Pressure Reaction Distance to End Pressure Reaction SUMMARY OF MOMENTS LOADING End Pressure, Face Pressure, Floating Hd. Load, Gasket Seating,

(

200778.0 14 .962

lb.

l\M

W HG DHG DHT DHD

383324.4 445306.1 0.6717 0.6796 0.6875

lb. lb.

in

in. in. in.

FOR EXTERNAL PRESSURE: Md Mt Mh Ma

Force 444688. 618. 779559. 383324.

Distance 0.0158 0.0079 1.5000 0.6717

Bolt Corr 1.0000 1.0000 1.0000 1. 0000

TOTAL MOMENT FOR OPERATION I External Pressure ) TOTAL MOMENT FOR GASKET SEATING I Ext. Pressure

)

Moment 584. ft.lb. o. ft.lb. -97445. ft. lb. 21458. ft.lb. 96861. it.lb. 214.58. ft. lb.

Required thickness for Main Flange, internal operating conditions: T F+SQRT{F* F + J ) per 1-6{g) 1.273 + SQRT( 1.273 * 1.273 + 18.257 ) T T 5.7316 in. Required thickness for Main Flange, internal bolt-up conditions: T F + SQRT ( F * F + J ) per 1-6{g) T 0.000 + SQRT( 0.000 * 0.000 + 7.150 T 2.6740 in. Required thickness for Main Flange, external operating conditions: T F+SQRT(F* F + J ) per 1-6(g) 1.628 + SQRT( 1.628 * 1.628 + 32.276 ) T 7.5385 in. T Required thickness for Main Flange, external bolt-up conditions: T F + SQRT( F * F + J ) per 1-6 (g) T 0.000 + SQRT( 0.000 • 0.000 + 7.150 T 2.6740 in. Required thickness for Backing Ring, internal operating conditions: T SQRT( M Y / S B ) Per App 2-7(b) (9) SQRT( 397426.9 * 24.888 / ( 20000.0 * 45.687 ) ) T 3.2901 in. T

(

Required thickness for Backing Ring, internal bolt-up conditions: T SQRT( M Y / S B ) Per App 2-7 (b) (9) T SQRTI 407282.2 * 24.888 / ( 20000.0 * 45.687 ) ) T 3.3306 in. SUMMARY OF REQUIRED THICKNESSES: Tubeside (Internal) Pressure Shells ide (External) Pressure Tubeside Gasket Seating Load

FloaLing Heads and Spherically Dished Covers

(

Head 0.4042 0.8217

Flange 5.7316 7.5385 2.6740

Backing Ring 3.2901 in. in. 3.3306 in.

8-7

Example


COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------~-------------------___________ Page 38 Flohead Analysis: EX-2 FLOHEAD Item: 1 9:51a Sep 21,2000 Shells ide Gasket Seating Load Maximum + Corrosion Allowance Actual Thickness as Given WEIGHT Weight Weight Weight Weight Weight Weight

OF of of of of of of

l.0717 1.1250

HEAD AND FLANGE: Spherical Head, Uncorroded Spherical Head, Corroded Flange Ring, Uncorroded Flange Ring, Corroded Backing Ring, Un corroded Backing Ring, Corroded


(c)

2.6740 7.7885 7.7500

WHO WHOCA

WFL WFLCA

WBR WBRCA

in. 3.3306 in. 3.7500 in.

544.9 lb. 428.9 lb. 691.8 ·lb. 610.4 lb. 292.2 lb. 252.6 lb.


(

(

8-8



Chapter 9: Heat Exchanger Tubesheets

lEMA TUBESHEETS Tubesheets are usually analyzed using the TEMA Standards (Standards of the Tubular Exchanger Manufacturers Associalion). The 7th Edition ofTEMA was issued in 1988, with addenda. The following TEMA tubeshcet types are illustrated in Figure 9-1: A stationary tubesheets, gasketed on both sides. B stationary tubesheets, integral with the shell. C stationary tubcshecls, integral with the channel. N stationary tubeshccts, integral on both sides. U V P S T W D

U-tube tubesheets gaskelcd on both sides. U-tube tubesheets integral with the channel. floating tubesheets, outside packed floating head. floating tubeshects, head with backing device. floating tubesheets, pull through floating head. floating head, externally sealed floating tubesheet. Divided floating tubeshee!.

F

Fixed tubesheet exchanger - two stationary tubcsheets.

The TEMA Standard has a very simple equation for lhe thicknes~ of the tubeshee!. The equation in bending (which controls far more often than shear) is

T is the required thickness. F is a correction faclor taking into account the fixity of the shell support for the tubeshee!. It is 1.0 for simply supported (gaskeled) tubesheets. G is the characteristic diameter of the tubesheel support point-either the 1D of the shell or the mean diameter of the gasket. Lowercase h is the ligament efficiency of the tubcsheet. The value ofh depends on the tube pattern, tube pitch and tube OD. This term was added to the equation (in place of a constant) in the Seventh edition ofTEMA. S is the allowable stress.

Heat Exchanger Tubcsheets

9-1

TEMA Tubesheets


Pis Ihe pressure. But P is really a wild card. TEMA aeeounls for the complexities of fixed lubesheet geometry and tubes heels extended as flanges by developing multiple eandidales for an equivalenl design pressure on each side oflhe tubeshee!. These candidates lake inlo aeeounlthe loading on Ihe lubesheet from differenliallhermal expansion, Ihe loading due to bolling, and Ihe aelual pressure loads.

fRONT INO "AT1OHA.lJ HEAD rn£5

~

,...-

A

~r-

AND REM

8

Aau COVER

e & ,:~

E

~r

F

t· N

G

~

~ ..

:L~

SPfCIAl HIOH 'RfSSURf: CIOSUIl£'

~

M

-

~ --m-l----- ~

~IT UKE

~ liKE

N

flX£D rUBHmn ~10- STAIIONAIlY HEAD

mefO TUBESHUT ·B~ nAflONAJiY H£AO

C~ flXfD TUUSHEET LIKE Y S'Al1ON101l'Y HEAD

!>f'l.1T F\OW

H

:1: .....

(HANNU INTfGIAL wnH lUUSHEet AND REMOVABLE (OVER

L

TWO p.us SHEll WITH lONGITUDINAl aAFflE

~~

Q:

I~

~~_._ ..__ ... -

80NNET (INTEGRAL COVER)

CHANNEl INTEGRAl WITH TUBESHEeT AND RfMOVABl£ COVER

TT

ONf 'ASS SHnL

':'_~'.

C

lEAl INO HEAD lYPU

SHIll rnu

I] ,jP ~ 50"

+

~ '-i-

OUTSIN '.0.0£0 1l0ATlIiG HEAOI

DOUBLE SPlIT nOW'

J

T

I~

5

~

T

~I

flOATING HE100 WITH IAO:ING D[VlCf

OMOfD flOW

K

KETTlE TYPE REBOILER

X

~

:

ClOSS flOW

~'.:~f4 . '~\'\.

r .... Jl. ~~gf'~...!~....

POll TtUIOUGH fLOAflN(; HE"'O

U

~

D

U·TUBE BUNDlE

~

w

~ "

UlUNAUY SfAUO nOAlING IOBESHUT

Figure 1--Standards of the Tubular Exchanger Manufacturers Association One of the most difficult areas of fixed tubesheet design is deciding on the actual metal temperature for the components. The most conservative approach is to choose metal temperatures which are based on the

maximum difference belween Ihe shellside and tubeside fluids. For example, if the shellside fluid enters the exchanger a1450°F and leaves Ihe exchanger al 340°F, then the shellside mclaltemperature would be assumed 10 be 450°F. Ifthe tubeside fluid enlers at 180°F and leaves al 250°F, then the tube metallemperature would be taken at 180°F. This would

9-2

HeaL Exchanger Tubesheers

(


TEMA Tubcsheets

result in a 270°F temperature difference between the shell and the tubes, which would almost certainly require an expansion joint. An approach that is one step Icss conservative takes thc metal temperatures at the mean of the inlet and outlet temperatures for the fluids. In the case described above, this results in a shell metal temperature of 395°F and a tube metal temperature of 215°F. The best approach is to know something about the actual metal temperatures through heat transfer considerations. We recommended asking a process engineer what the process design implies about the metal temperatures. Alternately, if you can find out the approximate shellside and tubeside heat transfer coefficients, you can estimate the tube metal temperature using a simple weight average as TrM ~

(

HsTsF + HrTrF H +H

s

r

In the case above, if the shellside fluid has a heat transfer coefficient that is five times greater than the tubeside fluid, then, starting from the average temperature for each side, the tube metal temperature would be 365°F. The shell metal temperature would remain 395°F. Thus it would be likely that the exchanger could be built with no expansion joint. Note

Typically a metal temperature difference of up to 50°F can be tolerated with no expansion joint. Note also that there may be operating scenarios (such as a loss of fluid) where the temperature difference is much higher. These must be taken into account in the design.

TEMA allows differenlial pressure design. The differential pressure is used as the design pressure on both the tubeside and the shellside. The exception to this rule is fixed tubesheet exchangers, which have special differential design pressure rules. When a tubesheet may be controlled by shear stress, TEMA requires the perimeter and area of the tubesheet for the shear calculation. The perimeter is the distance covered when stepping around all of the tubes on the periphery of the bundle. The area is the area enclosed by this process. It is difficult to empirically evaluate the perimeter and area. You will be conservative if you overestimate the area and underestimate the perimeter. TEMA calculates a pressure below which shear calculations need not be considered. However, even above this pressure, bending may control the tubesheet thickness. Fixed tubesheet heat exchangers frequently require expansion joints. For TEMA evaluation, all you need to know about the expansion joint is its spring constant and inside diameter. If you have a flanged and flued expansion joint-that is, one that is relatively stiffyou will have to know the exact spring constant. For metal bellows expansion joints you only need to know that the spring rate is below a certain value prescribed by TEMA. When a tubesheet is extended as a flange, additional bending moments may be imposed on the tubesheet. TEMA transforms these moments into an equivalent pressure. The bending moments on the tubesheet extension are calculated assuming the extension is a ring-type flange, and using the typical Code flange analysis. TEMA also provides rules for longitudinal stresses, both in the shell and in the tubes. You need to know the baille spacing and tube physical and material properties to determine the buckling stress in the tube.

Heat Exchanger Tubesheets

9-3

TEMA Tubeshects


The following table shows lhe tube joinl reliability factor used by TEMA to compare the actual to the maximum tube pull-au I load. IS98 SECTION VII DIVISION 1 TABLEA-2 EFPICIENCIESfr Dew:iption (l)

9-4

No ...

Heal Ex.changer Tubesheets


TEMA Tubesheels


Floating TubesllCet Design Determine the required thickness for the following floating tubesheet. Tubesheet type - Floating tubesheet with backing ring, SA-105 Forging Gasket diameter for tubesheet - 38 in. (965 mm) Design pressure - 250-psi (1.72 N/mm 2) shell at 480°F (249°C), 525-psi channel at 200°F (93°C) 3/4-in. tubes on 15116 pitch, triangular Shell: SA-516,70, 37-in. diameter, 0.5-in. thick, 0.125-in. CA Channel: SA-5 I 6,70, 37-in. diameter, 0.5-in. thick, 0.125-in. CA Gasket between Channel and Tubesheet Flange Face OD ~ 38.5 in. (978 mm) Flange Face ID ~ 37.5 in. (953 mm) Gasket OD ~ 38.5 in. (978 mm) Gasket ID ~ 37.5 in. (953 mm m ~ 2.75, y = 3700.0-psi (2.55 N/mm 2) Notes:

(

Questions: What is the required thickness of the tubeshect? What is the maximum pressure allowed on the tubeside for a 4-in.(102 mm) thick tubesheet?


Heal Exchanger Tubeshecls

9-5

TEMA Tubesheels


COADE Engineering Software PVElite -4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------------------- Page 51 Tubesheet Analysis: EX-) Float TS Item: 1 9:51a Sep 21,2000

Input Echo, TubeSheet Item

Description: EX-) Float TS

1,

Shell Design Pressure Shell Temperature for Internal Pressure Shell Material Shell Allowable Stress at Temperature Shell Allowable Stress at Ambient Shell Thickness Shell Corrosion Allowance Inside Diameter of Shell

PS TEMPS

250.00 480.00 SA-516 70 20000.00 20000.00 0.5000 0.1250 37.0000

psig

Channel Design Pressure Channel Temperature for Internal Pressure Channel Material Channel Allowable Stress at Temperature Channel Allowable Stress at Ambient Channel Thickness Channel Corrosion Allowance Inside Diameter of Channel

PC TEMPC

psig

DC

525.00 200.00 SA-516 70 20000.00 20000.00 0.5000 0.1250 37.000Q

TUBESHEET TYPE: Floating, Backing Ring Tube Outside Diameter DT Tube pitch (Center to Center Spacing) FT Tube Layout Pattern Tubesheet Design Metal Temperature TEMPTS Tubesheet Material (Not Normalized) Tubesheet Allowable Stress at Temperature SOTS Tubesheet Allowable Stress at Ambient SATS Thickness of Tubesheet TTS Tubesheet Corr. Allowance (Shell side) CATS Tubesheet Corr. Allowance (Channel side) CATC Depth of Groove in Tube Sheet GROOVE

0.7500 0.9380 Triangular -480.00 SA-lOS 19680.00 20000.00 4.0000 0.0000 0.0000 0.0000

SOS SAS TS CAS OS

SOC SAC

TC CAC

ADDITIONAL DATA FOR GASKETED TUBESHEETS: Flange Face Outside Diameter Flange Face Inside Diameter Flange Facing Sketch Gasket Outside Diameter Gasket Inside Diameter Gasket Factor, m, Gasket Design Seating Stress Column for Gasket Seating Tubesheet Gasket on which Side

38.5000 FID 37.5000 1, Code Sketch la GOD 38.5000 GID 37.5000 2.7500 M y 3700.00 2, Gode Column II SIDE BOTH FOD

F

psi psi in. in. in.

F

psi psi in.

in. in.

in. in. F


in. in. in. in. psi

INTERMEDIATE CALCULATIONS FOR GASKETED TUBESHEETS: ASME Code, Section VIII, Div. I, 199B, A-99 Appendix 2 Gasket Contact Width, Basic Gasket Width, Effective Gasket Width, Gasket Reaction Diameter,

9-6

N BO BE G

(GOD-GID)

I 2

N I 2.0 BO (GOD+GID) /

2.0

0.500 0.250 0.250 38.000

in. in. in. in.



TEMA Tubesheets

COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------------------- Page 52 Tubesheet 1\Ilalysis: EX-3 Float TS Item: 1 9:51a Sep 21,2000 TUBESHEET ANALYSIS, TUBESHEET NUMBER l,Description:EX-3 Float TS TEMA Standards, Eighth Edition, 1999, RCB-7 Tubesheets TEMA R-I.131 Minimum Tubesheet Thickness for R-type: THIN MAX ( DT, CONST - ( CATS ~ CATC ) ) TMIN =: MAX ( 0.750, 0.750 - ( 0.000 + 0.000) ) TMIN = 0.7500 in. Min. Thickness + CATS + MAX (CATC, GROOVE) TREQMIN = 0.7500 in. Shellside Fixity Factor, F, per RCB Shellside Effective Diameter, per RCB Tubeside Fixity Factor, F, per RCB Tubeside Effective Diameter, per RCB TEMA Eta factor used in calculation Shells ide Tubeside

7.132 7.132 7.132 7.132

Effective Pressure, Bending, Effective Pressure, Bending,

TEMA RCB-7.132 Required Thickness TRS FS I GS I SQRT { PSU / ( TRS = 1.0000 I 38.0000 I SQRT( TRS = 2.2025 in. TEMA RCB-7.132 Required Thickness TRC FC I GC I SQRT { PTU / ( TRC = 1.0000 I 38.0000 I SQRT( TRC = 3.1918 in. TEMA RCB-7.132 Required Thickness TREQ = 3.1918 in.

FS GS FC GC ETA

1. 0000 38.0000 1.0000 38.0000 0.4201

PSU PTU

250.0000 525.0000

in.

psig psig

for Shellside Pressure: ETA I SOTS) ) / 3.0 250.00 / ( 0.4201 I 19680 ) ) / 3.0 for Tubeside Pressure: ETA I SOTS) ) / 3.0 525.00 / ( 0.4201 I 19680 ) ) / 3.0 for Bending + CATS + MAX ( CATC,GROOVE):

No Shear Calculation, since Pressure is less than SUMMARY of RESULTS for Tubesheet calculations Reqd Tubsheet Thk. + CATS + MAX(CATC,GRV)TREQ Actual Tubsheet Thickness as Given TTS The PV Elite Program,

in.

1264.8911

3.1918 4.0000

psig

in. in.


(


9-7

TEMA Tubeshcets

9-8


Heal Exchanger Tubeshcels


TEMA Tubesheets


Fixed tubesheet design Determine the required thickness for the following fixed tubesheet. Tubesheettype - SA-lOS Forging Design pressure - 250 psi shell at 480°F (249°C), 525-psi channel at 480°F (274°C) 3/4-io.( 19 mm) tubes on 15/16 (24 mm) pitch, triangular Shell- SA-516,70 - 40.5lD (1029 mm) by 0.5 (12.77 mm) thick, 0.125 (3.175 mm) corrosion allowance

Channel- SA-516,70 - 40.5 lD (1029 mm) by 0.5 (12.77 mm) thick, 0.125 (3.175 mm) corrosion allowance

(

Tubesheet is welded to shell, extended as flange, and gasketed to channel. aD of extension is 47.65 (1210 mm), thickness of extension is 3.5 in. (89mm) 56 I-in. bolls on 45.5-in.(1156 mm) boll circle, SA-193, B7 40.5-in. (1029 mm) gasket lD, 41.5-in. (1054 mm) aD, flexitallic (m~3.0, y~10000.)

40.5-in. (1029 mm) face lD, 42.5-in. (l080 mm) face aD. Tubes (SA-214, allowable ~IO ksi (69N/mm 2), yield=23 ksi (159 N/mm2» Tube metal temperature 241°F (116°C) Shell metaltemperalure 259°F (126°C) Tubesheet melal temperature 250°F (121°C) No expansion joint 1511 tubes, 0.075-in. (1.9056 mm) thick, 96-in. (2438 mm) overall length 44-in.(l1I8 mm) spacing all baillles Tube to tubesheet joinl factor (ASME) 0.65 Notes:

(

Questions: What is the required thickness of the tubcsheet? What is the maximum pressure allowed on the tubeside for a 4-in.(l02 mm) thick tubesheet? DO NOT TURN THE PAGE UNTIL YOU HAVE COMPLETED YOUR ANALYSIS

Heal Exchanger Tubesheels

9-9

TEMA Tubesheels


CQADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------------------- Page 53 Tubesheet Analysis; EX-3 f-'ixed TS Item: 2 9~51a Sep 21, 2000

Input Echo, TubeSheet Item

2,

Shell Design Pressure Shell Temperature for Internal Pressure Shell Material Shell Allowable Stress at Temperature Shell Allowable Stress at Ambient Shell Thickness Shell Corrosion Allowance Inside Diameter of Shell Channel Design Pressure Channel Temperature for Internal Pressure Channel Material Channel Allowable Stress at Temperature Channel Allowable Stress at Ambient Channel Thickness Channel Corrosion Allowance Inside Diameter of Channel

Description: EX-3 Fixed TS

PS

250.00 480.00 SA-516 70 20000.00 20000.00 0.5000 0.1250 40.5000

TEMPS

SOS SAS TS CAS

OS

PC

OC

525.00 480.00 S1I-516 70 20000.00 20000.00 0.5000 0.1250 40.5000

TUBESHEET TYPE: Fixed Tubesheet Exchanger Tube Outside Diameter OT Tube Pitch (Center to Center Spacing) PT Tube Layout Pattern Tubesheet Design Metal Temperature TEMPTS Tubesheet Material (Not Normalized) Tubesheet Allowable Stress at Temperature SOTS Tubesheet Allowable Stress at Ambient SATS Thickness of Tubesheet TTS Tubesheet Corr. Allowance (Shell side) CATS Tubesheet Corr. Allowance (Channel side) C/l,TC Depth of Groove in Tube Sheet GROOVE

0.7500 0.9380 Triangular 480.00 SA-lOS 19680.00 20000.00 4.0000 0.0000 0.0000 0.0000

TEMPC

SOC SAC TC CAC

ADDITIONAL DATA FOR FIXED TUBESHEET EXCHANGERS Actual Metal Temperature for Shell ACTUALl Actual Metal Temperature for Tubes ACTUAL2 Actual Metal Temperature for Tubesheet ACTUAL3 Number of Tubes TNU>lT Length of Tubes TLENT Tube Material Tube Allowable Stress at Temperature SOT Tube Allowable Stress At Ambient SAT Tube Yield Stress At Operating Temperature SYT Tube Wall Thickness TT ASME Tube Joint Reliability Factor Fl\SME Expansion Joint Spring Rate SJF Expansion Joint Inside Diameter at Bellows OJ Maximum Unsupported Tube Length SL Tube Span Between Baffle and Tubesheet SLT Is this a Kettle-type configuration

259.00 241.00 250.00

psig F

psi psi in. in. in. psig F

psi psi in. in. in.

in. in. F

psi psi in.

in, in. in.

F

F F

.1511

96.0009 SA-214 1l400.00 11400.00 21400.00 0.0750 0.6500 0.00 40.5000 44.0000 44.0000

in. psi psi

psi in. lbs./in. in.

in. in.

No

ADDITIONAL DATA FOR TUBESHEETS EXTENDED AS FLANGES:

9-10

Heal Exchanger Tubesheets


TEMA Tubesheets

COADE Engineering Software PVElite 4.00 Licensee: COhOE Inc., Local White Lock FileName: Seminar --------------------------------------- Page 54 Tubesheet Analysis: EX-3 Fixed TS Item: 2 9:51a Sep 21,2000 Outside Diameter of Flanged Portion OF 47.6500 Diameter of Bolt Circle DB 45.5000 Thickness of Extended Portion of Tubesheet TF 3.5000 Nominal Bolt Diameter DBOLT 1.0000 Type of Threads (Thread Series TEMA Thread Series Number of Bolts NUMBER 56 Bolt Material SA-193 B7 Bolt Allowable Stress At Temperature SBO 25000.00 Bolt Allowable Stress At Ambient 25000.00 SB' Weld between Flange and Shell/Channel WLDH 0.0000

(

ADDITIONAL DATA FOR GASKETED TUBESHEETS; Flange Face Outside Diameter Flange Face Inside Diameter Flange Facing Sketch Gasket Outside Diameter Gasket Inside Diameter Gasket Factor, m, Gasket Design Seating Stress Column for Gasket Seating Tubesheet Gasket on which Side

FOD

42.5000 40.5000 1, Code Sketch 1a GOD 41.5000 GID 40.5000 M 3.0000 Y 10000.00 2, Code Column. II SIDE CHANNEL FID

in. in. in. in.

psi psi in.

in. in. in. in. psi

INTERMEDIATE CALCULATIONS FOR TUBESHEETS EXTENDED AS FLANGES: ASME Code, Section VIII, Div. 1, 1998, A-99 Appendix 2 Gasket Contact Width, Basic Gasket Width, Effective Gasket Width, Gasket Reaction Diameter,

(

N

BO BE G

(GOD-GID) / 2 2.0 BO (GOD+GID) / 2.0

N /

0.500 0.250 0.250 41.000

in. in. in. in.

BASIC FLANGE AND BOLT LOADS: Hydrostatic End Load due to Pressure: H 0.785*G*G*PEQ H = 0.7854 * 41.0000 * 41.0000 * 525.0000 H = 693134. lb. Contact Load on Gasket Surfaces: HP 2 * BE * PI • G * M * PEQ + HP = 2 * 0.2500 • 3.1416 • 41.0000 * 3.0000 • 525.00 HP = 101434. lb. Hydrostatic End Load at Flange 10: HD 0.785 * Bear * Bcor * PEQ HD = 0.785 * 40.7500 * 40.7500 * 525.0000 HD = 684706. lb. Pressure Force on Flange Face: HT H - HD HT = 693133 - 684706 HT = 8427. lb. Operating Bolt Load: WMI H + HP + HPP WM1 = ( 693133 + 101434 + 0 ) WMI = 794568. lb. Gasket Seating Bolt Load: WM2 = y * (t BE • PI * G ) + (BEPG * GLPG) + HPGY


9-11

\ TEMA Tubesheets


~U1d

Analysis - Seminar Notes

\ \


4.00

Licensee: COADE Inc., Local White Lock

FileName: Seminar ------------------------~-------------- Page 55 Tubesheet Analysis: EX-3 Fixed TS Item: 2 9:51a Sep 21,2000 WM2

~

lOOOO_OO·((O.2S00~).141*41.000>~10.00*O.OOOO)}+O.OO

WM2

~

32201J. lb.

\

Required Bolt Area: AM AM AM

~

Maximum of WMl/ABSTR, WM2/ABASTR Maximum of 794567 I 25000 31.7821 in

I

Bolting Information tor TEMA Thread Series: Total Area of Bolts, AS

** WARNING ** Required Bolt Area Minimum Minimum Minimum Actual Maximum

radial distance radial distance circumferential circumferential circumferential

between between spacing spacing spacing

>

\

322013 I 25000

30.356

\

in

Actual Bolt Area

hub and bolts bolts and edge between bolts between bolts between bolts

1. 375 1.063

2.250 2.551 6.000

in. in. in. in.

\

in.

\

Flange Design Bolt Load, Gasket Seating: 1I.BlI.STR * ( AM + AB ) I 2.0 ~ 25000.00 * ( 31.7827 + 30.6560 ) I 2.0 "" 7e298J.BB lb. Gasket Seating Force: HG WMI - H 794567 - 69)133 HG HG 101434.19 lb.

" "

\

"

\

MOMENT ARM CALCULATIONS: Distance to Gasket Load Reaction: DHG (C - G ) j 2.0 DHG "" ( 45.5000 - 41.0000 ) DHG "" 2.2500 in.

\

I 2.0

_-I

Distance to Face Pressure Reaction: nUT '" I R + GleOR + DHG ) I 2.0 DHT '" ( 2.0000 + 0.3750 DHT "" 2.3125 in.

~

2.2500 l

\

I 2.0

Distance to End Pressure Reaction: DHD R + ( GICOR I 2,0 I DHD '" 2.0000 + < 0.3750 DHD 2.1875 in.

\

I 2.0 )

SUMMARY OF MOMENTS FOR INTERNAL PRESSURE: Force LOADING Distance End Pressure, MD 684706. 2.1875 Face Pressure. MT 8427. 2.3125 Gasket: Load, 101434. 2.2500 MG Gasket Seating, MA 782984. 2.2500

TOTAL MOMENT FOR OPERATION, RMO TOTAL MOMENT FOR GASKET SEATING, RNA

Bolt Carr 1.0000 1.0000 1.0000 1.0000

Moment 124.816. 1624. 19019 . 146809.

ft. lb. ft.lb. ft.lb. it.lb.

(

\

\

1<15459. ft .lb. 14.6809. ft-Ib.

TUBESHEET ANALYSIS, TUBESHEET NUMBER 2.Descriptian:EX-3 Fixed TS TEMA Standards, Eighth Edition, 1999, RCB-7 Tubesheets

\ \ \ \

9-12


\ \

TEMA Tubesheets

Pressure Vessel Design and AIl<\lysis - Seminar Notes

COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------------------- Page 56 Tubesheet Analysis: EX-3 Fixed TS Item; 2 9:51a Sep 21,2000

TEMA R-7.131 Minimum Tubesheet Thickness for R-type: TMIN MAX( DT, CONST - ( CATS + CATC ) ) TMIN;; MAX( 0.750, 0.750 - ( 0.000 + 0.000 ) ) TMIN ;; 0.7500 in. Min. Thickness + CATS + MAX(CATC,GROOVE) TREQMIN;; 0.7500 in. Shells ide Fixity Factor, F, per RCB Shellside Effective Diameter, per RCB Tubeside Fixity Factor, F, per RCB per RCB Tubeside Effective Diameter, TEMA Eta factor used in calculation

7.132 7.132 7.132 7.132

FS GS FC GC ETA

1.0000 40.7500 1.0000 40.7500 0.4201

MATERIAL PROPERTIES FOR THERMAL EXPANSION ANALYSIS: Shell - TEMA : C-Mn-Si, 1-1/4Cr-1!2Mo & 3Cr-1Mo Coeff. of Thermal Exp. at Metal Temp. 0.6108E-05 Elastic Modulus at Metal Temp. 0.2870E+08 Tubes - TEMA : Plain Carbon Stl & C-Mn Stl. 0.6752E-05 Coeff. of Thermal Exp. at Metal Temp. 0.2829E+08 Elastic Modulus at Metal Temp. TubeSheet - TEMA : C-SI, C-1/2Mo & Cr-l/2Mo 0.6260E-05 Coeff. of Thermal Exp. at Metal Temp. 0.2825E'l-08 Elastic Modulus at Metal Temp. TEMA TEMA TEMA TEMA

(

in. in.

/ deg F psi / deg F psi ! deg F psi

0.2045 RCB-7.161 K Factor for Thermal Expansion 3.7930 RCB-7.161 Fq Factor for Thermal Expansion -0 .12E-06 RCB-7.161 Differential Expansion / Length RCB-7.161 Equivalent Differential Thermal Expansion Pressure: PD 4 * RJ * ELAS * TSCA * DLL / ( DOS - 3.0*TSCA ) • ( 1.0 + RJ * RK ,., FO l FD 4 * 1.0000 * 0.287E+08 * 0.3750 * 0.0000 / ( 41.5000 - 3.0 * 0.3750 ) * ( 1.0 + 1.0000 ., 0.2045 • 3.7930 ) PD;; -0.07 psig

RCB 7.162 Equivalent Shellside Bolting Pressure: PBS 6.2. RMA / ( F*.2 * G**3 ) h A PBS = 6.2 * 0.176E+07 / ( 1.0000 2 * 40.7500 3 PBS = 161.42 psig RCB 7.162 Equivalent Tubeside Bolting Pressure: PST 6.2. RMO / ( F**2 * G**3 ) A 6.2 * 0.175E+07 / ( 1.0000 A 2 * 40.7500 3 PBT PBT 159.93 psig TEMA RCB 7.163 Effective Shellside Design Pressure: TEMA RCB-7.163 Fs Factor for Shells ide Design 0.4882 PSP1 0.4 * RJ * ( 1.5 + RK * (1.5 + FFS ) ) 0.4 * 1.0000 * ( 1.5 + 0.2045 • (1.5 + 0.4882 ) ) PSP2 « 0.5 - RJ / 2.0 ) * ( DJ**2 / G**2 - 1.0 ) ( ( 0.5 - 1.0000 / 2.0 ) * ( 40.7500**2 / 40.7500**2 - 1.0 ) PSP3 1. 0 + RJ * RK * FQ 1.0 + 1.0000 * 0.2045 * 3.7930 PSP PS * ( ( PSPI - PSP2 ) / PSP3 )

Heal Exchanger Tubcsheets

9-13

TEMA Tubes/wets Pressure Vessel Design and Analysis _Seminar Notes COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName 'Seminar ------------------_____________________ Page 57 Tubesheet Analysis, EX-3 Fixed TS Item, 2 9,51a Sep 21,2000 PSP

250.00 ,0; ( ( 0.7627 - 0.0000 ) /1.7758 107.37 psig

TEMA RCB-7.163 Eff. Shell Side Design Pressure: PSU = Max (absolute) of PSP 107.37 or PBS PSP - PD) / 2 PD - PBS) /2 or ( PBS + PD ) / 2 or ( PSP - PBS)

161.42 (107.37 -0.07 / 2.0 (107.37 -0.07 161.42 ) / 2.0 (161.42 + -0.07 (107.37 161.42 ) / 2.0 psu '" ) TEMA RCB 161.42 psig 7.163 Shellside Shear Design Pressure, PSS : 107.37 psig or or ( PSP

RCB 7.164 Effective Tubeside Design Pressure: TEMA RCB-7.164, Ft Factor for Tubeside Design PTPI 1.0 ~ 0.4 * RJ * RK .. ( 1.5 + FFT ) PTP2 PTP PTP

0.6724

1.0 ++ RJ 0.4 * *RK1.0000 1.0 ,0; FQ .. 0.2045 * ( 1.5 + 0.6724 ) 1.0 + 1.0000 * 0.2045 ... 3.7930 PT * PTP1 / PTP2 525.00 ... 1.1777/1.7758 348.19 psig

TEMA RCB-7.164 Effective Tube Side Design Pressure: PTU = Max (absolute) of ( PTP + PBT + PD ) / 2 ( 348.19 ~ 159.93 + -0.07 ) / 2.0 or ( PTP + PBT ) ( 348.19 + 159.93 PTP - PSP + PBT + PO) / 2 or ( PTP - PSP + PBT ) (348.19 107.37 + 159.93 + -0.07 ) / 2.0 PTU (348.19 107.37 + 159.93 RCB 508.12 psig 7.164 Tubeside Shear PTS Design Pressure (PBT=O): : 348.19 psig TEMA RCB-7.132 Required Thickness for Shellside Pressure: TRS FS'" GS ,0; SQRT ( PSU / ( ETA * SOTS) ) / 3.0 TRS '== 1. 1.0000 < 40.7500 < SQRTI 161.42 / I 0.4201 < 19680 ) ) / 3.0 TRS 8979 in. TEMA RCB-7.132 Required Thickness for Tubeside Pressure: TRC FC,o; GC ... SQRT ( PTU / ( ETA" SOTS) ) / 3.0 TRC '"= 3.3673 1.0000 in. < 40.7500 < SQRTI 508.12 / I 0.4201 < 19680 TRC

) / 3.0

TENA Thickness for Bending < CATS + MAxI CATC, GROOVE) , TREQRCB-7.132 = 3.367] Required in. No Shear Calculation, since Pressure is less than

1264.8971

psig

TENA RCB-7.134 Required Thickness for Tubesheet Flanged Extension, TPREQ .98* SQRT( (RM* fR"'2 - 1 + 3.71'" R"'2 ... LN IR») / ( SOTS .. ( DF - G -' .. ( 1.0 + 1.86 ... R.. . 2 ») TFREQ .98< SQRTII 0.176E+07<1 1.17'2 I + 3.71< 1.17'2
9-14

Heat Exchanger Tubesheels

TEMA Tubesheets


COADE Engineering Software PVElite 4.00 Licensee; COADE Inc., Local White Lock FileName Seminar --------------------------------------- Page 58 Tubesheet Analysis; EX-3 Fixed TS Item; 2 9:51a Sep 21,2000

RCB-7.22 Shell Longitudinal Stress: Max. Effective Pressure for Longitudinal Stress,

284.18

psig

TEMA RCB-7.22 Maximum Shell Longitudinal Stress: STSMAX PSSMAX * CS * ( DOS - TSCA )/( 4 * TSCA ) STSMAX ~ 284.18 * 1.00 * ( 41.5000 - 0.3750 ) / ( 4 * 0.38 ) STSMAX ~ 7791. psi TEMA RCB-7.22 Allowable Shell Longitudinal Stress STSALL ~ 20000. psi RCB-7.23 Tube Longitudinal Stress Results: Max. Effective Pressure for Longitudinal Stress, Min. Effective Pressure for Longitudinal Stress,

255.12 -75.27

psig psig

TEMA RCB-7.23 Maximum Tube Long. Stress (Tension): STTMAX ~ PTTMAX*CT*FO*G*G /(4*TNUMT*TT*(DT-TT) STTMAX ~ 255.12* 1.00* 3.79*-40.7500* 40.7500 / ( 4* 1511* 0.0750*( 0.7500- 0.0750) ) STTMAX ~ 5251.62 psi TEMA RCB-7.23 Allowable Tube Long. Stress (Tension): STSALL = 11400.00 psi TEMA RCB-7.23 Minimum Tube Camp. Longitudinal Stress STTMIN = PTTMIN*CT*FO*G*G/{4*TNUMT*TT*(DT-TT») STTMIN = -75.27* 1.00* 3.79* 40.7500* 40.7500 / ( 4* 1511* 0.0750*( 0.7500- 0.0750) ) STTMIN -1549.40 psi

Modulus of Elasticity of Tubes at Mean Tube Temperature External Pressure Chart CS-l Elastic Modulus for Tube Material,

at ET

300.00 F 29000000.00 psi

TEMA RCB-7.24 Allowable Tube Compressive Stress STTCOM -PI**2 * ET / ( FSAF * RKLR**2 ) STTCOM - 3.14**2 * 0.290E+08 / ( 1.35 * 183.24**2 STTCOM -6297.79 psi RCB-7.25 Tube-To-Tubesheet Joint Load: Effective Pressure for Tube-to-Tubesheet Load

255.12

psig

TEMA

RCB-7.25 Actual Tube-To-Tubesheet Load WJ PI * FQ * PTLOAD * G * G / ( 4.0 * TNUMT ) WJ = 3.14 * 3.79 * 255.12 * 40.75 • 40.75 / ( 4.0 * 1511 ) WJ = 835.23 lb. TEMA RCB-7.25 Allowable Tube-To-Tubesheet Load: WJA (PI/4.0)* (DT 2-(DT-2*TT) **2)*SOT*FASME WJA (3.14/4.0)*( 0.7500 A 2-( 0.7500-2.0 * 0.0750)**2) * WJA 11400 * 0.65 WJA 1178.51 lb. A

LONGITUDINAL TUBE STRESS SUMMARY


Actual

Allowable

9-15

TEMA Tubesheels Pressure Vessel Design and Analysis - Seminar NOles COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar ---------------------__________________ Page 59 Tubesheet Analysis; EX-3 Fb-:ed TS Item: 2 9:51a Sep 21,2000 Shell Long. Stress (Tens.+J Tube Long. Stress (Tens.+) Tube Long. Stress (Comp.-) Tube-to-Tubesheet Load.

7791.29 5251.62 -1549.40 935.23

SUMMARY of RESULTS for Tubesheet Calculations Reqd Tubsheet Thk. + CATS + MAx(CATC.GRV)TREQ Actual Tubsheet Thickness as Given TTs Required Thickness for the Flanged Portion TFREQ Actual Thickness for the Flanged Portion TF Equivalent Differential Ther. Expansion Pressure Shells ide Prime Pressure Tubeside Prime Pressure The PV Elite Program,

20000.00 114 00.00 -6297.79 1179.51

3.3673 4.0000 2.0204 3.5000 -0.0740 107.3695 349.1878

psi psi psi lb.

in. in. in. in. psig psig psig


':c.

9-16


}

. ;t

."


ASME Tubesheels

ASME TUBESHEETS Beginning in 1989, the ASME Code began to include rules for some kinds of tubesheels. These rules have now been expanded 10 include bolh U-Iube exchangers and mosl kinds of fixed tubesheet exchangers. These are located in appendix AA of the Code, an oplional appendix, and most people continue to use the TEMA rulcs for most exchangers. However, there are many occasions in which the ASME rules can result in a thinner tubeshcct, at the expense of a slightly more complicated analysis.

0,

(

(til Typkil Sholl.Sldo Inleg';l1 ;lnd ChannQI·SldG G;I~hlQd TypQ Wrth Twuhul £:t'ond~d n a Flango

Tld>II~hwl

ConHnxllon

Figure 2--Typical Fixed Tubesheet Type Constructions The design considerations and calculations for ASME tubesheets are similar 10 those for TEMA tubesheets which we have already considered. Notice, however, the following additional considerations:

(

The ASME method requires a specific and independent calculation of each design case (temperature plus pressure, pressure only, temperature only). The TEMA method takes these cases into account by means of the various candidates for design temperature and pressure. The ASME method uses a model in which the tubesheet is considered to be a flat plate on an elastic foundation. This model typically results in a thinner tubesheet than caleulated by the TEMA melhod. In cases where the stresses at the junetion between the tubesheet and the shell are high, the ASME method allows the introduction of plastiticity at this joint. Recalculating using a reduced stiffness al the joint typically results in a significant reduction in the stresses. The CodeCalc program performs this type of recalculation automatically. This Appendix was rewritten in 1989.

Heat Exchanger Tubeshecls

9-17

ASME Tubeshee.s


(

9-J8


Pressure Vessel Design and Analysis - Selllll1ar Noles

ASME Tubesheels

Pressure Vessel Design and Analysis Seminar Component Design Problem ASME Tubesheet Calculation: Recalculate the required thickness of the fixed tubeshcet analyzed using lhe ASME program. Tubesheel type - SA-105 Forging. Design pressure - 250-psi shell at 480°F (249°C), 525-psi channel at 480°F (249°C) 3/4-in.(19mm) tubes on 15116 (24mm) pitch, triangular Shell - SA-516,70 - 40.5 (I 029mm) !D by 0.5 (12.7mm) thick, 0.125 (3.175mm) corrosion allowance

Channel- SA-516,70 - 40.5 (l029mm)!D by 0.5 (l2.7mm) thick, 0.125 (3.175mm) corrosion allowance

Tubeshect welded to shell, extended as flange and gasketed to channel. aD of extension is 47.65 (1210 mm), thickness of extension is 3.5 in. (89mm) 56 I-in. bolts on 45.5-in. (1156 mm) bolt circle, SA-193, B7 40.5-in. (1029 mm) gasket!D, 41.5-in. (1054 mm) aD, flexitallic (m~3.0, y=IOOOO) 40.5-in. (1029 mm) facc!D, 42.5-in. (1080 mm) face aD. Tubes (SA-214, allowable ~10 ksi (69N/mm 2 , yield~23 ksi (159 N/mm 2 ) Tube metal temperature 241°F (116°C) Shell metal temperature 259°F (126°C) Tubesheet metal temperature 250°F (121°C) No expansion joint

1511 tubes, 0.075 thick (1.9050 mm), 96-in.(2438 mm) overall length. 44-in.(l118 mm) spacing all barnes. Tube to tubesheet joint factor (ASME) 0.65

( Notes:

Questions: What is thc rcquircd thickness oflhe tubeshect? How does this requircd thickness compare to the rcquired thickness calculated using the TEMA program?


Hettl Exchanger Tubesheets

9-19

\ Pressure Vessel Design and Ana!ysis - Seminar Notes

ASME Tubesheels

\

coADE Engineering Software PVElite 4.00 Licensee; COADE Inc., Local White Lock FileName; Seminar -----------------------~--------------- Page 93

ASMETube Analysis: EX-2 FIXED COM

INPUT ECHO, TUBESHEET NUMBER

shell Design Pressure Shell Thickness

Shell Corrosion Allowance Inside Diameter of Shell Channel Design Pressure Channel Thickness Channel Corrosion Allowance Inside Diameter of Channel

1

Item~

Description: EX-2 FlXED COM Ps Hs CAS Os

250.00 0.5000 0.1250 -40.5000

psig in. in. in.

Pt HC CAC

525.00 0.5000 0.1250 40.5000

psig

DC

Tubesheet Design Metal Temperature TEMPTS Tubesheet Material Specification Tubesheet Allowable Stress at Temperature S Tubesheet

~llowable

Stress at Ambient

Thickness of Tubesheet Tubesheet Carr. Allowance (Shell side) Tubesheet Carr, Allowance (Channel side) TUBESHEET TYPE:

9:51a Sep 21,2000

480.00 S1\-105 1.96BQ.OO

S1\TS

20000. 00

H

4.0000 0.0000 0.0000

CATS

CATS

\

in.

in. in.

\

F

psi psi in.

in. in.

\ (

Fixed Tubesheet Exchanger

\

ADDITIONAL DATA FOR fIXED TUBESHEET EXCHANGERS Shell Temperature for Internal Pressure DELTs shell Material Shell Allowable Stress at Temperature Ss Shell Allowable Stress at Ambient SAS Channel Temperature for Internal Pressure DELTC Channel Material Channel Allowable Stress at Temperature So SAC Channel Allowable Stress at Ambient Length of 'I'tJbes RL Tube Fixity condition k Tube Yield Stress At Operating Temperature SYT Expansion Joint Spring Rate ( Ibs./in. ) Sj Expansion Joint Projection from Shell 00 wj Maximum Unsupported Tube Length SL Tube Span Between Baffle and Tubesheet SLT Mean Shell Metal Temp. along Shell len. TEMPSM Mean Tube Metal Temp. along Tube length TEMPTM Tubesheet Metal Temp. at Rim TDASH Shell Metal Temp. at Tubesheet TDASHS TDASHC Ch~nnel Metal Temp. at Tubesheet ASME Tube Joint Reliability Factor FASME TJNT Classification for Tube Joint Type Number of Tubes Tube Wall Thickness Tube Outside Diameter Design Temperature of the Tubes Tube Layout Pattern Tube Material

4BO.00 S1\-516 70

20000.00 20000. 00 480.00

psi psi

\

F

SA-516 70

20000.00 20000.00 96.00 O.QO 2HOO.00

o. 0.0000 44.00 4.4.0000 0.00 0.00 0.00 0.06 0.00

psi psi in.

\

psi lbs./in.

in.

\

in. in. F F

\

F F

F

0.65

r

\

(

RNTUBS

1511

t D

0.0,50 0.7500

in. in.

480.00

F

\

Triangular SA-214

\ \ \ "-.c.;"'-

9-20


\

\ \


ASME Tubesheets

COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------------------- Page 94 Item: 1 9:51a Sep 21/2000 ASMETube Analysis: EX-Z FIXED COM Tube Allowable Stress at Temperature Tube Allowable Stress At Ambient Tube Pitch (Center to Center Spacing) Radius to Outermost Tube Hole Center Tube-side pass partition groove depth Largest Center-to-Center Tube Distance Length of Expanded Portion of Tube

SOT SAT

11400.00 11400.00 0.9375 18.0000 0.0000 0.0000 0.0000

P

rc hg

Ul Pedd

ADDITIONAL DATA FOR TUBE SHEETS EXTENDED AS FLANGES: DF 47.6500 Flanged Tubesheet Outside Diameter Diameter of Bolt Circle DB 45.5000 Flange Thickness TF 3.5000 Nominal Bolt Diameter DBOLT 1.0000 Type of Threads TEMA Thread Series 56 Number of Bolts Bolt Material SA-193 87 Bolt Allowable Stress At Temperature SBO 25000. 00 SBA 25000.00 Bolt Allowable Stress At Ambient Weld between Flange and Shell/Channel WLDH 0.0000 ADDITIONAL DATA FOR GASKETED TUBESHEETS; Flange Face Outside Diameter Flange Face Inside Diameter Flange Facing Sketch Gasket Outside Diameter Gasket Inside Diameter Gasket Factor, m, Gasket Design Seating Stress Column for Gasket Seating Tubesheet Gasket on which Side

FOD 42.5000 FID 40.5000 I, Code Sketch 1a GOD 41.5000 GID 40.5000 3.00 M y 10000.00 2, Code Column II CHANNEL SIDE

Installation Temperature of Assembly

70.00

psi psi in. in. in. in. in.

in. in. in. in.

psi psi in.

in. in.

in. in. psi

F

ASME TubeSheet Results per APP. AA. 1998 Addenda A-99 Gasket Contact Width, Basic Gasket Width, Effective Gasket Width, Gasket Reaction Diameter, Flange Design Bolt Load

N

80 8E G

(GOD-GID) I 2.0 80 (GOD+GID)

I 2

N

I 2.0 W

0.500 0.250 0.250 41. 000 782983.88

in.

in. in.

in. lb.

Note: Since ASHE does not provide an explicit formula for flanged extensions, the TEMA formula will be used.

(

Req. Thk. of Flanged Extension for Fixed Tubesheet Per TEMA 8th Edition FTREQ .98ISQRT((RMICRAZ-l+3.71IRA2ILN(R»)/((S*(A-G)I(1+1.86*RA2» FTREQ .98+SQRT( 0.176E+07+( 1.1693 A2-1+3.71* 1.1693 A2 I LN( 1.1693»)/ ( 19680+( 47.6500- 40.7500)*(1+1.86* 1.1693 A Z» FTREQ 2.0204 in. Elasticity/Expansion Material Properties

Heal Exchanger Tubcshccts

9-21

ASME Tubesheels


COADE Engineering Software PVElite 4.00 Licensee; COADE Inc., Local White Lock FileName: Seminar --------------------------------------- Page 95 ASMETube Analysis EX-2 FIXED COM Item; 1 9:51a Sep 21,2000 Shell - TEMA C-Mn-Si, 1-1/4Cr-1/2Mo & 3Cr-1Mo Coeff. of Thermal Exp. at Metal Temp. along Length 0.5530E-05 Elastic Modulus at Design Temp. 0.2758E+08 Channel - TEMA : C-Mn-Si, 1-1/4Cr-l/2Mo & 3Cr-lMo Coeff. of Thermal Exp. at Metal Temp. at Tubesheet 0.5530E-05 Elastic Modulus at Design Temp. 0.2758E+08 Tubes - TEMA : Plain Carbon Stl & C-Mn Stl. Coeff. of Thermal Exp. at Metal Temp. along Length 0.6076E-05 0.2708E+08 Elastic Modulus at Design Temp. TubeSheet - TEMA : C-SI, C-1/2Mo & Cr-1/2Mo Coeff. of Thermal Exp. at Metal Temp. at Rim 0.5666E-05 0.2708E+08 Elastic Modulus at Design Temp. Shell - TEMA : C-Mn-Si, 1-1/4Cr-l/2Mo & 3Cr-lMo Coeff. of-Thermal Exp. at Metal Temp. at Tubesheet 0.5530E-05

/ deg F psi / deg F psi / deg F psi / deg F psi / deg F

Results for ASME Fixed Tubesheet Calculations, Per 1998 Revisions, Original Thickness : Results for Step 1

rho rho rho d* d* d* ao ao

ao

Pedd / h 0.0000/ 4.0000 0.0000 ( must be 0

rho

<=

1

MAX ( dt 2tt'" ( Et/E ) ( St/S ) ( rho), dt - 2tt) MAX ( 0.7500-2* 0.0750*( 0.271E+08/ 0.271E+08)* ( 11400/ 19680 )"'( 0), 0.7500-2* 0.0750 ) 0.7500 in. ro + dt / 2 18.0000 + 0.7500 / 2 IB.3750 in.

pstar pstar pstar mustar mustar mustar

9-22

<~

p I sqrt{ 1 - 2*ro*UL/(Pi '* ao~2) ) 0.9375 / sqrt( 1 - 2* 18.0000* 0.0000/ ( 3.142 * 18.3750~2) ) 0.9375 In. 1 - dstar / pstar 1 - 0.7500 / 0.9375 0.2000

Ks Ks Ks

as / ao 20.5625 / 18.3750 1.1190

Kc Kc Kc

ac / ao 20.5000 / 18.3750 1.1156

xt

1

(

- Nt*{ (dt - 2*tt) / (2*ao) )'2



ASME Tubesheets

CO~DE Engineering Software PVElite 4.00 Licensee: CO~DE Inc., Local White Lock FileName; Seminar --------------------------------------- Page 96 ASMETube Analysis: EX-2 FIXED COM Item: 1 9:51a Sep 21,2000

xt xt

1 - 1511""( ( 0.7500 - 2"- 0.0750)/(2"- 18.3750) 0.5972

xs xs xs

1 - Nt*( dt/(2.ao) l"2 1 - 1511*( 0.7500/(2"- 18.3750) )"2 0.3707

)"2

Results for Step 2 Ks Ks Ks

Pi * ts( Ds + ts )Es/t 3.14"- 0.3750( 40.7500 + 0.3750 ) 0.276£+08/ 96.000 13919066.0000 psi * in.

Kt Kt Kt

Pi"- tt( Dt - tt )Et/L 3.14"- 0.0750( 0.7500 44863.4180 psi * in.

Kst Kst Kst

Ks / ( Nt ,.- Kt ) 0.139E+08 / ( 1511 * 44863.4180 1 0.2053

s 0;

s s c =

c c

0.0750) 0.271E+08/ 96.000

( ( 3 * (l rnus A2) )A O • 25 ) / (Os + ts)/2 * ts ) .... 0.5 ( ( J * (l - 0.30"2) )"0.25 ) / ( { 40.7500 + 0.3750)/2. * 0.3750) .... 0.5 0.4629 (= 0.0 if Gasketed 1 (

( ( 3 * (1 - rouc""2l )""0.25 ) / ( (Dc + tc)/2 • tc lA O. 5 ((3"- (1 - 0.30""2) )""0.25 ) / ( ( 40.7500 + 0.3750l/2. * 0.3750)""0.5 0.0000 (= 0.0 if Gasketed )

Results for Step 3 : J J

J

1 / ( 1 + Ks/Kj ) 1 / ( 1 + 0.139E+08/ 0.0000 ) 1.000000000 ( = 1.0 if No Exp. Joint)

Lambdas Lambdas

,

(

Lambdas Lambdat Lambdat

HeaL Exchanger Tubesheets

(

rnus*as/ts + Es/Et * ( (rnut*dt)/(2"-tt) + ao""2"- xs/( Nt * tt "-(dt-tt) ) 0.3* 20.5625/ 0.3750 + 0.276E+08/ 0.271E+08 ,.( ( 0.3* 0.7500)/(2* 0.0750) + 18.3750""2 • 0.3707/ ( 1511 ,.. 0.0750 ,.. ( 0.7500 - 0.0750) ) 19.6441 rnut*( dt - 2*tt )/( 2*tt ) + ao A2 * xt/( Nt * tt *(dt-tt) ) 0.3*( 0.7500 - 2* 0.0750 l/{ 2* 0.0750 ) + 18.3750....2 * 0.5972/( 1511 * 0.0750 * ( 0.7500 - 0.0750))

9-23

ASME Tubesheels


COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------_____________ Page 97 ASMETube Analysis: EX-2 FIXED COM Item: 1 9:51a Sep 21,2000 Lambdat

=

gamma gamma

{alphatm*( Ttm - Tarnb ( 0.608E-05*( 0 - 70) 96.00 -0.0037 in.

gamma Qe

=

Qe

Qe

3.8361 - alphasm*( Tsm - Tamb ))* L 0.553£-05*( 0 - 70)"

J*gamma/L + J*Lambdas " (Ps/Es) _ (Pt/Et) * (J*Lambat + Et"as/ (2"Es*ts») PS*(1~J)/2 * hj* (Ds + hj)/(Es*ts*Os) 1.0000* -0.0037/ 96.00 + 1.0000* 19.6441*{ 250.00/ 0.276E+08) ( 525.0/ 0.271E+08)*( 1.0000* 3.8361 + 0.271E+08* 20.5625/ (2* 0.276E+08* 0.3750») - 250.0*(1 - 1.0000)/2 * 0.00* ( 40.7500 + 0.00)/( 0.276E+08* 0.3750* 40.7500) -0.000456451

Results for Step 4 A ( E*/E ) * ( 1 nu 2)/( 1 - nu*A 2 ) ( 0.1553 )" ( 1 - 0.3000 A2 )/( 1 _ 0.4586*A 2 0.1790

eta eta eta

Xa Xa Xa v

= =

V

V

=

({24 * (1 rnu*"'2) "Nt * Et* tt+ (dt - tt)" aO"'2) / A ( E* " L * H 3 ) )A O . 25 A (24 *(1 0.459 2)* 1511* 0.271E+08" 0.0750" (0.7500 0.0750)* 18.3750 A2)/( 0.421E+07 * 96.00 * 4.000 A3 ))A O . 25 4.7588

A A 2/(E * h 3)* ( s" ts 3* Es* as/(1 _ rnus A2)* (1 + s*h + SA2 * h A2 /2) + c* tc A3* Ee* ac/ (1 - rouc"'2)* (1 + c*h + c A2* h A2 /2) ) A 2/( 0.271E+09* 4.0000 3)* ( 0.463* 0.3750 A3 * 0.276E+08* 20.5625/ A (1 - 0.300 2)*(1A + 0.463* 4.0000 + 0.463 A2* 4.0000 A2 /2) + ( 0.000* 0.3750 3 * 0.276E+08* 20.5000/ (1 - 0.300"'2)*(1 + 0.000* 4.0000 + O.OOOA Z * 4.0000 A2 /Z) 0.08015467Z

gamastars gamastars gamastars gamas gamas gamas

A (Ks 2 - 1) * (Ks -1)/4 - gamastars ( 1.119"'2 - 1) * ( 1.119 - 1)/4 - 0.OZ21 -0.014545569

gamastarc gamastarc gamastarc

9-24

A SAZ * Ats 2 * Ks"'3 *(1 + s*h)/(6*(1 _ rnus"'Z» A 0.463 Z * 0.3750 Z * 1.119 A3* (1 + 0.463* 4.0000)/ ( 6* ( 1 - 0.300 A2 ) ) 0.022053536

CAZ * te"'2 " KC"'] *(1 + c*h)/(6*(1 _ rnuc A2l) A 0.000A 2 " 0.3750 2 * 1.116"'3* (1 + 0.000* 4.0000)/ ( 6* ( 1 - 0.300 A2 ) ) 0.000000000


(

ASME Tubesheels


COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local_ White Lock FileName; Seminar --------------------------------------- Page 98 ASMETube Analysis: EX-2 FIXED COM Item: 1 9:51a Sep 21,2000 (Kc A 2 _ 1) w (Kc -1)/4 - (Kc A 3 Ks)/2 ( 1.119 A 2 - 1) w ( 1.119 - 1)/4 ( 1.116 A 3 - 1.119)/2 - 0.0000 . -0.005373001

gamac gamac gamac

gamastarc

Results for Step 5 : Values from Table AA-2.4 0.013938 Zd ZV 0.043256 Zm 0.296594 (1 - rnu A 2)!eta * ( In( A/{2*ao) ) ~ V) (1 - 0.3000 A 2)/ 0.1790 * (In( 47.65/(2* 18.3750) ) 0.0802 ) 1. 727864861 Q1 Q1 Q1

(Ks - 1 - * Zv) / ( l + * Zm) (1.1190 - 1 - 1.7279* 0.0433)/(1 . ,. 1.7279* 0.2966) 0.029294359

Tr Tr Tr

(T + Ts + Tc)/3 (0.0 + 0.0 + 0.0) 0.0000 F

T*s T*s T*s

(Ts + Tr)/2 (0.0 + 0.0) / 2 0.0000 F

T""c TIC T*c

(Tc ~ Tr) /2 (0.0..,. 0.0) / 2 0.0000 F

Pstars Pstars Pstars Pstarc Pstarc

(

Pstarc Q2 Q2

Q2

/ 3

Es*ts/as * ( alphas * (TIs Tamb)alpha ..,. (Tr - Tamb) ) 0.276E+08* 0.3750/ 20.5625 * ( 0.553E-05 "" ( 0.0 - 70.0) 0.567E-05 * ( 0.0 - 70.0) ] 4.8049 psig Ec*tc/ac * ( alphac "" (Tirc Tamb)alpha * (Tr - Tamb) ) 0.276E+08* 0.3750/ 20.5000 * [ 0.553E-05 * ( 0.0 0.567E-05 * ( 0.0 - 70.0) ) 4.8196 psig

70.0)

-

(ao A 2*( Pt*gamac + Pstarc*gamastarc ~ Ps""gamas pstars,o,gamastars) + W*gamab!(2*Pi) ) ! (1 + *Zm) (18.3750 2*( 525.00* -0.00537 + 4.82* 0.00000..,. 250.00* -0.01455 4.80* 0.02205) + 782983.9* -0.12245/ (2* 3.14) ) / (1 + 1.7279* 0.2966) -11549.908203125


(

~

A

9-25

ASME Tubesheets


COADE Engineering Software PVElite 4.00 Licensee: COADE Inc .• Local White Lock FileName: Seminar --------------------------------------ASMETube Analysis: EX-2 FIXED COM

Item:

1

Page 99

9:51a Sep 21,2000

Results for Step 6 : QZl

(Zd + Ql*Zv) /2 "Xa"'4

Qzl

(0.01394 + 0.02929* 0.04326)/2 * 4.75884"'-4

Qzl

3.8992

Qz2 Qz2 Qz2

(Zv + Ql*Zm) /2 *Xa 4 (0.04326 + 0.02929· 0.29659)/2 * 4.75884 ....4

A

13.3203 A

Ux

(Zv +

Ux

(0.0433 + ( 1.1190 - 1)* 0.2966)* 4.75884"'4/2

Ux

20.1466

Pe

(Ks -

1) *Zm) ·Xa 4/2

2/ao"'2" (as*Es*ts*Qe - J*Kst*Q2*Ux

O.5*(Ps - Pt)*ao"'2*(Ks"'2 - 1»/(1 + J*Kst*

Pe

(Qzl + (Ks -I)*Qz2) ) 2/ 18.3750"'2 * ( 20.5625* O.276E+OB* 0.3750* -0.0005 1.0000* 0.2053* -11549.91* 20.15 - 0.5*( 250.00 - 525.00)* 18.37s0~2*( 1.119~2 - 1)/(1 + 1.0000* 0.2053*{ 3.899 + ( 1.119 - 1)* 13.320) ) -104.6997 psig

Q3 Q3 Q3

Q1 + 2*Q2/{pe*ao A2) 0.02929 + 2*-0.115E+Os/( -104.700* 18.3750 A2) 0.6827

Pe

Results for Step 7 : Curve Value from Table AA-2.4 Fro '" 0.360388 hmin hmin

h (For pressure plus thermal case and thermal only case) 4.0000 in.

The Tubesheet Bending Stress - original Thickness: Sigma (1.5*Fm/mustar) * (2*ao/hmin)A 2 * Pe Sigma (1.5* 0.3604/ 0.2000) * (2* 18.3750/ 4.0000)A 2 ... -104.70 Sigma -23887.5625 psi The Allowable Tubesheet Bending Stress ; Sigma allowed 1.5 * Omega * S Sigma allowed 1.5 * 2.6667 * 19680.00 Sigma allowed 78720.0000 psi The Tubesheet Bending Stress - Final Thickness: Sigmaf (l.s*Fm/mustar) * (2*ao/hmin) AZ * Pe Sigmaf (1.5* 0.1963/ 0.2000) * {2* 18.3750/ 1.9270)A 2 * 147.01 Sigmaf 78719.8125 psi Required Tubesheet Thickness, for bending stress :

9-26


ASME Tubesheels


COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar ----~---------------------------------- Page 100 ASMETube Analysis: EX-2 FIXED COM

Item:

1

9:51a Sep 21,2000

(Including CATS and CATC) H + CATS + CATC 1. 9270 + 0 _0000 + 0.0000 1.9270 in.

H reqd _ bend. H reqd. bend. H reqd. bend.

mu mu

(p - dt) /p (0.9375 0.7500)/0.9375

mu

0.2000

The Tubesheet Avg. Shear Stress - Original Thickness: Tau (1/ (2*mu) ) * { ao/h l • Pe Tau

(1/{2* 0.200)

Tau

-1202.4100 psi

)

* (

18.3750/4.000

1 *

-104.700

The Allowable Tubesheet Shear Stress

Tau allowed Tau allowed Tau allowed

0.8 * S 0.8 * 19680.00 15744.0000 psi

The Tubesheet Shear Stress - Final Thickness: Tauf (1/ (2*mu) ) * ( ao/h ) * Pe Tauf (1/ (2* 0.200) ) * ( 18.3750/ 0.976 ) * 333.879 Tauf 15715.6035 psi Required Tubesheet Thickness, for shear stress : (Including CATS and CATC) H reqd. shear H + CATS + CATC H reqd. shear 0.9759 + 0.0000 + 0.0000 H reqd. shear 0.9759 in. Required Tubesheet Thickness for Given Loadings, (Including CATS and CATC) H reqd. Max{ H_re'L-bend, H~reqd_shear H reqd. Max{ 1.9270, 0.9759) H reqd. 1.9270 in. Results for Step 8 ; Fq Fq Fq

(

The Tube Stress in the outermost Tube row. ( (Ps*xs - Pt*xt) - pe*Fq) / (xt - xs) Sigmato ( ( 250.00* 0.3707 525.00* 0.5972) sigmato ( 0.5972 - 0.3707) Sigmato 4176.6797 psi The Allowable Tube Sigmato allowed Sigmato allowed Sigmato allowed

(

A

(zd + Q3*Zv)*Xa 4/2 A (0.0139 + 0.6827* 0.0433)* 4.7588 4/2 11.1474


-104.70* 11.147)/

Stress omegat * Sot 2 * 11400.00 22800.0000 psi

9-27

ASME Tubesheets


COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------------------- Page 101 ASMETube Analysis: EX-2 FIXED COM Item: 1 9:51a Sep 21,2000 The Force on TubeForce TubeForce TubeForce

the Outermost Tube ; TubeStress * Tube Area 4176.68 * 0.15904 664.2722 lb.

The Allowable Tube Force Per Appendix A VIII 1 Force Allowed 0.1590 * 11400.00 * 0.65 * 1.00 * 1.00 Force Allowed = 1178.5096 lb.

Results for Step 9 Qz2star Qz2star Qz2star

( (Zv + Q3*Zm)*Xa A 4 )/2 ( ( 0.0433 + 0.6827* 0.2966)* 4.7588 A 4 )/2 63.0190

The Shell Membrane stress due to joint interaction A Sigmasm ao 2*Pe/(2*as*ts) + Pt*as/(2*ts) + (Ps - Pt)/2 * ao*(Ks A 2 - 1)/(Ks*ts) Sigmasm 18.3750 A 2* -104.70/(2* 20.5625* 0.3750) + 525.0* 20.5625/(2* 0.3750) + ( 250.0 - 525.0)/2 * 18.3750*( 1.119 A 2 - 1)/( 1.119* 0.3750) Sigmasm = 10582.6562 psi

(

The Shell sending stress due to joint interaction Sigmasb = 12* (1 + s*h/2)* (Es/E)* (s*ts/eta) * A A (ao 3/(h 3*Xa A 4»)* Pe*Qz2star + 1/(1 - rnus A 2)*(Ps - rnus*Sigmasm*ts/as + Pstars)* SA 2 * as A 2 Sigmasb 12* (1 + 0.463* 4.000/2)* ( 0.276E+08/ 0.271E+08)* ( 0.463* 0.3750/ 0.179) + { 18.3750 A 3 / ( 4.000 A 3* 4.7588 A 4))* -104.700* 63.019 + A 1/(1 - 0.30 2)* (250.0 0.30* 10582.7 * A 0.3750/20.5625 + 4.80) * 0.4629 A 2* 20.S62S 2 Sigmasb = -8859.6250 psi SHELL STRESS SUMMATION VS. ALLOWABLE ISigmasm/ + ISigmasbl =< 1.5 * Omega * Ss I 10582 I + I -8859 I =< 1.5 * 2.6667 • 20000 19442.28 must be < or = 80000.0000 Computations completed for ASME Tubesheet configuration b Summary of stress/force comparisons Stress Description Tubesheet bending stress Average Tubesheet shear stress Tube stress in the outermost tube Force on the outermost tube Shell stress due to j t. inter.

9-28

Stress Value 23887.6 1202.4 4176.7 664.3 19442.3

Allowable <~

<~ <~ <~

<~

78720.0 15744.0 22800.0 1178.5 80000.0

psi psi psi lb. psi


(

ASME Tubesheels


COADE Engineering Software PVElite 4.00 Licensee; COADE Inc., Local White Lock FileName: Seminar --------------------------------------i\SMETube Analysis: EX-2 FIXED COM

Item:

SUMMARY of RESULTS for ASHE Tubesheet Calculations HREQ Reqd Tubsheet Thickness + CATS + CATC H Actual Tubsheet Thick. as Given Required Thickness for the Flanged Portion FI'REQ Actual Thickness for the Flanged Portion TF The PV Elite Program,

(

Heat Exchanger Tubeslleets

1

Page 102

9:51a Sep 21,2000

1.9270

in.

4. 0000

in.

2.0204 3.5000

in. in.


9-29

ASME Tubcshects

Pressure Vessel Design and An'alysis - Seminar NOles

>

---,'~

9-30


Pressure Vessel Design amI Analysis - Seminar Notes

Chapter 10: Expansion Joints

FLANGED AND FLUED EXPANSION JOINTS These are also commonly called "thick-walled" expansion joints. They are characterized by having a small number of large bellows with relatively thick walls on the order of 0.25 m. Thick-walled expansion joints are analyzed using "Expansion Joints for Heat Exchangers," by S. Kopp, American Locomotive Company, Alco Products Division, and M.F. Sayre, Union College, Schenectady, N.Y. This method is commonly referred to as the "Kopp and Sayre" method. TEMA has modified the Kopp and Sayre method to reflect the use of plate and shell theory, and has implemented it as paragraph RCB-8. This is also the method that is implemented in the CodeCalc program. Evaluation of the stresses calculated by this method is based on the ASME Code, Section VJII, Division I, Appendix CC, Paragraph CC-3. The formnlas contained in the program are applicable based on the following assumptions: Applied loadings are axial. Torsional loads are negligible. The flexible elements are snfficiently thick to avoid instability. The flexible elements are axisymmetric. The sequence of calcnlation used by the program is as follows:

Expansion Joints

I.

Select a geometry for the flexible clement per RCB-8.21 (user input).

2.

Determine the effective geometry constants per RCB-8.22.

3.

Calculate the flexibility factors per RCB-8.3.

4.

Calculate the flexible element geometry factors per RCB-8A.

5.

Calculate the overall shell spring rate with all contributions from flexible shell elements per RCB-8.S.

6.

Calculate Fax for each condition as shown in Table RCB-8.6. This requires that you run the CodeCalc Tubesheet program to determine the differential expansion and shellside and tubeside equivalent pressures.

?

Calculate the flexible element stresses per RCB-8.?

8.

Compare the flexible element stresses to the appropriate allowable stresses per the Code, for the load conditions as noted in Step 6.

9.

Modify the geometry and rerun the program if necessary.

to-t

Flanged and Flued Expansion Joints


Figure lO-1 shows geomelry for the THICKJNT program. (TEMA Figure RCB-8.2l and RCB-8.22) Both the input geomelry and the equivalent geometry used for the analysis are shown. The discussion of input data below uses the nomenclature shown on this figure. I.

I.

I.

t.

t.

2r~

<:l

C

t,

1

<:l

C

I,

I,

/ ... 0

1.-0

r.-O

r." 0

I,

I,

I,

"

til.

I.)

I,

I,

I, (--"

Ii

(

-

-

TI

_.

.

-Ir

+ I.

.Q

y,

~ Y.

a

~X+

Figure 1--Geomelry for Flanged and Flued Expansion Joints The Kopp and Sayre Method uses continuity considerations alent stresses in the expansion joint.

'0 elastically calculate equiv-

The bellows outside diameter, inside depth, and the radius of the inside and oulside corners are shown in the sketch on the following page. The analysis technique also applies to joints with a square comer (zero radius) at either 'he inside diameter or the outside diameter. When a flanged-only expansion joint or a square expansion joinl is analyzed, the bellows thickness should be the same as the shell thickness, and the web thickness should be the thickness of the annular plate.

10-2

Expansion Joints

Pressure Vessel Design and AnCllysis - Seminar Noles


For fully flanged and flued joints, thc bellows web thickncss is typically a lillie bit less than thc nominal bcllows thickness due 10 thinning in the forming operation. The Kopp and Sayre method analyzcs a square cornercdjoint, and approximatcs radiused corners by substituting the length of a square corner wilh the samc stiffness. Thus Ihe method produces an equivalcntlcngth and span for use in the analysis. Allowable stress for internal and external pressure are based on the ASME Code, Section VIII, Appendix CC, Paragraph CC-3, and is 1.5 limes the basic material allowable stress in Ihe flat pari of the annulus and 3.0 times the basic allowable slress near the corners. Individual stress intcnsitics are based on the algcbraic addition ofthc various cases, including differential expansion pressure, shellside and lubeside design pressurc, and the equivalent pressures due 10 othcr loadings. All ofthesc loads are found in the output of the CodeCale Tubesheet Program. Since thc CodeCale Tubesheel program is separate from the flanged and flued expansion joinl program, a new expansion joint design requires an iterative approach: I.

Select expansion joint geometry using the expansion joint program based only on shcllside and tubeside pressure. Detcrmine the spring constant for this joint.

2.

Run Ihe Tubeshcet program with this spring constant and determinc the required tubesheet Ihickness plus thc differential and equivalent pressures.

3.

Enter the differential and equivalcnt pressures into the cxpansionjoinl program and determine if the joint is adequate from a strcss standpoint. If il is nOI, rcsize the joinl and repeal from Stcp I.

The stress resulting from intcrnal pressure (shellside pressure) lends 10 increase as the joint OD increases, and decrease as the joint thickness increases. The stress due to differ-

cnlial cxpansion and tubeside pressure lends to decrease as joint OD increases and increase as joint thickness increases. Workable designs will adequately balance Ihesc competing demands. The maximum stress range is automatically calculated for each design case by the TEMA procedures. The cycle lifc is bascd on the material category, using the fonnulas given in the ASME Code, Section VIII, Division I, Appendix CC, Paragraph CC-4. This paragraph includes a power law formula for fatigue evaluation as follows:

(

2.17

<

2.2

- ( 14.2Kg Sn

E

)

0.03

b

for series 3XX high alloy steels, nickel-chromium iron alloy, nickel-iron chromium alloy and nickel-copper alloy. 2."

< 2.0 - [ 15K S ) +-0.011 b

for carbon, low alloy, series 4XX, and high alloy steels. N is the number of cycles, the cycle life. Sn is the maximum stress range resulting from the analysis.

(

Expansion Joints

10-3



E b is the elastic modulus for the material at the temperature under consideration.

Kg is the fatigue strength reduction factor. Use a fatigue strength reduction factor of 1.0 if both thc inside and the outside comcr radii are greater than 3 times the joint thickness. If either radius is 0.0 usc a factor of4.0. Ifboth radii arc greaterthan 0.0 but less than 3.0 times the thickness, use a value of2.5.

10-4

Expansion Joints


Pressure Vessel Design and Analysis Seminar Componenl

PRESSURE VESSEL DESIGN AND ANALYSIS SEMINAR COMPONENT DESIGN PROBLEM . Flanged and Flued Expansion Joint Analysis Evaluate the stresses in the following expansion joint: 10-1\ (3048mm) tit shell, 0.375-in. (9.5 mm) thickness, 25.25-in.(641 mm)!D, SA516-70 at 200°F (93°C) 34-in. 00, 33-in.!D Expansion Joint, 0.4375 thick, SA-240 316 SS at 200°F At the joint!D there is a I-in. (25.4 mm) straight flange and 1.5-in. (38.1 mm) radins bend. At the joint 00 there is a l-in.(25.4 mm) straight flange and 1.5-in. (38.1 mm) radius bend. There is a short I-in. (25.4 mm) outer cylinder connecting the two halves of the expansion joint. The outer cylinder thickness is 0.5 in. (12.7 mm), SA-240 316 SS. From the Tubesheet Program you learn the following Information: PO = 515.9200 psi (3.55 Nmm2) PS ~ 200.0000 psi (1.372 Nmm2) PSP ~ -38.7500 psi (-0.267 N/mm2) PT ~ 100.0000 psi (0.689N/mm2) PTP = 86.9000 psi (0.559 Nmm2) Notes:

(

Questions: What is the minimum cycle life for the bellows?

_


(

Expansion Joints

10-5

Pressure Vessel Design and Analysis Seminar Componelll Design Problem

Pressure Vessel Design and Analysis-

COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., ,Local White Lock FileName : Seminar --------------------------------------- Page 83 Thickjnt Analysis: EX-4 JOINT 1 Item: 1 9:51a Sep 21,2000

Input Echo, Thick Joint Item

I,

Description: EX-4 JOINT 1

Design Temperature for Expansion Joint TEMP Inside Diameter of Expansion ~oint RID Outside Diameter of Expansion Joint ROD Wall thickness of Expansion Joint TE Corrosion Allowance for Expansion ~oint CAE Material for Expansion Joint Operating Allowable Stress, Expansion Joint SOE Ambient Allowable Stress, Expansion Joint SAE Inside Diameter of Shell Wall thickness of Shell Corrosion Allowance for Shell Length of Shell Cylinder Material for Shell Operating Allowable Stress, Shell Ambient Allowable Stress, Shell Length Radius Length Radius

of of of of

Straight Knuckle, Straight Knuckle,

Flange, Joint ID Joint ID Flange, Joint 00 Joint 00

Thickness of Outer Cylinder Corrosion Allowance for Outer Cylinder Length of Outer Cylinder Outer Cylinder Material Name Operating Allowable Stress, Outer Cylinder Ambient Allowable Stress, Outer Cylinder Differential Expansion Pressure Shells ide Design Pressure Shellside Prime Design Pressure Tubeside Design Pressure Tubeside Prime Design Pressure Analyze Analyze Analyze A.nalyze Analyze Analyze Analyze

for for for for for for for

G

TS CAS

RLI SOS SAS FA

200.0000 33.0000 34.0000 0.4375 0.0000 SA-240 316 20000.00 20000.00 25.2500 0.3750 0.0000 40.0000 SA-516 70 20000.00 20000.00

RA

1. 0000 1. 5000

FB RB

1. 0000 1.5000

TO CAO RLa

0.5000 0.0000 1. 0000 SA-240 316 20000.00 20000.00

SOO SAO PD PS PSP PT

PTP

Differential Expansion Shellside Pressure Tubeside Pressure Shellside + Tubeside Pressure Shellside + Differential Expansion Tubeside + Differential Expansion Shells ide + Tubeside + Differential

515.9200 200.0000 -38.7500 100.0000 86.9000

F

in. in. in. in. psi psi in. in. in. in. psi psi in. in. in. in. in. in. in. psi

psi psig psig psig psig psig

y y y y y N

(

N

Results per TEMA 1999 TEMA Paragraph RCB B.22 Effective Geometry Constants: Ta IF(RA<>O,Te,Ts) 0.4375 Tb IF (MoO, Te, To) 0.4375 A ( G + TA ) / 2.0 12.B437 B (ROD-TB)/2 16.7B12

10-6

in. in. in. in.

Expansion Joints

Pressure Vessel Design ilnd Analysis - Semin
Pressure Vessel Design and Analysis Seminar Component

COADE Engineering Software PVEl-ite -:1.00 Licensee: COADE Inc., Local White Lock PileName ; Seminar --------------------------------------- Page 84 Thickjnt Analysis: EX-4 JOINT 1 Item: 1 9:51a Sep 21,2000

I B - A) FA + RA+ TE I 2.0 FB + RB I 4.0 + TE RADASH " RA + o. 5 TE RBDASH " RB + 0 .5 • TE YA MIN I RLA + RLI, 2 YB "" MIN I RLB + RLO, 2

H RLA RLB

3.9375 2.7188 2.7188 1.7l88 1.7188 4.7409 3.7188

I 2.0

•

•

SQRT SQRT

A B

•

TA TB

in. in. in. in. in. in. in.

TEMA Paragraph ReB 8.30 Element Flexibility Factors: Elastic Modulus for Expansion Joint Material EE 0.273E+08 psi Elastic Modulus for Shell Material ES O. 290E+08 psi

Elastic Modulus for Outer Cylinder MatI. EA EB Ba Bb Da Db De Wa

RJIA RJ2A ZZA

RKOA RKIA RK2A RK3A wb

RJIB RJ2B ZZB

RKOB RKIB RK2B RK3B

1.285 / SQRT( A * TA 1. 285 / SQRT! B * TB 0.0916 • EA • TA**3 0.0916 EB * TB**3 0.0916 • EE * TE .... 3

COS ( WA RJIA**2 SINH{ COSH { SINH( COSH! BB • YB

• SINH { WA * COSH ( WA + RJ2A**2 WA ) + SIN ( WA WA ) + COS ( WA SIN{ WA WA ) COS { WA WA )

0.273E+08 psi 0 .273E+08 psi 0 .273&+08 psi 0.5421 l/in. 0.4742 l/in. 209407.53 in. lb. 209407.53 in. lb. 209407.53 in. lb.

IF (RA <> 0, EE, ES) IF (RB <> 0, EE, EO)

BA • YA SIN( WA

EO

I RKOA

2.5700 3.5134 -5.5266 42.8877 7.0356 0.8144 0.8462 1.0535

I RKOB I RKOB I RKOB

1.7636 2.7785 -0.5753 8.0511 3.8124 0.7373 0.4851 0.8378

I RKOA I RKOA

.

SIN( WB ) SINH{ WB COS( WB ) '* COSH{ we RJIB*"'2 + RJ2B"'·2 SINH{ we ) + SIN( WB COSH( we ) + COS ( WB SINH{ WB ) SIN ( HB COSH( WB ) COS I WB

TEMA paragraph RCB 8.31 Cylinder Flexibility Factors: CIA (RLA / SQRT ( A • TA ) ) (TS / TA ) C2A C3A (EA / ES ) C4A -0.364661+{0.338172/C2A)-(O.0366351/C2A**2 C5A -1.06871+(1.01164/C2A)-(0.122627/C2A··2) C6A 0.0696709+(1.76415*C2A)-(S.46103"'C2A"''''3) C7A -0.142734+{0.918656*C2A)-(2.00749"'C2A**3) C8A (C5A/C1A'*·2-C6A/CIA"'*3+C7A/CIA*"'4-C4A) /(C3A H O.2) EAA 2.718"''''C8A CIB C'B

(

Expansion Joints

RLB / SQRT( B • TB ) ) TO / TB )

1.1469 0.8571 0.9414 -0.0200 -0.0554 -1. 8572 -0_:6195 0.8612 2.3658 1.0034 1.1429

10-7



COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar ------------------------------------ Page 85 Thickjnt Analysis EX-4 JOINT 1 Item: 1 9:51a Sep 21,2000

C3B C4B C'B CGB C7B C'B

EBB

( EB I EO I

1.0000 0.0017 -0.0777 -0.0908 -0.0296

(3.3731-1.707962*C2B+0.226216*C2B**2)/1000 -0.403287+0.3200J7*C2B-0.0307508*C2B**2 -0.684978+0.582549*C2B-0.0547812*C2B**2 -0.201334+0.168201*C2B-0.015728*C2B**2 (CSB/C1B**2-C6B/C1B**3+C7B/C1B**4-C4B) / (C3B H O.2) 2.718**C8B

-0.0183 0.9819

TEMA paragraph RCB 8.40 Element Geometry Factors: RYl (EAA* (RK3A-RK2A H 2/ (2*RK1A)) / (DA*BA)) RY2 (EBB* (RK3B-RK2B H 2/ (2*RK1B)) / WB*BB)) C {A**2/ (B**2 - A**2») D

( B /

RXI RX2 RX3

-A * C * (0.769 + 1.428 * D * D) / DE 2.2 * A * C * D * D / DE -A*A*(1.538+LN(D)*{2+C*(2+3.714*D*D))) / (4*OE) ( -2.2 * B * C ) / DE {B * C * (0.769 * D * D + 1.428) / DE) (-A*B*{1.538 + 5.714*C*ln(D)) / (4*DE») (RX1 RY1) * (RX5 + RY2) - (RX2 * RX4) RX3 * RXS RX3 * RY2) / XBOT (RX2 * RX6 (RXJ * RX4 - RX1 * RX6 + RX6 * RY1) / XBOT

RX4 RX'

RXG XBOT RX7 RX' RQ1 RQ2 RQ3 GG GST RM1 RM2 RM3

12.7948 6.7051 1.4142 1.3066 -278.1458 325.7556

A )

-1.0294 -249.3214 310.6066 -951.7537 -0.0111 -1.4941 1.8255

0.38S * A * A + 1.429 * C * B * B * In{D) (-0.385 - 1.429 * C * In(D)) * B * B .2S*A*B*B*(1.269/fC*D*D)+3.714*C*ln(D)"'2) (A/B) (GG**4 * In(GG) / (1-GG**2)) 0.51 - 0.635 * GG**2 + GST 0.635 (1 - GG**2 ) + GST 2.357 * GG**2 + 3.714 * GST

I 10-"6 I 10"'6 I 10"'6 I 10"'6 I I I I I

10"'3 10"'6 10"'6 10-"6 10"6

215.6894 -260.5991 814.9036 0.7654 -0.2215 -83.4889 I 10"'3 41.5111 I 10"'3 0.5580

TEMA Paragraph RCB 8.50 Stiffness Multiplier: TEMA Paragraph ReB 8.51 Y Geometrical Parameters: YA IF( YA/G >00 0.075) YA 1.0 YB 00 IF( YB/G >'" 0.075) YB = 1.0

1. 0000 1. 0000

TEMA Paragraph RCB 8.52 Equal Knuckle Radii Case {RA M = Fig RCB-8.51 ( RADASH, H, TE ) ALPHA '" IF{G/TE < 160) ALPHA '" 4.3*(G/TE)"'-0.287 K '" ALPHA * M * YA * YB

RB) : 1.6949 1.3427 2.2758

TEMA RCB 8.60 Flexible Element Stiffness : SJ 2*PI*A*DE*K/{ RX7*RQl + RX8*RQ2 + RQ3 SJF", {1.0 / (1.0/ SJ + 1.0 / SJ))

2272059.250 lb./in. 1136029.625 lb./in.

Analysis of Differential Expansion TEMA Paragraph RCB 8.70 Induced Axial Force: Tubeside Equivalent Pressure for this case, PIC

10-8

0.000 psig

Expansion Joints


Pressure Vessel Design and Analysis Seminar Componcnl

COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------------------- Page 86 Thickjnt Analysis: EX-4 JOINT 1 Item: 1 9:51a Sep 21,2000 Shellside Prime Pressure for this case, PSPC Differential Expansion Pressure for this Case, PDC Shellside Pressure for this Case, PSC Equivalent Pressure, PSS PIC + PSPC - PDC Induced Axial Force, FAX = A + PSS / 2.0 TEMA Paragraph RCB S.80 Flexible Element Moments THA «(Ps+S"3) / (8·DE) ) • (-2+GG+RM2-RM3/GG-GG"J/2-2 I GG"J·LN(GG)) THB (Ps·B++J)+{-2·RM2-RM3+0.5-GG++2)/(S+DE) ZA (pS·A++2-0.J·1'.+FAX)/(EA·TA) (PSIS "'2- 0.3 + (A+FAX+ { (B ....2 -A "2) /2) IPS 1) ZB / (ES+TB) MTlA (RXS+RY2)·(-THA-FAX·RX3-BA·ZA) MT21'. RX2*(FAXIRX6+THB-(BS*RK2B+ZB/RK1B») RMA (MTlA+MT21'.)/XBOT MT1B (RYl-RX1) + (THB+FAX*RX6-(BB*RK2B*ZB/RKlB)) MT2B RX4 * (FAX·RX3+THA+BA*ZA) RMB (MTlB+MT2B)/XBOT TEMA paragraph RCB 8.82 Flexible Element Moments RAIl -C*RMA+C*O*O·RMB+0.65*A+C*FAX*ln(GG) RA12 PSS*(0.32S.RM2*B*B+0.412S*A*Al RA1 (RA1D1-RAI2) RA21 C*RMA-C*RMS-0.6S.A*C*FAX*ln(GG) RA22 0.087S*RM3*PSS*B*B RA2 B*B.(RA21+RA22) RA3 O. 206*PSS RA4 0.65*A*(FAX-0.S*A*PSS)

0.000 515.920 0.000 -515.920 -3313.174

0.0000 0.0000 1. 0688

psig psig psig psig lb. in.

I 10"'6 I 10.... 6 I 10.... 3

1.0688 I 10.... 3 -1.0823 I 10"'3 1. 0271 I 10.... 3 4976.483 in.lb./in. 917.3J31 I 10 .... 6 -850.4341 I 10 ....6 -6026.405 in.lb./in.

-11.1267 0.0000 -11.1267 5.1003 0.0000 1. 4363 0.0000 -27.6598

•

10"3

I 10"6 I

• I

•

10"3 10 ....3 10.... 6 10.... 6 10.... 6 10.... 3

Stress in Expansion Joint Flange: SBX(R) = (6.0 / TE+*2) * (RAI + RA2/(R*R) + RA3*R*R + RA4 I ln{R/S») Stress Suounary: Location: (A) 12.8437 in. , 155997.55 psi Location: (B) 16.7812 in., -188909.36 psi TEMA Paragraph RCB 8.83 Cylindrical Element Stresses For the DEA BIA B2A

SMYA

Inner Cylinder at X = YA +RRA*(PSS*RRA-0.3*F2A)/(EA*TYA) (l/ZZA)* {RJ21'.IRMA/{2*8A"'2*EAA*OA)-RJIA*DEA (1/ZZA)*(-RJIA*RMA/(2*BA*BA*EAA*OA) -RJ2A*DEA) +BA* (YA-YA) 811'.*SIN(U1A)*SINH(UI1'.)+B21'.*COS{UIA) *COSH(Ul1'.) EAI(OEA+U2A)/RRA

-1.2395 I 10"'J 16.003 psi

For the DEA BIA B2A

Inner Cylinder at X = LA RRA*{PSS*RRA-0.J*F2A)/(EA*TY1'.) (1/ZZA1*{RJ2A*RMA/(2*BA"'2*EAA*OA)-RJ1A*OE1'. (1/ZZA)·(-RJIA·RMA/(2*BA"'2 I EAA*OA)-RJ2A*DE

1.2470 -2.J046 -1. 2395

UlA U2A

(

Expansion Joints

1.2470 -2.3046

I 10.... 3 I 10.... 3

-1.2395 0.0000

I 10.... 3 I 10"'6

I 10"'3 I 10"'3 I 10.... 3

10-9


Pressure Vessel Design and Analysis -

CQADE Engineering Software PVElite 4.00 Licensee: COADE Inc., LOcal White Lock

FileName: Seminar

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

Thickjnt Analysis: EX-4 JOINT 1

DIA D2A SMYA

Item:

1

BA*(YA-LA) BIA*SIN(UIA) *SINH (UIA) +B2A*COS (UIA) *COSH (U EA*(DEA+U2A)/RRA

Page 87

9:51a Sep 21,2000

-0.9317 -3.0679 / 10.0. 3 -3870.467 psi

For the Outer Cylinder at X DEB BIB B'B DIB D'B SMYB For the DEB BIB B2B UlB D2B SMLB

= YB RRB*(PSS*RRB-O.3*F2Bl!(EB*TYB) (1/ZZB)*(RJ2B*RMB!C2*BS"'2*EBS*DB)-RJlS*OEB (1/ZZB)*(-RJlS*RMB/(2*Ss"'2*E8B*DB)-RJ2S*DE

BB*(YB-YB) BIB*SIN(UIBj*SINH(UIB)+B2B*COS(UIB) ·COSH(UIB) EB·(DEB+U2B)/RRB

1. 0688

/

10"'3

4.2870 / lOA) 22.5635 / lOA) 0.0000 / 10"'6 A

22.5635 / 10 3 38445.434 psi

Outer Cylinder at X '" LB RRS*(PSS*RRB-0.3*F2B)/(ES*TYB) (1/ZZB)·(RJ2B*RMB/(2*BB ....2*EBB*DB)-RJIB*DEB (1/ZZB)*(-RJIB*RMB/(2*BB*BB*EBB*DB)-RJ2B*O BB*(YB-RLB) BIB*SIN(U1B)·SINH(U1B)+B2B*COS(U1B) *COSH(U1B) EB*(DEB+U2B)/RRB

1.0688 / 10....) 4.2870 / 10 .... 3 22.5635 / 10 .... 3 -0.8151 23.7395 / 10 .... ) 40358.613 psi

TEMA Paragraph RCB 8.84 Maximum Cyclic Stresses Evaluated per ASME A-99 Appendix CC, Paragraph CC-3(C): For the Inner Cylinder: SCLA = ABS(6.0*RMA/ (T*T)) + ABS(F2A/T) 163570.484 psi 3.2217 * 10 .... 3 RNA= 47.11*EXP«(ln(SCLA*28.3E6/EB)-14.12)**2/1.023 For the Outer Cylinder: SCLB = ABS(6.0*RMB/ (T*T) + ABS(F28/T) 194705.406 psi 1.6345 * 10....3 RNB= 47.11*EXP«ln{SCLB*28.3E6/EB)-14.12)*·2/1.023 Analysis of Shellside Pressure TEMA Paragraph RCB 8.70 Induced Axial Force: PIC 0.000 -38.750 PSPC= PSC '" 200.000 -38.750 PSS '"

PDC FAX

TEMA pa-ragraph RCB 8 .80 Flexible Element Moments THA -0.438 THB -0.410 ZB 2A 2.843 /10 .... 3 4.503 /10 ....3 A MT1A 57.324 /10 .... 6 MT2A -56.875 /10 6 RMA A /10 .... 6 MT18 -50.797 45.041 /10 6 MT2B RMB TEMA Paragraph RCB 8.82 Flexible Element Moments 14.369 *10 3 HAl RAIl 2.094 *10 3 RA12 2.750 *10 3 HA2 RA21 -1.576 ·10 3 RA22 A RA3 41.200 RA4 -12.800 *10 3 Stress in Expansion Joint Flange: SBX(R) '" (6.0/ TE·*2) • (RAI + RA2/(R*R) Stress Summary:

10-10

+

0.000 -248.848

-40.404 518.487

( A

-12.275 *10 3 0.331 0010 .... 6

RA3*R*R + RA4*ln(R/B»

Expansion Joints

\

Pressure Vessel Design and Analysis Seminar Component


COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------------------Thickjnt Analysis

Location:

(A)

Location: Location:

(R) (B)

EX-4 JOINT 1 12.8437 14. 0081 16.7812

Item:

1

-1621.83 -6070.31

in., in .• io.,

Page 88

9:51a Sep 21,2000

15717 _07

psi psi psi

TEMA Paragraph RCB 8.83 Cylindrical Element Stresses For the Inner Cylinder at X = YA DEA 3.316 /10 ....3 BIA -253.799 /10"'6 438.722 /10"'6 OlA 0.000 /10"'6 U2A

SMYA

438.722 /10"6 7.982 *10"'3

For the Inner Cylinder at X = LA DBA 3.316 /10"'3 BIA -253.799 /10"'6 UIA -0.932 U2A 165.245 /10"'6

B2A SMLA

438.722 /10"'6 7.400 *10"'3

For the Outer Cylinder at X DEB BIB 4.503 /10"'3 UlB U2B 0.000 /10"'6

YB -1.955 /10"'3 -1.613 /10"'3

B2B SMYB

-1.613 /l0"'3 4.701 *10"'3

For the Outer Cylinder at X DEB 4.503 /10"'3 BIB UIB U2B -0.815

LB -1.955 /10"'3 -2.787 /10"'3

B2B SMLB

-1 .613 /10"'3 2 .792 *10"'3

B2A

TEMA Paragraph RCB 8.64 Maximum Cyclic Stresses Evaluated per ASME A-99 Appendix CC , paragraph CC-3fc): For the Inner Cylinder: SCLA RNA 1. 635 *10"'3

1.000 *10"'6

For the Outer Cylinder: SCLB 17.406 *10"'3 RNB

1.000 *10"'6

Analysis of Tubeside Pressure TEMA Paragraph RCB 8.70 Induced ;r>.xial Force: 0.000 PIC 13.100 PSPC= PSC ':0 0.000 13.100 PSS = TEMA paragraph RCB 8.80 Flexible Element Moments 0.000 /10"'6 THB THA 0.000 /10"'6 -27.140 /10"'6 ZA -27.140 /10"'6 ZB -26.080 /10"'6 RMA MT1A 27.482 /10 .... 6 MT2A 21.594 /10"'6 RMB MT1B -23.292 /10 .... 6 MT2B

(

TEMA Paragraph RCB 8.82 Flexible Element Moments 0.000 /10"'6 RAI RAll 282.523 RAl2 0.000 /10"'6 RA2 RA21 -129.504 RA22 RA4 702.325 RAJ 0.000 /10"'6

PDe

0.000 84.127

FAX

-126.361 153.020

282.523 -36.470 "'10"'3

Stress in Expansion Joint Flange: SBX(R) = (6.0/ TE*"'2) * (RA1 + RA2/(R"'R) + RA3*R*R + RA4*ln(R/B») Stress Summary: -3961.02 psi Location: (A) 12.8437 in. , 4796.70 psi Location: (B) 16.7812 in. ,

Expansion Joints

(

10-11



COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------------------Thickjnt Analysis: EX-4 JOINT 1

TEMA

Item;

1

Page 89

9:51a Sep 21,2000

Paragraph RCB 8.83 Cylindrical Element Stresses

For the Inner Cylinder at X DEA -31.663 /10 .... 6 BIA UIA 0.000 /10"'6 U2A For the Inner Cylinder at X DEA -31.663 /10"'6 BIA UIA -0.932 U2A

Y~

58.517 /10"'6 31.472 /10"'6

B2A SMYA

31.472 /10"'-6 -0.406

B2A

31.472 /10"'6

LA

58.517 /10"'-6 77.899 /10""6.

98.277

SMLA

For the Outer Cylinder at X YB -27.140 /10"'6 DEB BIB -108.853 /10"6 UIB 0.000 /10"6 U2B -572.921 /10"'6

SMYB

-572.921 /10 .... 6 -976.188

For the Outer Cylinder at X = LB -27.140 /10 .... 6 DEB BIB -108.853 /10"-6 UIB -0.815 U2B -602.782 /10.0. 6

B2B SMLB

-572.921 /10"-6 -1.025 ... 10 .... 3

B2B

TEMA paragraph RCB 8.84 Maximum Cyclic Stresses Evaluated per ASME A-99 Appendix CC , Paragraph CC-3(c): For the Inner Cylinder: seLA 4.153 ... 10 .... 3 RNA

1.000 ... 10.... 6

For the Outer Cylinder: SCLB 4 _944 ... 10.... 3 RNB Analysis of Shellside + Tubeside Pressure TEMA Paragraph RCB 8.70 Induced Axial Force: PIC 13 .100 -38.750 PSPC= PSC = 200.000 -25.650 PSS = TEMA paragraph RCB 8.80 Flexible Element Moments THA -0.438 THB -0.410 2A 2.815 /loA] ZB 4.476 /10 .... 3 -82.955 /10.... 6 MT1A 84.806 /10.... 6 MT2A RMA MTIB -74.089 /10 ....6 MT2B 66.635 /10.0. 6 RMB

-166.764 671. 506

TEMA Paragraph RCB 8.82 Flexible Element Moments

RAll RA2l RA3

2.377 "'10.0. 3 -1.705 "'10"-3 41. 200

RA12 RA22 RA4

14.369 ... 10 .... 3 2.750 '""10.0. 3 -12.098 "'10"-3

Stress in Expansion Joint Flange: = (6.0/ TE...... 2) ... (RAI + RA2/(R"'R) Stress Summary: Location: (A) 12.8437 in. I Location: (R) 13.6136 in. , Location: (B) 16.7812 in ..

SBX(R)

RAI RA2

+

0.000 -164.721

PDC FAX

-11.992 ~10A3 0.294 "'10.0. 6

RA3"'R'""R

+

RA4"'ln(R/B))

-5582.82 -7494.45 20513.79

psi psi psi

TEMA paragraph RCB 8.83 Cylindrical Element Stresses

10-12

Expansion Joinls


Pressure Vessel Design and Analysis Seminllr Component

COADE Engineering SoftwaFe PVElite 4.00 Licensee; CO~DE Inc., Local White Lock FileName; Seminar ---~------------------~---------------- Page 90 Thickjnt Analysis: EX-4 JOINT 1 Item: 1 9:51a Sep 21,2000

For the Inner Cylinder at X = YA DEA 3.285 /10"'3 BIA -195.282 /10 ..... 6 UlA 0.000 /10"'6 U2A 470.194 /10"'6

B2A SMYA

470.194 /10"'6 7.981 ·10"'3

For the Inner Cylinder at X = LA DEA 3.285 /10"'3 BIA -195.282 /10..... 6 ~O. 932 UIA U2A 243.144 /10"'6

B2A SMLA

470.194 /10"'6 7.499 *10"'3

For the Outer Cylinder at X 4.476 /10 ..... 3 BIB DEB UIB 0.000 /10 ..... 6 U2B

YB -2.063 /10"'3 -2.186 /10"'3

B2B SMYB

-2. lB6 /10"'3 3. 725 ·10"'3

For the Outer Cylinder at X 4.476 /10 .... 3 BIB DEB UIB -0.815 U2B

LB -2.063 /10 ....3 -3.389 /10"'3

B2B SMLB

~2.186

/10..... 3 1.767 *10"'3

TEM~ Paragraph RCB 8.84 Maximum Cyclic Stresses Evaluated per ASME A-99 Appendix CC, Paragraph CC-3(c);

For the Inner Cylinder: SCLA 5.604 *10"'3 RNA

1.000 *10"'6

For the Outer Cylinder: 22.350 *10"'3 RNB seLB

1.000 ·10"'6

Analysis of

Sh~~lside +

Differential Expansion

TEMA Paragraph RCB 8.70 Induced Axial Force: PIC 0.000 -38.750 PSPC= PSC = 200.000 -554.670 PSS =

PDC

515.920 -3562.022

FAX

TEMA paragraph RCB 8.80 Flexible Element Moments -0.4 3 8 THB -0.410 THA ZA 3.911 /10"'3 ZB 5.572 /10 ..... 3 MT1A -1. 025 /10"'3 MTZA 970.229 /10 ..... 6 RMA RMB MT1B 866.536 /10"'6 MT2B -805.393 /10"'6 TRMA Paragraph RCB 8.82 Flexible Element Moments -9.032 *10"'3 RA12 RAI RAll 14.369 *10 .....3 RA2l 3.524 *10 ..... ) RA22 2.750 ·10"'3 RA2 41.200 RA4 -40.460 *10"') RA3

4.936 *10 ..... 3 -5.508 *10"'3

·,.10"'3 1.767 ·10 .... 6

~23.401

Stress in Expansion Joint Flange: SBX(R) = (6.0/ TE**2) * (RA1 + RA2/{R·R) + RA3*R*R + RA4*ln(R/B) Stress Summary: Location: (A) 154375.70 psi 12.8437 in. I Location: (8) -173192.11 psi 16.7812 in. , TEMA paragraph RCB 8.83 Cylindrical Element Stresses For the Inner Cylinder at X DEl'. 4.563 /10"'3 B1A

Expansion Joints

YA -2.558 /10 ..... 3

B2A

-800.740 /10"'6

10-13




FileName: Seminar

-----~---------------------------------

Thickjnt Analysis; EX-4 JOINT 1

UtA

0.000

/loA 6

U2A

-800.740

Item:

1

Page 91

9:51a Sep 21,2000

/loA 6

SMYA

7.998 *10.0.)

For the Inner Cylinder at X DEA 4.563 /10"'3 BIA UIA -0.932 U2A

LA -2.558 /10""'3 -2.903 /l0"3

B2A SMLA

-800.740 /10"'6 3.530 *10")

For the Outer Cylinder at X DEB 5.572 /10"'3 BIB UIB 0.000 /10"'6 U2B

YB A 2.332 /10 ) 20.951 /10.0. 3

B2B SMYB

20.951 /10""'3 43.147 ·10"'3

For the Outer Cylinder at X DEB 5.572 /10.0.) BIB UIB -0.815 U2B

LB B2B SMLB

20.951 /10"'3 43.151 "'10""3

2.332 /10"3

20.953 /l0")

TEMA Paragraph RCB 8.84 Maximum Cyclic Stresses Evaluated per ASME A-99 lI.ppendix CC, Paragraph CC-3(c); For the Inner Cylinder: SCLA 0.163 *10"'6 RNA

3.278 *10 .... )

For the Outer Cylinder: SCLB 0.177 *10"'6 RNB

2.337 *10"'3

STRESS SUMMARY: Analysis of Differential Expansion Inside Junction Outside Junction Actual Allowed Actual Allowed Annular Element 155998. (Fatigue) -188909. (Fatigue) Annular Ring at R 16.781 188909. (Fatigue) CyL at point Y 16. (Fatigue) 38445. (Fatigue) CyL at point L -3870. (Fatigue) 40359. (Fatigue) Max. Cycle Stress 163570. (Fatigue) 194705. (Fatigue) Max. Cycle Life 3222. Cycles 1635. Cycles STRESS SUMMARY; Analysis of Shells ide Pressure Inside Junction Outside Junction Actual Allowed Actual Allowed Annular Element -1622. 60000. 15717. 60000. Annular Ring at R 14.008 -6070. 30000. CyL at point y 7982. 20000. 4701. 20000. CyL at point L 7400. 20000. 2792 . 20009· Max. Cycle Stress 1835. (Fatigue) 17406. (Fatigue) Max. Cycle Life 999999. Cycles 999999. Cycles STRESS SUMMARY: Analysis of Tubeside Pressure Inside Junction Outside Junction Actual Allowed Actual Allowed Annular Element -3961. (Fatigue) 4797. (Fatigue) Annular Ring at R 16.781 4797. (Fatigue) CyL at point y o. (Fatigue) -976. (Fatigue) Cyl. at point L 98. (Fatigue) -1025. (Fatigue) Max. Cycle Stress 4153. (Fatigue) 4944. (Fatigue) Max. Cycle Life 999999. Cycles 999999. Cycles

10-14

psi psi psi psi psi

psi psi psi psi psi

psi psi psi psi psi

Expansion Joints

Pressure Vessel Design aud Analysis - Seminar Noles

Pressure Vessel Design and Analysis SeminarCompollcnl

COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------------------- Page 92 Thickjnt Analysis: EX-4 JOINT 1 Item: 1 9:51a Sep 21,2000 STRESS SUMMARY: Analysis of Shellside ~ Tubeside Pressure Outside Junction Inside Junction Actual Allowed Actual Allowed 20514 _ (Fatigue) -5583. (Fatigue) Annular Element -7494. (Fatigue) 13.614 Annular Ring at R 3725. (Fatigue) Cyl. at point Y 7981. (Fatigue) 1767. (Fatigue) 7499. (Fatigue) Cyl. at point L 22350. (Fatigue) 5604. (Fatigue) Max. Cycle Stress 999999. Cycles 999999. Cycles Max. Cycle Life

psi psi psi psi psi

STRESS SUMMARY: Analysis of Shellside + Differential Expansion Outside Junction Inside Junction Actual Allowed Actual Allowed -173192. (Fatigue) psi 154376. (Fatigue) Annular Element 43147. (Fatigue) psi 7998. (Fatigue) Cyl. at point Y 43151. (Fatigue) psi 3530. (Fatigue) Cyl. at point L 171299. (Fatigue) psi 162873. (Fatigue) Max. Cycle Stress 2337. Cycles 3278. Cycles Max. Cycle Life The PV Elite Program,

Expansion JoinLs

(e) 1989-2000 by COADE Engineering Software

10-15

Metal Bellows Expansion Joints

Pressure Vessel Design and Analysis - Seminar NOles

METAL BELLOWS EXPANSION JOINTS These arc also called "thin-walled" expansion joints. They characteristically have multiple bellows with thin walls on the order of 0.05 in. There arc several ways to analyze metal bellows expansion joints. Many of them use clastic analysis of an equivalent structure, constrained by continuity effect. The most widely used method is the one presented in the standards of the Expansion Joint Manufacturers Association (EJMA). The method we will usc is the one presented in the ASME Code, Section VIII, Division I, Appendix 26. (This used to be Appendix BB, but it was moved up in the An revision to the Code.) These formulas take into account both internal pressure and axial joint movement. They do not include lateral or angular joint movements. The appendix covers expansionjoints up to liS-in. thick, with multiple convolutions, and includes both reinforced

and unreinforced expansion joints. Circumferential and meridional stresses are calculated in the bellows for both the reinforced and unreinforced geometries. In addition, for reinforccd expansion joinls the stress in the reinforcing clement and in any bolted fastener that may be on the reinforcing element arc calculated.

(

The ASME approach calculates the maximum stress for the worst combination of internal pressure andjoinl opening, and uses this value to calculate the maximum cycle life for the joint. Notice that the thickness you use in the analysis is the thickness of the sheet from which the bellows is formed, measured before the forming process occurs. The method itself calculales the thinning due to forming. The method in the Code also calculates stresses in a reinforcing ring and in a bolted faslener for the reinforcing ring. {Sec the sketch on the next page for details of this geometry.} The faclor q/2w is used to identify the location on the X axis of curves 26-3, 26-4, and 26-5. In the same way the factor q/(2.2* SQRT«d+w)-*t» is used to select the curve on these figures. Figure 26-3 is then interpolated linearly for Cpo Figures 26-4 and 26-5 arc interpolated logarithmically for Cf and Cd respectively. The meridional membrane plus bending stress due to pressure (Smmp+-Smbp) is compared to 1.5 times the basic allowable stress for unreinforced bellows and 3.0 times the basic allowable stress for reinforced bellows. The sum of the meridional stresses (Smmp+Smbp+Smmd+Smbd) is compared to the allowable stress (Sn) at the givennumber of design cycles for the given material type. The actual stresses in the reinforcement arc compared to the basic allowable stress for the reinforcement. The actual stresses in the bolt (fastener) arc compared to the allowable stress in the bolt. These values arc only calculated if AR and/or AF are greater than zero. The stress amplitude for cycle life evaluation is the sum of all the meridional stresses. This value is used with the appropriate formula in Paragraph 26-3 to calculate lhe cycle life.

10-16

Expansion Joints

(


Metal Bellows Expansion Joints

The cycle life calculation should be ba.sed on the fonnulas in the ASME Code, Section VIII, Division I, Appendix 26. These formulas implement a power law for fatigue life as follows:

N< 2.5 - [ _I4_.2_K--",-s'.::.. _ 0.02

E.

]

~.oo

for series 3XX high alloy steels, nickel-chromium iron alloy, nickel-iron chromium alloy, and nickel-copper alloy.

N~[15K ;.0 ]"1 _..!'~. -0.011

E.

for carbon, low alloy, series 4XX, and high alloy steels. N is the number of cycles, the cycle life. S. is the maximum stress range resulting from the analysis. Eb is the elastic modulus for the material at the temperature under consideration. K, is the fatigue strength reduction factor. Use a fatigue strength reduction factor of 1.5 for lInreinforeed bellows, and 3.0 for reinforced bellows. Alternative factors may be used if substantiated by test data, for example by hudrotes for design temperatures below the creep range. The nominal bellows thickness must be greater than the thickness required for pressure as calculated by the ASME method.

(

Expansion Joints

10-17

Melal Bellows Expansion Joints

Pressure Vessel Design and Analysis Seminar NOles R

The knuckle radius for the bellows must be grcater than threc limcs tm, the minimum unformed thickness. If r<3 'tm you should increase the K factor for fatigue and stress calculations from 1.0 to 4.0, usually resulting in unacceptably high slresses or low cycle life.

n Li-L

drt ~-------L---------.J (ill Unrelnfarced BellowJ

( ~A

~A'

/ - - .....

I

f L.A I

V

A-A

\

EqU3Iilir>g"ng

I E,.u t'<.!ualiling ring

~h GENERAL NOTE' Nominal t ;0 :3'm .

(b! Reinlor..,.Q' 8flllows

t

d'''m.

Figure 2--Geometry for Melal Bellows Expansio'n Joints

(

10-18

Expansion Joints


Pressure Vessel Design and Analysis Seminar Componenl

PRESSURE VESSEL DESIGN AND ANALYSIS SEMINAR COMPONENT DESIGN PROBLEM Metal Bellows Expansion Joint Analysis Evaluate the following expansion joint for a design life of 1800 cycles. Pressure is 150 psi (1.034 Nmm2) at 200°F (93°C). The joint opens .125 in. (3.175 mm) per convolution. Bellows ID is 24 in. (610 mm) The joint is fabricated from SA-240-304. The convolutions are I-in. (25.4 mm) deep on a 1.l25-in. (28.57 mm) pitch. The thickness of the bellows before forming is 0.05 in. (1.27 mm) This is an unreinforced bellows. There is no provision for corrosion.

Notes: The elastic modulus for SA-240-304 at ambient temperature is 28.3e6 (195121) The elastic modulus for SA-240-304 at 200°F is 27.7e6 (190984)

Questions: What is the cycle life for the bellows?

DO NOT TURN THE PAGE UNTIL YOU HAVE COMPI~ETED

Expansion Joinls

YOUR ANALYSIS

10-19

Pressure Vessel Design and Analysis Seminar ComponcnL Design Problem

Prcssure Vessel Design and Analysis-

COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------------------- Page 80 Thinjnt Analysis: EX-4, JOINT 2 Item: 1 9:51a Sep 21,2000

Input Echo, Thin Joint Item

Description: EX-4, JOINT 2

1,

Design Cycle Life, Number of Cycles Design Temperature of Expansion Joint Design Internal Pressure Expansion Joint Opening Per Convolution

1800 200.0 150.0 0.1250

NCYC TEMP P E

Material of Expansion Joint Bellows Bellows Operating Allowable Stress Bellows Elastic Modulus at Design Temp. Bellows Elastic Modulus at Ambient Temp.

MATI

ALLST1 EMOD1 EAMBI

SA-240 304 20000.0 . 27700E+08 0.2830E+08

F

psig in.

psi psi psi

Material Category for Cycle Life Evaluation AUSTENITIC STAINLESS Inside Diameter of Bellows Convolution Depth Convolution pitch Bellows Minimum Thickness before Forming Fatigue Strength Reduction Factor

24.0 1.0000 1.1250 0.0500 1. 0000

D W Q

TM

Kg

in. in. in. in.

(

THIN JOINT RESULTS, ASME VIII DIV. I App.26, 1998, A-9B Nominal Bellows Thickness Based on given Minimum T (TM ~ SQRT ( D/ ( D + W ) ) ) T (0.0500 ~ SQRT( 24.00/( 24.00 + 1.0000 ) ) ) T 0.0490 in. Cross Sectional Area of Bellows AS (0.571 * Q + 2. 0 ~ W ) * T AB (0.571 ~ 1.1250 + 2.0 * 1.0000 ) AB 0.1294 in Final Reinforcement Ratio

~

0.0490

~

1.0000

Minimum Required Thickness for Unreinforced Bellows Tu

P~(D+W)/(ALLST1*(1.14+4.0*W/Q»

Tu

150.0*( 24.0+ 1.0000)/( 0.0399 in.

Tu

20000.0~(1.14+4.0*

Required Bellows Thickness for Pressure

=

1.0000/ 1.1250»

0.0399 in.

Actual Knuckle Radius for one Convolution RKN (Q/4-TM/2) RKN (1.1250 / 4 - 0.0500 / 2 ) RKN 0.2562 in. Allowed Knuckle Radius for one Convolution TM3 = 3.0 * TM = 0.1500 in. Stiffening Factor for Bellows Under Pressure RKS 0.3-( 100.0/( 0.6 * P**1.5 + 320.0 ) )**2 RKS 0.3-( 100.0/( 0.6 ~ 150.0**1.5 + 320.0) )**2 RKS 0.2951

10-20

Expansion Joints


Pressure Vessel Design and Analysis Semill
COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName ; Seminar --------------------------------------- Page 81 Thinjnt Analysis: EX-4, JOINT 2 Item: 1 9;51a Sep 21,2000

SUMMARY for PRESSURE CALCULATIONS ; Nom. Bellows Thickness based on given Minimum T Required Bellows Thickness for Pressure TREQ Actual Knuckle Radius for One Convolution RKN Allowed Knuckle Radius for One Convolution 3"*TM Cross-Sectional Area for One Convolution Ab Fraction of Pressure Resisted by Bellows R Stiffening Factor for Bellows under Pressure RKS SUMMARY of FIGURES B8-3, 8B-4, BB-5 Horizontal Figure Factor Vertical Figure Factor q/(2.2«(d+w)* Value Interpolated from Figure B8-3, Cp Value Interpolated from Figure 88-4, Cf Value Interpolated from Figure B8-4, Cd

q/2W t)

"'* .5) Cp

Cf cd

0.0490 0.0399 0.2562 0.1500 0.1294 1.0000 0.2951

in. in. in. in.

in

0.56250 0.46207 0.63207 1.74880 1. 78330

STRESS RESULTS for UNREINFORCED BELLOWS Circumferential Stress in the Bellows SCMPU P*(D+W)/(T*(1.14+4.0*W/Q) SCMPU 150.0*( 24.0+ 1.0000)/( 0.0490*(1.14+4.0* 1.0000/ 1.1250)) SCMPU 16301.9 psi Meridional Membrane Stress due to Pressure SMMPU (P*W )/( 2.0*T ) SMMPU (150.0 * 1.0000 )/( 2.0 * 0.0490 ) SMMPU 1530.9 psi Meridional Bending Stress due to Pressure 5MBPU (p * W "* W * CP ) / ( 2.0 * T * T ) 5MBPU (150.0. 1.0000* 1.0000* 0.63207 )/( 2.0 "* 0.0490 * 0.0490 ) 5MBPU 19752.2 psi Meridional Membrane Stress due to Deflection SMMDU (EAMB1*T*T*E )/(2.0*W"3*Cf ) SMMDU (0.283E+08* 0.0490* 0.0490* 0.1250 )/(2.0* 1.0000.. . 3* 1.74880 ) SMMDU 2427.4 psi Meridional Bending Stress due to Deflection 5MBDU 5.0·EAMB1"*T*E/(J.0*W*W*CD) 5MBDU 5.0* 0.283E+08* 0.0490"* 0.1250/(3.0* 1.0000"* 1.0000* 1.78330) 5MBDU 161967.2 psi Summation of all Meridional Stresses STOTU = (SMBDU+SMMDU+SMBPU+SMMPU) = 185677 psi The Total Cycle Life CYCU (FM/ (FN*RKG*STOTU/EMOD1) -FO») ** 2.00 CYCU ( 2.5/« 14.2* 1.0000* 1856771 0.277E+08)- 0.020»)**2 CYCU n06. Maximum Stress for Given Cycle Life

Expansion Joints

10-21

Pressure Vessel Design and Anfllysis Seminar Camponelll Design Problem


COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------------------- Page 82 Thinjnt Analysis : EX-4, JOINT 2 Item: 1 9:S1a Sep 21,2000 SMAXC SMAXC

(FM/{RNCYC**(1/2.00)+FO)*(EMODl/FN*RKG)

SMAXC

153960.4 psi

( 2.S!( 1800**(1/2.00))+ 0.020)*( O.277E+08/ 14.2* 1.0000)

STRESSES in BELLOWS, psi Circumferential membrane, pressure Meridional Membrane, pressure Meridional Bending , pressure Meridional Membrane, deflection Meridional Bending , deflection Meridional Membrane + Bending , pressure Summation of all Meridional Stresses Stress Amplitude for Cycle Life Evaluation Cycle Life for Bellows The PV Elite Program,

Actual Scmp

Smmp 5mbp Smmd 5mbd

16301 1530 19752 2427 161967

21283 185677

Allowed 20000

20000

30000 153960 185677 1105

Ie) 1989-2000 by CQADE Engineering Software

(

10-22

Expansion Joints

\


Chapter 11: Stresses

STRESSES IN SHELLS DUE TO LOADS ON ATTACHMENTS This analysis is based on "Local Stresses in Spherical and Cylindrical Shells due to External Loadings," Welding Research Council Bulletin 107, August 1965, based on the prior work ofP.P. Bijlaard, and with revisions through 1979. In this section, you will calculate slrcsses at eight points around the periphery of an externally loaded attachment to a shell. You will analyze two shell types-cylinders and spheres. All head types are approximated as spheres. You will also evaluate several attachment types. On spheres, attachments can be round or square, hollow or solid. On cyUndcrs, attachments can be round, square, or rectangular, but all are considered solid.

IThe program does not calculate stresses in attachments, only in the shell of the vessel. , Stresses in attachments can, in fact, be higher than the stresses in the shell. liThe WRC107 analysis does not specifically address reinforced nozzles. An approach that I we believe to be eonservative is to analyze the nozzle at the nozzle OD and the vessel plus \ pad thickness, and then to analyze the nozzle at the reinforcing pad OD and the vessel · thickness. Ifboth of these analyses yield reasonable stresses, then the nozzle is OK. If this 1,approach does not work, a more sophisticated analysis is required. Loads for analysis of cylinders include the following: Design pressure on the cylinder Radial load on the cylinder (axial load on the nozzle) Cireumferential shear, dragging the attachment around the vessel Longitudinal shear, dragging the attachment along the vessel Torsional moment, twisting the attachment Circumferential moment, tipping the attachment around the vessel Longitudinal moment, tipping the attachment along the vessel The ratio of cylinder radius to wall thickness (the shell parameter, gamma) must be in the range 5 < RMITH < 300 or the analysis willllot work. In other words, the analysis is not valid for very thick or very thin vessels. The ratio of attachment radius to cylinder radius (the attachment parameter, beta) must be in the range 0.0571 < RO/RM < 0.571 or the program wiII give an error message. In other words, the program is not valid for very small or very large attachments. The value beta must be adjusted when calculating several of the figures for rectangular attachments. Thus, the error for a beta greater than 0.571 may occur even when the nominal beta is less than 0.571.

Slresses

It-I

Stresses in Shells due to Loads 011 Allachmcllls


The analysis e~lculales stresses at four points around the periphery of the allaehment, and at the inside of the cylinder and the outside of the cylinder at each point. The computer program makes the following modification: The last page of Appendix A (to WRC-107) indicates that curves IC and 2C were uneonservalive by as much as a factor of three, and were truncated for that reason. Now, however, instead ofjust using the values that exisl at the end of the truncated curves, we compared those points to the same point in Figure IC-I and 2C-I, multiplied by 3.0. For example, if the value from IC-I is more than three times less than the truncated value from IC, the value of IC-I multiplied by 3.0 will be used. The typical allowable stress for the combined stresses is 1.5 times the basic material allowable stress. If all your results are less than 1.5*SA then the configuration and loading are acceptable. If the load is self-relieving (that is, if it would disappear after only a small rotation or translation of the attachment), then the allowable stresses may be increase to 3*SA. Loads for the analysis of spheres include the following:

(

Design pressure on the cylinder. Radial load, along the axis of the attachment. Shear from point A to point B. Shear from point C to point D. Torsional moment, twisting the attachment. Moment tipping point A toward point B. Moment tipping point C toward point D. The shell parameter for spheres is Upsilon, the radius allachment divided by the decay length of the sphere. The ratio of sphere thickness to allaehment thickness must be in the range 0.25 < TslTn < 10.0 or the analysis will not work. The ratio of allachment radius to allachmentthiekness must be in the range 5 < RnlTn < 50.0 or the analysis will not work. If the above two parameters for a hollow allachment are out of range, the attachment can probably still be analyzed as a solid allachment. ' The analysis yields stress at both the inside of the sphere and the outside of the sphere at ) four points around the periphery of the nozzle. Stresses are not calculated in the nozzle wall. .

11-2

Stresses

(



DISCUSSION OF RESULTS WRC107 Stress Calculations CodeCale computes stress intensities in accordance with WRCI07 and includes the effects of longitudinal and hoop stresses due to internal pressure. If the geometry includes a circular reinforcing pad, CodeCale will perform two analyses on the geometry. The first analysis will compute the stresses at the edge of the nozzle. The second stress analysis will be at the edge of the reinforcing pad. CodeCale uses the Lame equation to determine the exact hoop stress at the upper and lower surface of the cylinder around the edge of the altaclunenl. The hoop stress equations, as well as the longitudinal stress equation, are as follows:

SHoop(Upper) = 2Sl.ong

Where

P

Design Pressure

r;

Inner diameter of shell

ro

Outer diameter of shell

For spherical shells the program uses the following equation:

SHoop = SLong

For each run performed, a table of dimensionless stress factors for each loading will be displayed for review. Any table figure followed by an exclamation point (!) means that the curve figure for that loading has been exceeded.

Why are the Stresses at Edge of the Pad the Same as at the Edge of the Nozzle? Since the stress is a direct product of the stress factor, the stresses computed at the edge of the pad may be same as those at the edge of the nozzle ifthe curve parameter for that type of stress has been exceeded.

What are the Allowable Stresses? The stress intensities computed should typically be between 1.5 and 3.0 times the hot allowable stress for the vessel material at operating temperature. If the results arc less than 1.5 Sa then the configuration and loading are acceptable. If the load is self-relieving (that is, if it would relax or disappear after only a small rotation or translation of the attachment), the allowable stress intensity would increase to 3.0 Sa.

(

Stresses

11-3



Since many geometries do not fall within the acceptable range of what WRClO7 will accept, it may be necessary to use a more sophisticated 1001 to solve the problems where the diameter of the vessel is very large in comparison with the nozzle or where the thickness of the vessel or nozzle is small. An example of a more sophisticated tool would be a FEA (finite elemeut analysis) program.

CYLINDRICAL SHELLS

SPHERICAL SHELLS

To Define WRC Axes: I. P-axis: Along the Nozzle centerline and positive entering the vessel.

2.

M I-axis: Perpendicular to the nozzle center-

3.

M2-axis: Cross the P-axis into the MI axis and the resull is the M2-axis.

To Define WRC Axes: I. P-axis: Along the Nozzle centerline and posi-

2.

live enleTing the vessel. Me-axis: Along the vessel centerline and pasi-

3.

live to correspond with any parallel global axis. M2-axis: Cross the P-axis with the Me axis and

line along convenient global axis.

To Define \VRC Stress Points: u-upper, means stress on outside ofvesscl wall at junction. I-lower, means stress 011 inside of vessel atjunction. A-Position on vessel at junction, along negative Ml axis. B-Position on vessel atjunclioll, along posilive M2 axis. C-Position on vessel at junction, along positive M2 axis. D-Position on vessel at junction, along negative M2 axis.

the result is the ML-axis.

To Define WRC Stl"CSS Points: u-upper, means stress on outside of vessel wall at junction. I-lower, means stress on inside of vessel a!junction. A-Position on vessel at junction, along negative MCaxis. B-Position on vessel at junction, along positive MCaxis. C-Position on vessel at junction, along positive MLaxis. D-Position on vessel aljunction, along negative ML axis. Note: Shear axis "VC" is parallel, and in the same direction as the bending axis "ML." Shear axis "VL" is parallel, and in the opposite direction as the'bending axis "MC."

11-4

Stresses


Discussion or Results

WRC107 Stress Summations ASME Seclion VIII, Division 2, provides for a fairly elaborate procedurc to analyzc the local stresses in vesscls and nozzles (Appendix 4-1 "MandatOlY Design Based On Stress Analysis"). Only the elastic analysis approach will bc discussed here. The user should always refer to the applicable code if any of the limits described in this seclion arc approached, or ifany unusual material, weld, or stress situation exists, or thcre are non-linear concerns such as the material's operation in the creep range. The first step in the procedure is to determine if the clastic approach is satisfactory. Seclion AD-160 contains the exact method and basically states that if all of the following conditions are met, then fatigue analysis need not be done: a) The expected design number offull-range pressure cycles does not exceed the number of allowed cycles corrcsponding to an S, value ons m(4S mfor non-integral allachments) on the material fatigue curve. The Sm is the allowable stress intensity for the material at the operating temperature. b) The expected design range of pressure cycles other than startup or shutdown must be less than 1/3 (1/4 for non-integral attachments) the design pressure times (S,I Sm), where S, is the value obtained on the material fatigue curve for the specificd number of significant pressure fluctuations. c) The vessel does not experience localized high stress due to heating. d) The full range of stress intensities due to mechanical loads (including piping reactions) does not exceed S, from the fatigue curve for the expecled number of load fluctuations. Once the user has decided that an elastic analysis will be satisfactory, either a simplified or a comprehensive approach may be taken to the vessel stress evaluation. Both melhods will be described in detail below, after a discussion of the Section VIII Div. 2 Requirements.

Stresses

11-5

Discussioll of Results


ASME Section VIII Division 2-Elastic Analysis of Nozzle Ideally in order to address the local allowable stress problem, the user should have the endurance curve for the material of construction and complete design pressure / temperature loading information. If any of the elastic limits are approached, or ifthere is anything out of the ordinary about the nozzlelvessel connection design, the code should be carefully consulted before performing the local stress analysis. The material Sm table and the endurance curve for carbon steels are given in this section for illustration. Only values taken directly from the code should be used in design. There are essentially three criteria that must be satisfied before the stresses in the vessel wall due to nozzle loads can be considered within the allowables. These three criteria can be summarized as

Where Pm. P r, Pb' and Q are the general primary membrane stress, the local primary membrane stress, the local primary bending stress, and the total secondary stresses (membrane plus bending)} respectively; and K, Smh' and Smavg are the occasional stress factor, the hot material allowable stress intensity, and the average material stress intensity (Smh + Sm,) / 2. Because of the stress classification defined by Section VIII, Division 2, in the vicinity of nozzles, as given in the Table 4-120. I, the bending slress terms caused by any external load moments or internal pressure in the vessel wall ncar a nozzle or other opening, should be classified as Q, or the secondary stresses, regardless of whether they were caused by sustained or expansion loads. This causes P b to disappear, and leads to a much more detailed classification: Pm

General primary membrane stress (primarily due to internal pressure)

PI

Local primary membrane stress, which may include the following: Membrane stress due to internal pressure Local membrane stress due to applied sustained forces and moments

Q

Secondary stresses, which may include the following: Bending stress due to internal pressure Bending stress due to applied sustained forces and moments Membrane stress due to applied expansion forces Bending stress due to applied expansion forces and moments Membrane stress due to applied expansion moments

Each of the stress terms defined in the above classifications contain three parts: two stress components in nonnal directions and one shear stress component. To combine these stresses, the following rules apply: Compute the normal and shear components for each of the three stress types, i.e. Pm, PI, and Q. Compute the stress intensity due to the Pm and compare it against kSmh.

tl-6

Stresses

(

Discussion ofResulls

Pressure Vessel Design c1nd Analysis - Seminar Noles

Add Ihe individual normal and shear stress components due 10 Pmand PI, compute Ihe resullant slress intensily, and compare ils value against 1.5kSmh. Add Ihc individual normal and shear slress componenls due to Pm, PI, and Q, compute Ihe resuhant slrcss inlensity, and compare ils value 10 against 3Smavg. If Ihere is an occasional load as well as a suslaincd load, Ihese Iypes may be repeated using a k value of 1.2. These criteria can be readily found from Figure 4-130.1 of Appendix 4 of ASME Seclion Vl1l, Division 2 and the surrounding lexl. Nole Ihallhe primary bending slress lerm (Pb) is not applicable 10 the shell stress evalualion, and Iherefore disappears from Ihe Scclion Vl1l, Division 2 requirements. Under Ihc same analogy, Ihe peak slress limil may also be

written as PI + P b + Q + F < S,

The above equation need not be salisfied, provided Ihe elaslic limit crileria of AD-160 is met based on the slalement explicitly given in Seclion 5-100:

(

"If the specified operation ofthe vessel meets all ofthe conditions ofAD-160, no analysis for cyclic operatioll is required and it may be assumed that the peak stress limit discussed in 4-135 has been satisfied by compliance with the applicable requirements for materials, design,jabricafion, testing and inspection afthis division. "

,"

I

J.

~'"

II) E - 30" to'poi.

1211tl1.'~.lo
131 Tlbhl5-110.1

~ ......

.. buI .."" ...... _

•

Iooonul.ofor.n

I l«u,,'.",I.'~,IionQl"'"''

........ 0-

., /_'OfUTs" ~ ~~

eo ...

,,,.

"'--

!"<0-

I-

..•..

' ,

'"

,

,

,

,,,

------- --r...

~

,

UTS Il!-UOb'

'" ..' , ;- "':;;"

FIG. s-nO.1 DE-SICH fAll(,UE CURVES fOR CARBON. LOW AllOY, SERIES 4XlC, HleH ALLOY STEELS AHD HleH TENSILE STEELS FOR TEM~ERATURES NOT EXCEEDINC 100'F

Figure 1--Example Fatigue Curve (For Values of Sa) The equations used in CodeCale to qualify the various stress components can be summarized as follows: Pm(SUS) < Smh Pm(SUS + acC) < I.2S mh

Stresses

11-7

Discussion of Resulls


Pm(SUS + OCC) + Pj(SUS + OCC) < 1.5( 1.2)Smh Pm(SUS + OCC) +NSUS + OCC) + Q(SUS + EXP + OCC) < 1.5(Sm, + Smh) If some of the conditions listed in AS ME VIII Div.2 (in which AD-I 60 is not satisficd), you probably need to perfonn the fonnal fatigue analysis. Peak stresses are required to be calculated or estimated. You may consider using AD-560 "Alternative Rules for Nozzle Design" instead of Article 4-6 "Stresses in Openings for Fatigue Evaluation" to calculate the peak pressure stress for the opening. Ifall conditions of AD-560.1 through AD-560.6 are satisfied, the stress indices given in Table AD-560.7 may be used. Ifuser says "Yes" to indicate the conditions have been satisfied, the program will use these pressure stress indices to modify the primary stress due to internal pressure (hoop and longitudinal stresses). For external loads, the highest peak stresses are usually localized in fiIlets and transitions. If the user decides to use WRCI07 stress concentration factors (Kn, Kb), the fillet radius between the Vessel and Nozzle is required. (If a reinforcing pad is used, the program assumes the same pad fillet radius.) The program will make a crude approximation and use WRCI07 Appendix-B equations (3) and (4) to estimate Kn and Kb. The tension and bending stresses are thus modified using Kn and Kb respectively. The program outputs the local stresses for 4 pairs of points (upper and lower) at the intersection. Note

The user should 1101 direct the program to perform the stress summations. Instead the user should detennine which stresses should be added based on locations in order to obtain the peak stress level, then use Appendix-4 & 5 rules and fatigue curves depending on operation cycles.

Based on comparisons with finite element analysis, it is known that the top tip of the fiIlet weld on the nozzle usually experiences the highest peak stress due to external loads. So it is conservative to add all the peak stresses after including both pressure stress indices and concentration factors. Note

The stress summation may ollly be used to check stress intensities, not stress !evels.

You need the peak stress level to perform fatigue analysis. The current stress summation routine does not compare stress level with fatigue allowables per Appendix-5. However,

J 1-8

Stresses



you may find the stress summation results useful to compare the combined effect due to the stress concentration factor and pressure stress indices. ""------"'Ll:%'lOITUOIN.'
o GEOMETRY

POINTS FOR STRESS RECOVERY

(

Figure 2--Geometry for Local Stresses in Cylindrical Shells

Stresses

11-9



NOZZLE I. ""pl•., Lud,' Ilooli.1 Lood. SI._L•• "" SI..., L ... d. 0.- ......",".......,,'. O.... ~, ... ftt .140 ....... I ......

r.,,'....

l.

c

u,.

,

r __

Vl~

_ _ lb, "'--~I!.:,. vl~

',- _ _ If_,t

"

..,

..

T T

P

U

P.,_OI."

'., " - '"

\'''1r.T

.

0.' -,-

_ _ : .... _ _ : •. _ _ :,.- •

,--

.t.T

f

_ _ ;f.. __

1-0. '.'

i".

.. _~ ..... I... d.IC .. "' _ _ b",lin.I..d. llo ...- _ _ "'OTE, I!ftlo •• n IOU Iuoo I.. u .... donee -
c-,.... ~ .. h.....1.....1

R.d.~ ....

f;.

" 10

,.

.. s"... C ... c ......II011 Foe,...

I Thlch.... T '" III __ A••;~ •• Il..~ ~ I. TMd".... • .. .......1. lol.uR.d;y., '",= .....,.I.OuIJ;d. Rd;u •• I.=-

,..

C........ic

-_.'I.,';.

", V V

lP~1

,. 'b.

_ _ I>.

Do DL

-,-

..

,.

,. ,

..

~

.. ,-"')- "

o.

Rm

STllESSES _ lilood', .,,..';1. ,h." .t.o

.1 ....................1'·

. (";').

HOLLOW ATTACHMENT

+-

iT

+

+

+

1.. _1 I. 10

-.-,-

••

(N.r, 1i;l) -.-,-

•1'\1'it:i

II., Il..I

:b ( ... ,.~)

.~

N.I, i;i" -',-

01.

." .

-,.Y,

Hr'.,tI;t ~

Y,

IO."L' -Y-,-' H",r.=f" ~

Y,

II.~-"&i

-.-,-'

_

q ..

('~')

. . .

(".t~~)

"'

.. I

....,

" ... ,

; f

W,

W•

11.

Y,

."..[W

-II,- -

"V ••'

,.."

"ya;f

(Y'~)

... ~d al ••I... I..II, hll ....... lIn 01 aJ. •• 1

~

'Y'

T S.o.l '. It

I ....· ...'

.. ',.• "(.!'f) " . .. • (II,\IW) .... . '- , .. (",,\Ji;"i) .__,,_ ..

"'dd al,"".lcoll, lor • _ _ ~ ... f IP_I •• 10

o.

¥

~,

',.yt;T

~

COKBINEO STP.E:SS INTmSJrY - S.

11. When '( "I- o. s .. laraC!.':t llbsoll.lte 1l\.1'lnltuc1C! of eithC!r S .. 1/2 [Ox+Oy! {lOll: Oy)2 .. 4Tl)or {(all. - Oy)2~ 2) When t ~ O. S .. lIIrqest Abi!;Ol\Ite Illaqnitude of either S .. 0.:c' 0y or (Ox - a I •

7

Figure 3--AII Attachments to a Spherical Shell

ll-JO

Stresses



p 1. ",,,11'" 1. ..1.' ,. =---1'"

••iIl.II.. ,.

101.", _ _ 1... 1", l1li1.:: _ _ 1•• 1\,

CI ....... ,~'.

1..",.101_.". T...I..........,. Sh... L.. ~. .. _I. .....

I f . : - - I... I",

v.

............••

'"

",,,,,"-. .1,,,11 •••

Y....l .. ~, ...

.••

or

)C

.. ..

.."....4_ •• , ..

•

•

-

11./ ....•

..":!-

10

•

~'b

..

" •• 18 •• Jlv··fJ =

(

-it ;:--

·1.' ....

, •• I •• ..' Ko K\

..1.·

nIlE~U~

--

Om

--

lwII I. ,. .

CYLINDRICAL SHELL

_ "1. . 4 ..., ....11. ,h•• t.o .... ,........., .........

"'.~L

...

BLc..CLDw

_

P/...

• .., -

•

..

,.., . . . ~

C

'HOT(, (", n 10 • 1 1,10. .1

:.--"

•• (..!!2. . ...!...

--'!t = "b

Ie or 1C·1

tUJSl

.1 .. _\_.1 \J \ ...1'", '''1.

,.

Y.... I,"'.~~....

.. ".

St

-_ ... ... -_ ..

2. G_....,

,

{J

=__ "

VI. : _ _ " •

+

..-

..-

• .... {jt ;;:;.

(~).~ .., ....1' ••• {jt

I I •• -11.1

.. " .. . .... • . •

.....1

,1..... -

•1+ "

+_

or

3("

..,..;

.....

.. (~). -!.- <

~

P/....

.

.. (.!!..).

1C-1 01'

T

2e·

II.'

~

"

.0=:

··/~·fJ ~

... ....

.......,ll

IIL/••

,:II_I' A~

·U

"1./"'fJ

.II_'..

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-'l.'....... .,.. No... ,

u/_1I.a J'~;Il

=

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10. _ _

• ...flr·

~

....

" r.." so. •• 1

IIl.

. V.

J.IlA"

.,....,;.., ,

n

__ II_

1'".

COMBINED STRESS INTENSITY - S 1) When T # 0, S ;; largest absolute ma~nitude of either S;; 1/2 [Ox+04l ~ .,t(Ox - 0tP)2 + 4T 2 Jor I(a~ - 0'41)2 +

(

2) When S'-D

~

ox'

;; 0, S = largest absolute magnitude of either o~

or (ax - °41)

Figure 4--AII Attachments to a Cylindrical Shell

(

Slresses

1t -11


Pressure Vessel Design (lnd Analysis - Seminar Notes

(

11-12

Stresses

(




Stresses in a Cylinder due to Loads on a Nozzle Analyze the cylinder for Drum D-IOI under the following loads: Design pressure is 200 psi (1.378N/mm 2) Radial load is 5,400 lbs (24021 N) Longitudinal shear load is 2, I00 Ibs (9342 N) Longitudinal bending moment is 3500 ft-Ibs (4.74e6Nmm) The cylinder is 144 in. (3658 mm) ID and I-in. (25.4 mrn) thick The attachment is ronnd and has an OD of 12.75 in. (324 mm)

( Notes:

Questions:

(

What is the highest stressed point? Is the stress at Ihis point acceptable?


(

Stresses

11-13

Pressure Vessel Design and Analysis ~ Seminar NOles


COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------------------- Page 71 WRCI07 Analysis: Nozzle/Cylinder Item: 1 9;51a Sep 21,2000

Input Echo, WRC107 Item

Description: Nozzle/Cylinder

"

VBASIS CYLSPH

Diameter Basis for Vessel Cylindrical or Spherical Vessel Corrosion Allowance for Vessel Vessel Diameter Vessel Thickness

ID

Cylindrical 0.0000 in. 14-<1.000 in. 1.000 in.

CAS

DV TV

Attachment Type

TYPE

Round

N8ASIS

Diameter Basis for Nozzle Corrosion Allowance for Nozzle Nozzle Diameter Nozzle Thickness

ID

D"

in.

TN

0.0000 12.000 1.000

DP

200.00

psig

CAN

Design Internal Pressure

External Forces and Moments in WRC 107 Convention: Radial Load P Circumferential Shear VC Longitudinal Shear VL Circumferential Moment MC Longitudinal Moment ML Torsional Moment MT Compute Compute Compute Compare

Maximum Maximum Maximum Maximum

5400.00 0.00 2100.00 0.00 3500.00 0.00

Radial Force Circumferential Moment Longitudinal Moment Stress Intensity to

"0 "0 26250.00

0.00 lb. 44.00 lb. 44 .00 ft.lb. 0.00 ft .lb. 0.00 fLlb. 0.00 psig

P

INTACT VERSION

"0

"0

March 1979 ( 81 & B2 )

"0

Include WRC107 SIF(Kn,Kb)--concentration factors Include Pressure Stress Indices per Div. 2

( P/Rm ( P )

0.084 0.084

N{PHI) / M(PHI) /

11-14

4C

No

72.50

Figure

2C1

lb.

-LOI

My Mz

Dimensionless Parameters used: Gamma = Dimensionless Loads for Cylindrical Shells

psi

Fx Fz

Beta

ft.lb. .Ib. ft.lb.

n

Fy Mx

Curves read for

lb. lb. lb.

Yes

Global Force (SUS) Global Force (SUS) Global Force (SUS) Global Moment (SUS) Global Moment (SUS) Global Moment (SUS) Internal Pressure (SUS) Include Pressure Thrust Use Interactive Control WRCI07 Version

in. in.

value 12.282 0.078

Slresses

(



COME Engineering Software PlJElite 4.00 Licensee: COADE Inc. , Local White Lock FileName , Seminar WRC107 Analysis Nozzle/cylinder Item: 1

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

,

N (PHI) H(PHI) N(PHI) M(PHI)

I I I I

N(x)

I I I I I I

M(x)

N(x) M(x) N(x) M(x)

MC/(Rm**2 MC/ (Rm ML/(Rm**2 ML! (Rm

I I I

N(x)

M(x) M(x)

Beta) Beta) Beta) Beta)

0.084 0.084 0.084 0.084

3A 1A 3B 1B

Beta) Beta) Beta) Beta)

0.084 0.084 0.084 0.084 0.084 0.084

3C 1C1 4A 2A 4B 2B

10.437 0.118 3.410 0.051 2.289 0.063

0.084 0.084 0.084

3C 1C 1B1

10.437 0.115 0.042

0.084 0.084 0.084

4C 2C 2B1

12.282 0.078 0.064

P/Rm ) p

)

MC!(Rm**2 MC! (Rm ML!(Rm**2 ML! (Rm

STRESS POINTS C & 0 N(PHI) I ( P!Rm ) M{PHI) I ( p ) M(PHI) I ( ML/ (Rm

(

•

• • • •

·

BETA)

)

)

ML/ (RID

2.280 0.093 8.115 0.043

(MARCH 1979)

P!Rm ) p

Page 72 9:51a Sep 21,2000

•

BETA)

Stress Concentration Factors Kn = 1.00,

Kb

1.00

Stresses in the Vessel at the Attachment Junction [ [

Type of

Stress Values at (psi)

---------------1-------------------------------------------------------Stress

Load I

Au

Al

Bu

Bl

Cu

Cl

Du

D1

---------------1-------------------------------------------------------Circ. Cire. Cire. Cire. Cire. Cire. Cire.

Memb. Memb. Bend. Memb. Bend. Memb. Bend.

DP P P MC MC ML ML

I I I I

I I I

14300 -914 -2523 0 0 -767 -1739

14500 -914 2523 0 0 -767 1739

14300 -914 -2523 0 0 767 1739

14500 -914 2523 0 0 767 -1739

14300 -777 -3711 0 0 0 0

14500 -777 3711 0 0 0 0

14300 -777 -3711 0 0 0 0

14500 -777 3711 0 0 0 0

8357

17081

13369

15137

9812

17434

9812

17434

7150 -777 -3817

7150 -777 3817

7150 -777 -3817

7150 -777 3817

7150 -914 -2541

7150 -914 2541

7150 -914 -2541

7150 -914 2541

0 0

0 0

-216 -2648

-216 2648

0 0 216

0 0 216

2648

-2648

0 0 0 0

0 0 0 0

0 0 0 0

0 0 0 0

-308

12622

5420

7758

3695

8777

3695

8777

o o o

o o o

o o o

o o o

o

o

o

o

-95

-95

95

95

o

o

o

o

I Tot. Circ. Str.[ Long. Long. Long. Long. Long. Long. Long.

(

Mernb. Mernb. Bend. Memb. Bend. Memb. Bend.

OP P P

MC MC ML ML

1 1 1

1 1 1 1 1 Tot. Long. Str·1 Shear VC Shear VL Shear MT

Stresses

11-15



COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Semiriar ---------------------------------------

WRCI07 Analysis: Nozzle/Cylinder

I

Item:

1

Page 73

9:51a Sep 21,2000

Tot. Shear I

o

o

o

o

-95

-95

95

95

I

8665

17081

13369

15137

9813

17435

9813

17435

Str.

Int.

With the Given Forces and Moments the

11-16

Maximum Radial Force

(Edge of Nozzle) is .....



21621.90

lb.

Siresses




Stresses in a sphere due to Loads on a Nozzle Analyze the elliptical head for Drum D-IOI under the following loads: .Design pressure is 200 psi. (1.378 N/mm2) Radial load is 5,400 Ibs. (24021 N) Longitudinal shear load is 2,100 lbs. (9342 N) Longitudinal bending moment is 35,000 ft-lbs. (4745 N-m) The elliptical head is 144 in. (3658 mm) ID and I-in. (25.4 mm) thick. The attachment is round and has an OD of 12.75 in. (324 mm) (

Notes:

Questions:

(

What is the highest stressed point? Is the stress at this point acceptable?


Stresses

11-17

Discussion

or Results


COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --~------------------------------------ Page 74 WRCI07 Analysis: Nozzle/Head Item: 2 9:51a Sep 21,2000

Input Echo, WRCI07 Item

2,

Description: Nozzle/Head

Diameter Basis for Vessel Cylindrical or Spherical Vessel Corrosion Allowance for Vessel Vessel Diameter Vessel Thickness

VBASIS CYLSPH CAS

10

Spherical 0.0000 in. 259.200 in. 1.000 in.

OV TV

Attachment Type

TYPE

Diameter Basis for Nozzle Corrosion Allowance for Nozzle Nozzle Diameter Nozzle Thickness

Round

NBASIS

10

in.

TN

0.0000 12.750 0.6B7

OP

200.00

psig

CAN

ON

Design Internal Pressure

External Forces and Moments in WRC 107 Convention: Radial Load P Circumferential Shear (VC) V2 Longitudinal Shear (VL) V1 Circumferential Moment (MC) Ml Longitudinal Moment (ML) M2 Torsional Moment MT Compute Compute Compute Compare

Maximum Maximum Maximum Maximum

5400.00 0.00 2100.00 0.00 3500.00 0.00

Radial Force Circumferential Moment Longitudinal Moment Stress Intensity to

Use Interactive Control WRCI07 Version

lb. lb. lb. it.lb. it.lb. ft.lb.

Yes

No No 26250.00 INTACT VERSION

psi

No March 1979 ( Bl & B2 )

Include WRCI07 SIF(Kn,Kb)--concentration factors Include Pressure Stress Indices per Div. 2

No No

Dimensionless Parameters: U ;

1.46

0.62 TAU

in. in.

9.78 RHO

Dimensionless Loads for Spherical Shells Curves read for N(x)

•

M(x)

T /

p

/ P

N{x) • T • SQRT(Rm M(x) • SQRT(Rm N(x) • T • SQRT(Rm M(x) • SQRT(Rm N(y) M(y)

N(y)

II-IS

• T /

•

Figure

*

• •

T T T T

/ ML

SP SP SM SM SM SM

2 2 2 2 2 2

0.05315 0.0598B 0.10030 0.13465 0.10030 0.13465

/ MC

SP 2 SP 2 SM 2

0.15573 0.03152 0.14582

/ MC / MC / ML

P

/ P

T • SQRT(Rm • T )

Value

Stresses


Discussion of Resulls


4.00

Licensee: COADE Inc., Local White Lock ---~--------------------------------

FileName: Seminar

WRCI07 Analysis : Nozzle/Head M(y) N(y) M(y)

Item:

I MC I ML I ML

,.. SQRT(Rm T ... T ... SQRT(Rm T SQRT(Rm ... T

Stress Concentration Factors Kn

=

2

SM 2 SM 2

0.09009

SH 2

0.09009

Page 75

9: 51a Sep 21,2000

0.14582

1.00,

Kb == 1. 00

Stresses in the Vessel at the Attachment Junction

I I

Type of

Stress Values at (psi)

---------------1-------------------------------------------------------Stress

Loadl

Al

Au

Bu

81

Cl

Cu

Dl

Du

---------------1-------------------------------------------------------Rad. Rad. Rad. Rad. Rad. Rad.

DP P P MC MC

I

-2974

12960 -287 1940 0 0 -369 2974

12960 -287 -1940 0 0 369 2974

12960 -287 1940 0 0 369 -2974

12960 -287 -1940 0 0 0 0

12960 -287 1940 0 0 0 0

12960 -287 -1940 0 0 0 0

12960 -287 1940 0 0 0 0

Str.1

7390

17218

14076

12008

10733

14613

10733

14613

DP 1 P 1 P 1

12960 -840 -1021 0 0 -536 -1990

12960 -840 1021 0 0 -536 1990

12960 -840 -1021 0 0 536 1990

12960 -840 1021 0 0 536 -1990

12960 - 840 -1021 0 0 0 0

12960 -840 1021 0 0 0 0

12960 -840 -1021 0 0 0 0

12960 -840 1021 0 0 0 0

Tot. Tang. Str·1

8573

14595

13625

11687

11099

13141

11099

13141

I

0 0 0

0 0 0

0 0 0

0 0 0

0

0

0

-9'

-9'

9'

9.

0

0

0

0

0

0

0

0

-9'

-9'

9.

9'

8573

1721B

14076

1200B

11121

14618

!11121

14618

Rad.

Memb. Mernb. Bend. Memb. Bend.

I I I I I

Memb. ML Bend. ML 1

12%0 -287 -1940 0 0 ~369

1

Tot. Rad. Tang. Tang. Tang. Tang. Tang. Tang. Tang.

Memb. Memb. Bend. Memb. Bend. Memb.

MC MC ML Bend. ML

shear VC Shear VL Shear MT Tot.

I I I

1

I 1 Shear I

Str. Jnt.

(

1

I

I

With the Given Forces and Moments the Maximum Radial Force (Edge of Nozzle J is The PV Elite Program,

(

Slresses

0.0"

34893.37

0

lb.

(cl 19B9-2000 by COADE Engineering Software

11-19



STRESSES IN HORIZONTAL PRESSURE VESSELS The Zick analysis method calculates stresses in horizontal pressure vessels created by the combination of internal pressure and the weight of the vessel and its contained liquid. The method is based on "Stresses in Large Horizontal Cylindrical Pressure Vessels on Two Saddle Supports," The Welding Journal Research Supplement, J95 I, and subsequent interpretations of that work. The geometry analyzed by the Zick method is a horizontal cylinder with torispherical, hemispherical, or spherical heads, supported at two locations on saddles. The saddles extend between 100 degrees and 180 degrees around the shell and mayor may not have wear pads separating them from the shell. The melhod calculates the stresses at several points in a horizontal vessel, including the following: Longitudinal stresses at the top and bottom of the shell near the saddles, including pressure stresses Longitudinal stresses at the top and bottom of the shell at the midspan, including pressure stresses Tangential shear stresses in the shell Tangential shear stresses in the head Circumferential stress above the saddle, including pressure stresses Stress in the head if it is used as a stiffener Compressive stress in the shell The method also analyzes additional stiffening rings that may be required to reduce the stresses to acceptable levels. The saddle is generally located close 10 the head in order to take advantage of the stiffening effect of the head. In general, 'A' should be less than 0.2*L. Acceptable bearing angles are between 100 and 180 degrees. If you enter a value greater than or equal to 180 you will not be able to compute the result. The most typical value is 120 degrees, except for very small vessels, and the maximum is generally f50 degrees. If the dimensions of the wear pad and the saddle are appropriate, this method allows summing the thicknesses of the shell and pad for circumferential membrane stress calculations. For circumferential bending stress calculations the method will use the square root of the sum of the squares of the wear pad and shell thicknesses. If the wear pad extends above the horn orthe saddle more than R110 in., and if the saddle is within O.5R of the head, then the wear pad thickness may be included in the caleulation for the circumferential stress at the horn of the saddle. If the wear pad width is more than 1.56*SQRT(R*TS) then the wear pad thickness may be included in the calculation for the circumferential stress at the bottom of the vessel. This type of vessel may have a full encirclement stiffener either on the ID or the OD. The purpose of the stiffener is to keep the shell round, just as in external pressure caleulations. Several different kinds of stiffeners may be analyzed. These include simple plate stiffeners, tees, double plate stiffeners, and beam sections. Longitudinal stresses at the saddle include both the bending stresses across the vessel, and the longitudinal pressure stress (SLP+-SLS). The allowable stress in tension is the shell allowable stress times the shell joint efficiency. The allowable stress in compression is the allowable compression per the ASME Code, Paragraph UG-23 , using the materials chart

11-20

Stresses

(

Stresses ill Horizontal Pressure Vessels

Pressure Vessel Design and Analysis. Seminar Noles

associated with the given material. If you entered alternate stresses for the material, then the allowable compressive stress is based on chart CS-2. The top section of the shell for an unstiffened vessel is ineffective as a beam. Therefore, except for the stiffened case, the section modulus of the shell is lower at the saddles than away from them. Longitudinal stresses at the mid-span include both the bending stresses across the vessel, and the longitudinal pressure stress (SLP+-SLM). The allowable stress in tension is the shell allowable stress times the shell joint efficiency. The allowable stress in compression is the allowable compression per the ASME Code, Paragraph UG-23, using the materials chart associated with the given material. If you entered alternate stresses for the material, then the allowable compressive stress is based on chart CS-2. The tangential shear in the shell varies depending on whether the shell is stiffened or the head acts as a stiffener, or neither of these cases. Tangential stress in the head only exists if the head is close enough to the saddle to be used as a stiffener. The allowable stress in shear is 80% of the allowable tensile stress. The stress at the hom of the saddle depends on the location of the saddle and the equivalent thickness of the saddle and wear pad. It is zero if the shell is stiffened by rings. This stress is always compressive and the allowable stress is -1.5 times the allowable tensile stress.

Use of a head as a stiffener creates additional tensile stress in the head. The allowable additional stress in the vessel head is limited to 0.25 times the allowable tensile stress. If you add the pressure stress in the head (not shown) to this value, the resulting stress must be less than 1.25 times the allowable tensile stress. The ring compressive stress in the shell is the compression at the bouom of the saddle due to the weight of the vessel. The allowable membrane compressive stress is taken to be one half the yield stress or 1.5 times the tension allowable, whichever is less. If the stiffening ring is internal then the maximum stress in the shell is compressive. If the ring is external, the stress in the shell is tensile, and internal pressure adds to this stress. The slress should be limited to +/- 1.5 times the shell allowable stress.

(

POINTS FOR STRESS RECOVERY

..." LONG! TUOINAl. smESS l3 SHEAR STRI::SS

!

C1RCVUFERENT'~L STRESS

Figure 5--Geometry for the Analysis of Horizontal Vessels on Saddle Supports

(

Stresses

11-21


Pressure Vessel Design and Analysis Seminar Notes R

(

11-22

Stresses

Stresses in Horizonlal Pressure Vessels



Horizontal Vessels on Saddle Supports (Zick Analysis). Analyze Drum D-I 0 I for stresses as a horizontal drum: Design pressure 230 psi (1.585 Nmm 2 ) at 450°F (232°C) SA-516,70 (shell and heads) with 0.125-in. (3.175 mm) corrosion allowance 144-in. (3658 mm) ID with I-in. (25.4 mm) wall, full radiography 2:1 elliptical heads, I-in. (25.4 mm) thick, full radiography Vessel is 20 ft 6096 tangent to tangent. Saddle is 30 in. (762 mm) from the vessel tangent, and 12 in. (304.8 mm) wide. The saddle has an arc of 120 degrees.

(

The wear pad, 0.375, 9.5 extends only a small amount outside the saddle. There is no reinforcing ring. Weight of attachments and appurtenances

~

40,000.0 lbs. 177934N

Notes:

Questions: What happens if you extend the wear pad?


Stresses

11-23

Stresses in lIorizontal Pressure Vessels


COADE Engineering Software PVElite 4.00 Licensee; COADE Inc .• Local White Lock FileName: Seminar --------------------------------------- Page 60 Horizves Analysis; 0101 Zick Item: 1 9:51a Sep 21,2000

Input Echo, Horizontal Vessel Item

I,

Design Internal Pressure Design Temperature Corrosion Allowance for Vessel

Description: 0101 zick 200.00 450.00 0.1250

Shell Material Shell Operating Allowable Stress Shell Ambient Allowable Stress Head Material Head Operating Allowable Stress Head Ambient Allowable Stress

SA-516 70 20000.00 20000.00 SA-516 70 20000.00 20000.00

Density of Shell and Head Material Liquid Height in vessel Density of Stored Liquid Extra Weight

0.2830 144.2500 0.2830 40000.000

Baseplate Length Baseplate Thickness Baseplate Width Number of Ribs ( inc. outside ribs ) Rib Thickness Web Thickness web Location Height of Center Web

168.0000 1.0000 12.0000

Design Temperature of Base Structure Saddle\Baseplate\Rib\Web Material Operating Allowable Stress Ambient Allowable Stress

100.00 SA-516 70 20000.00 20000.00

Use ASCE 7-95 Force Coefficient Extra Area Importance Factor Wind Velocity Exposure Category Height above Grade Distance from Center of Vessel to SupportSeismic Loads Present

psig F

in.

psi psi psi psi lb. lin in. lb. 1ft lb. in. in. in.

3

1.0000 1.0000 Center 12.0000

in. in. in. F

psi psi

No 0.500 2880.0000 1.050 110.000

in mile/hr

C

7.0000 84.0000

ft.

in.

N

( Diameter Basis for Vessel Shell Diameter Shell Length Tangent to Tangent Thickness of Shell Shell Joint Efficiency Head Type Head Thickness Head Joint Efficiency

11-24

ID

144.0000 20.0000 1. 0000 1.0000 Elliptical 1. 0000 1.0000

in. ft.

in.

in.

Stresses



COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------------------- Page 61 Horizves Analysis: 0101 zick Item: 1 9:51a Sep 21,2000

30.0000 12.0000 120.0000 0.3750 0.0000 O. 0000

Distance from Saddle to Vessel Tangent Saddle Width Saddle Bearing Angle Wear Pad Thickness Wear Pad Extension above Horn of Saddle Wear Pad Width Stiffening Ring Present

in. in. degrees in. in. in.

N

Results for HORIZVES Number

Description: 0101 Zick

1,

Shell Allowable Stress used in calculation Shell Compressive Yield used in Calculation Head Allowable Stress used in Calculation

20000.00 31750.00 20000.00

psi psi psi

Volume of Vessel Weight of Vessel, Empty Weight of Vessel, Full

2722.20 83924.85 84695.23

ft.**3 lb. lb.

Shell Head Shell Head

Thickness, Thickness, M.A.W.P. M.A.W.P.

Long. Long. Long. Long.

Stress Stress Stress Stress

Reqd. Reqd. Reqd. Reqd.

VS. VS. VS. VS.

Actual Actual Actual Actual

Required 0.851 0.847 200.00 200.00

Saddles Saddles Midspan Midspan

Actual 824].46 8162.25 8116.58 8289.13

Tangential Shear in Shell Tangential Shear in Head circ. Stress at Horn of Saddle Addl. Stress in Head as Stiffener Ring Compressive Stress in Shell

642.27 642.27 -3360.27 292.74 -1044.44

WIND ( ASeE

at at at at

Top Bottom Top Bottom

of of of of

Actual 1.000 1.000 240.88 2'12.3'1 Allowable 20000.00 20000.00 20000.0020000.00 16000.00 16000.00 -28575.00 5000.00 -15875.00

in. in. psig psig

psi psi psi psi psi psi psi psi psi

#7 J and SEISMIC RESULTS

Transverse Wind Load Component Ft {AFT '" CF '" GH * QZ 1 * 0.5 Ft Ft {391.311 * 0.5000 * 1.3967 '" 27.3514 1 * 0.5 Ft 3737.1382 lb. Saddle Reaction Force due to Wind Ft Fwt 1.5 * Ft * B / E Fwt 1.5 * 3737.1 * 84.0000 / 126.6562 Fwt 3717.7756 lb. Longitudinal Wind Load Component Fl FI {AFL"'CF*GH*QZl Fl ~ { 159.671 * 0.5000 '" 1.3967 '" 27.3514

Stresses

11-25

Slresses in Horizontal Pressure Vessels


COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName; Seminar -------------------------------------- Page 62 Horizves Analysis: 0101 Zick Item: 1 9:51a Sep 21,2000 FI

~

3049.8240 lb.

Saddle Reaction Force due to Wind pI Fwl FI * B I La Fwl 3049.8240 '* 84.0000 / 180.0000 Fwl 1423.2512 lb. Load Combination Results for Q + Wind or Seismic Q WO!2 + Max( Fwl, Fwt, FsI, Fst ) Q 42347 + Max( 1423, 3717, 0, 0 ) Q 46065.3945 lb. FORMULAS and SUBSTITUTIONS for ZICK ANALYSIS RESULTS Shell and Head Required Thickness and MAWP : TR TR TR

MAWP MAWP MAWP TR TR TR

MAWP MAWP MAWP

(D/2+CA)) / (S*E-O. 6,o,P) per UG-27 (c) (1) ( 200.00 '* ( 144.0000 ! 2 + 0.1250 )) / ( 20000 '* 1.00 - 0.6 * 200.00 ) + CA 0.851 in. (PI

Shell

(S'*E* (T-cA) ) ! ( (D!2+CA) +0.6* (T-CA)) per UG-27 (c) (1) : Shell ( 20000 '* 1.00 * ( 1.0000 - 0.1250 ») I «( 144.0000 I 2 + 0.1250 ) + 0.6 '* ( 1.0000 - 0.1250 )) 240.9 psig (p* (D+2*CA) '*K) / (2·S·E-0. 2·P) per App. 1-4 (c) Elli. Hd. (200.00 '"' ( 144.0000 + 2 * 0.1250 ) * 1.00 ) I ( 2 * 20000 * 1.00 - 0.2 * 200.00 ) + CA 0.847 in. (2*S*E·(T-CA)!(K,o,(D+2*CA)+0.2'*(T-CA.)I per App. 1-4 (c) ; Elli. Hd. ( 2 * 20000 • 1.00 '* ( 1. 0000 - 0.1250 )) I ( 1.00 * ( 144.0000 + 2 * 0.1250 ) + 0.2 *( 1.0000 - 0.1250 ») 242.3 psig

Longitudinal Bending (+-) at Midspan ( 3 '"' Q * L '* K.2 / ( PI * R~2 '* ( TS - CA »)) ( 3 * 46065 * 20.00 * 0.4464 ) I ( PI * 72.1250 * 72.1250 * ( 1.0000 - 0.1250 I)) 86.27 psi Longitudinal Bending (+-) at Saddle A ( 3 * Q * L '"' K.l / { PI '* R 2 * (TS CA I)) ( 3 * 46065 * 20.00 * 0.2101 ) / ( PI * 72.1250 * 72.1250 '* ( 1.0000 - 0.1250 I)) 40.60 psi

(

Tangential Shear in Shell near Saddle Q '"' K.5 / ( R * ( TS - CA I») 46065 * 0.8799 / ( 72.1250 • ( 1.0000 - 0.1250 ) 642.27 psi Tangential Shear in Head used as Stiffener

11-26

Stresses

Stresses in Ilorizontal Pressure Vessels


COADE Engineering Software PVElite 4.00 Licensee: COADE Inc" Local White Lock FileName: Seminar ---------------------------------------

Horizves Analysis; 0101 zick Q .., K.5 I ( R ... ( TH - CA 46065 * 0.8799 / ( 72.12 ... 642.27 psi

Item:

Page 63 9:51a Sep 21,2000

1

») ( 1.0000 - 0.1250 )

Circumferential Stress at Horn of Saddle _Q!(4*TEM*(SADWTH+l.56*SQRT(R*CA»)-12*Q*R*K.7!(12*L*TEB) -46065!1 4'" 0.8750 ...

( 12.00 + 1.56 ... SQRT( 72.1250 ,.. 0.8750 »))

-12.0 ... 46065 ... 72.12 ... 0.0130 -3360.27 psi

I ( 12 .., 20.0000 ... 0.7656 )

Additional Tension in Head used as Stiffener (Q/(R*(HDTHK_CA»)*(3/8)*(SIN(AL)A 2 /(PI-AL+SIN(AL"*COS(ALI)~ ( 46065 I ( 72 .1250 ... ( 1. 0000 - 0.1250 )))"* (3/8)"* ( SIN ( 1.99 ) -"-2)

I

( PI - 1.99 + SIN{ 1.99 ) ... COS ( 1.99 ))) 292.74 psi Circumferential Compression at Bottom of Shell (Q*( K.9/( TEM9"*( B+1.56 "* SQRT( R * TEM9 I)~) ( 46065*( 0.7603/( 1.2500*( 12.00+1.56*SQRT( 72.25* 1.2500)))) -1044. 44 psi Longitudinal Pressure Stress DP"*«SID/2+CA)-0.4"'(TS-CA»/{2.0*(TS-CA)) 200.0'" « 144.00 I 2 + 0.1250 - 0.4 '" ( 1.000 - 0.1250 )) / ( 2.0 * ( 1.000 - 0.1250 )) 8202.86 psi RESULTS for VESSEL RIBS, WEB and BASE Moment of Inertia of Saddle

Shell wearplate Web BasePlate Totals Value Value Value

C1 I As

Y

A

AY

0.4375 1.0625 6.1250 11.5000 19.1250

10.8438 0.0000 9.7500 12.0000 32.5937

4.7441 0.0000 59.7187 138. 0000 202.4629

K1

(1+Cos(b)-.5*Sin(b)A 2 >/(pi-b+Sin(b)"'Cos(b))

Fh

(K1*Q)=

Stresses

9375.3076

Ay"'2 2.0756 0.0000 365.7773 1587.0000 1954.8529 6.2117 776.1425 21.7500

SumQfAy/SumOfA SumQfAy"'2 + SumOfIo - C1*SumOfAy SumOfA - Ashell

10 0.6919 0.0000 77.2383 1.0000 78.9301 in. in*"'4 in

0.2035

lb.

Tension Str .. St Allowed Str .• Sa

( 2.0 ... Fh/As ) .6 * Yield Stress

862.0974 22800.0000

psi psi

Bending Mom. ,

Fh

19124.1465

ft.lb.

M

Bending Str., Shnd Allowed Str. , Sa

(

Lateral Direction

d ( M

.66

C1 I I ) Yield Str.

1836.6779 25080.0020

psi psi

11-27




FileName: Seminar

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

Horizves Analysis: 0101 Zick

Item:

Minimum Thickness of Baseplate Baseplate Min .. ~ (3*Q*F/{4*A*SA»)A. 5

1

Page 64

9:51a Sep 21,2000

0.3137

in.

Calc. of Axial Load, Inter. Values and Compo Stress

e e

BPLEN - 1 ) / ( NRIBS - 1) 168.0000 - 1 ) / ( 3 - 1 )

83.5000 in.

AP AP

e * BPWID / 2 83.5000 * 12.0000 /

P

AP

P

501.0000 * 22.8499 = 11447.7988 lb.

501. 0000 in

* SF BPWID - 1 - WEBTK ) * RIBTK + e/2 * WEBTK 12.000 - 1 - 1.000 ) * 1.000 + 83.5000/2 * 1.000

AR AR

se se

2

P/AR 11447.7988/ 51.7500

221.2135 psi

Check of Outside Ribs Inertia of Saddle, Outer Ribs - Axial Direction Y A AY Rib 6. 0000 10.0000 60.0000 Web 6.0000 41.7500 250.5000 Values 6.0000 51.7500 310.5000 KL/R Sea Sea

51.750 in

<

Sea

Ay.... 2 0.0000 0.0000 0.0000

10

143.9167 6.9583 150.8750

Cc ( 27.8921 < 122.7360 ) per AISC E2-1 9th Edition (1-(Klr) .... 2/(2*Cc.... 2))*Fy/{5/3+3*(Klrl/(8*Cc)-(Klr"'3)/(8*Cc"'3l (1-( 27.89 )"'2/(2 * 122.74"'2 ») * 38000 / ( 5/3+3*( 27.89)/(8. 122.74)-( 27.89 .... 3)/{8* 122.74"'3) 21148.51 psi

AISC Unity Check on Outside Ribs ( must be <= 1.0 ) Check Sc/Sca + Sb/Sba Check 221.21 / 21148.51 + 334.94 / 250S0.00 Check 0.02 Check of Inside Ribs Inertia of Saddle, Inner Ribs - Axial Direction Y A AY Rib 5.5000 10.0000 55.0000 Web 5.5000 83.5000 459.2500 Values 5.5000 93.5000 514.2500 KL/R Sea Sea Sea

11-28

<

Ay"'2 0.0000 0.0000 0.0000

10

(

110.8333 6.9583 117.7917

Cc { 9.5776 < 122.7360 1 per AISC E2-1 9th Edition (1- (Klr) ....2/ (2*Cc"'2) 1*Fy/ (5/3+3* (Klr) / (S*Cc) - (Klr .... 3) / (S*CC .... 3) (1-( 9.58 ) .... 2/(2 * 122.74 .... 2 )) * 38000 / ( 5/3+3*( 9.581/(8* 122.74)-{ 9.58"'31/(8* 122.74"'3) 22339.15 psi

Stresses



COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName; Seminar --------------------------------------- Page 65 Horizves Analysis: DIOI zick Item: I 9:51a Sep 21,2000

~ISC

Check Check Check

Unity Check on Inside Ribs ( must be <= 1.0 Se/Sea + Sb/Sba 244.87/ 22339.15 + 177.54 / 25080.00 0.02

The PV Elite program,

(cl 1989-2000 by COADE Engineering Software

(

Stresses

11-29

Siresses in JIorizonlal Pressure Vessels

11-30


Stresses


Chapter 12: Tall Vertical Towers ENGINEERING

ALLOWABLE STRESSES ON TALL TOWERS The allowable stress for the hoop direction is the same as always-the labulated stress for thc material. The allowable stress for longitudinal tension can be increased when calculaling cases Ihal include wind or earthquakc loads. The factor for this increase is 1.2 The allowable compressive stress (which may also be multiplied by the above factor) is calculated as follows: I.

Allowable strain, A, is O.125/(RO/t).

2.

Allowable stress, B, is taken from the appropriatc external pressure chart.

3.

Joint efficiency for butt-welded joints is taken as unity.

The temperature limitations on the external pressure charts also apply 10 the compressive allowable.

Tall Vertical Towers

t2-1

SOfTWARE

Analyzing Tall Vertical Process Towers


ANALYZING TALL VERTICAL PROCESS TOWERS The analysis ofta]] vertical process towers is somewhat more complicated than that for simple pressure vessels. There are several reasons for this additional complication: Pressure

- Internal

-

Dead Loads

Thermal Moments

Load Case

Internal Pressure

External Hydrotest Vessel Weights Equipment & Appurtenances Contents Temperature Effects Wind Earthquake Piping Erection External Pressure

Erection Operating Operating

t2-2

x X

Dead Loads

Wind! Seismic

X

X

X

X

x

X

X

X

Thermal



Design Procedure

DESIGN PROCEDURE In light of the various loadings found on towers, a very carefully thought out design procedure must be followed. The following list shows typical steps in a procedure of this kind:

(

I.

Analyze vessel geometry and determine the X and Y coordinate for elements.

2.

Analyze or design the vessel for internal pressure-determine initial thickness estimates.

3.

Analyze or design the vessel for external pressure-verify original thickness estimates, and modify if necessary. When thicknesses are modified, return to Step 2.

4.

Calculate the weights of elements and details. This would include shells and heads, nozzles, flanges, skirts, baserings, insulation, internals, platforms, fireproofing, packing, liquid, trays, and other miscellaneous items. This turns out to be one of the most tedious steps you will perform, even ifusing a computer program, since the geometry of each of these items must be determined.

5.

Determine the support forces and/or moments. Based on the weights of each component, determine the force and moment at the support point (or, in a more sophisticated analysis, the support points.)

6.

Calculate the cumulative weight and bending moment for the vessel. In the case of a tower, this is calculated both for the upright and the horizontal position, since the horizontal weight and bending moment diagram is used to determine the natural frequency.

7. Determine the natural frequency for the vessel. This is most often done using a numerical method such as Freese's method. In a more sophisticated analysis, this might be done with a finite element or finite difference analysis.

(

8.

Calculate the wind load on the vessel. This is done using the rules found in national standards such as American Society of Civil Engineers, Standard 7, the National Building Code of Canada, or the Uniform Building Code.

9.

Calculate the earthquake load on the vessel. THis is done with the same national standards described above.

10. Calculate the cumulative weight and bending moments for the various load cases. These include the various combinations of wind, earthquake, internal pressure, external pressure and weight described above, plus hydrotest cases. I I. Calculate the stresses and required thicknesses for the various load cases. It may be that the designer will have to increase the thickness of one or more vessel components at this point. If so, then the designer will have to begin all over with his calculations, starting from Step 2.

Wind load Computations Most wind load calculations will be done using a national standard such as the American Society of Civil Engineers, Standard 7, the National Building Code of Canada, or the Uniform Building Code. The ASCE Standard contains the following formula for wind load computation: Where:

F q,


~

force on each segment of the vessel wind pressure

12-3

Design Procedure

Pressure Vessel Design and Analysis· Seminar NOles

gust response factor structural force coefficient

Wind Pressure Computation Where: KZ qz I V

Exposure Coefficient Wind Pressure Importance Factor Wind Velocity

Where: zg Gradient Height a Power Law Coefficient z Elevation I.

2.

"Category" classification Structure All Structures Not Covered Below Primary Occupancy Of Over 300 Essential Facilities ~inor Structures "Importance" Factor Category 100 Miles F/ Hurricane I 1.00 11 1.07 III 1.07 IV 0.95

. Category I

11 1Il IV

At Hurricane Line 1.05 1.11 1.11 1.00

3.

"Exposure" Constants Category Alpha : Zg A - Large Cities 3.0 1500 B - Suburban Areas 4.5 1200 C - Open Terrain 7.0 900 D - Flat, Coastal 10.0 700 When the wind pressure has been calculated, the gust and structural factors for each element must be determined. The gust response factor can be either static or dynamic. The static just response factors are given in tables similar to those above, evaluated at the building or structural height h. The dynamic gust response factor is calculated based on the natural frequency of the vessel, various damping factors, and the design wind speed. Calculations for this factor can be found in the commentary to ASCE-7. The dynamic gust response factor must be used when the natural frequency of the structure is less than 1hz. If the natural frequency is unknown, then a guideline given by ASCE is to use the dynamic gust response factor when the height to width ratio of the vessel is greater than 5.

The structural force factor is also found in tabular form. The most commonly used table in ASCE 7 looks like this: Force Coefficients Jar Chimneys, Tanks, and Similar Structures, CJ

12-4


(


Design Procedure

CHar hiD values of Shape

Type of Surface

I

7

25

Square (wind normal)

All

1.3

1.4

2.0

Square (wind diagonal)

All

1.0

l.l

1.5

Hexagonal or octagonal

All

1.0

1.2

1.4

Round, highly loaded

Moderately smooth 0.5

0.6

0.7

Rough 0.7

0.8

0.9

Very RoughO.8 1.0

1.2

Round, lightly loaded

All

0.8

1.2

0.7

Earthquake Load Computation ASCE-7 and several of the other standards have a relatively simple calculation for the lateral force imposed by an earthquake: Where: Z

Seismic Zone Coefficient

I

Importance Factor

C

T

Fundamental Natural Frequency

K

Horizontal Force Factor

S

Soil Profile Coefficient

Note

C' S need not exceed 0.14 for profile 3 in zones 3 and 4; C need not exceed 0.12.

'Z''Seismic Zone Coefficient

(

Zone 4 3 2 I

Z I 3/4 3/8 3/16

o

1/8

'l''Gccupancy Importance Factor Category I


11 III

1.0 1.25 1.5

IV

N/A

12-5

Design Procedure


'S''Soil Profile Coefficient Profile

Type

Rock, Stiff Soil

SILO

S

Deep Cohesionless, Stiff Clay S2

1.2

Soft - Medium Clay

1.5

S3

'K'!Horizontal Force Factor

12-6

Structure Type

K

Unreinforced Masonry Walls Reinforced Masonry Walls Light Wood or Metal Walls Building Frame Systems Moment Resisting, Concrete

NlA 1.33 1.0 1.0 N/A

Moment Resisting, Ordinary Steel Moment Resisting, Special Frames Dual Systems Elevated Tanks

1.0 0.67 0.80 2.5

Other (Pressure Vessels)

2.0

(


Pressure Vessel Design and Analysis - Semill
Design Procedure

(

(

(


12-7


Design Procedure

Pressure Vessel Design and Analysis Seminar

Component Design Problem Tall Vertical Process Towers Use PVElite to analyze the tower shown on the following page. (You need not perform nozzle calculalions) Assume the tower full of liquid 10 the bottom of the cone (s.g.

~

.9)

Notes:

'<

j

Questions: Arc the thicknesses adequate? What arc the dimensions of the basering?

12-8

_


Pressure Vessel Design and Analysis ~ Seminar Notes

Design Procedure

T101 LIQUID PHASE REACTOR lCOO FT LBS 250 LBS 2'1 ELLJPTlCAL

• .,,5'

1

11 ...0000

t-f------l

MATERIALS: SHELL & SKIRT SA516.70 HEADS SA516.70 BASE PLATE SA285.C REINFORCEMENT SA516.70 FORGINGS SA105 NOZZLE NECKS PIPE SA106.C NOZZLE NECKS PLATE SA516.70

"~1----1

(

l-

DESIGN PRESSURE = 225 PSIG EXTERNAL PRESSURE = 15 PSIG DESIGN TEMPERATURE ~ 350 F CORRISION ALLOWANCE = 0.125 JOINT EFFICIENCY: LONGITUDINAL ~ 100% CIRCUMFERENTIAL ~ 85%

NOZZLE SCHEDULE: MARK SIZE PURPOSE A-D 20" MANWAY E 18" INLET OUTLET F 32"

I-----------l o.~375·

H"'WOO

----~~----""T~ .2\1 ELLlPTlCAL O.~375·

F

0.625'

Figure 1--Geometry for Tower Design Example English


12-9

Design Procedure

Pressure Vessel Design and Analysis· Seminar Noles

101 LIQUID PHASE REACTOR 1.62.691e+006 1112

N-I'lM

N

2'1 ELLlPTICAL 16

MH

DESIGN PRESSURE = 1.55 N/mm' EXTERNAL PRESSURE = 0.103 N/mm' DESIGN TEMPERATURE = 177 C CORRISION ALLOWANCE = 3.175 mm JOINT EFFICIENCY: LONGITUDINAL = 100% CIRCUMFERENTIAL = 85% MATERIALS: SHELL & SKIRT SA516.70 HEADS SA516.70 BASE PLATE SA285.C REINFORCEMENT SA516.70 FORGINGS SA105 NOZZLE NECKS PIPE SA106.C NOZZLE NECKS PLATE SA516.70

'''''

2-4

t-iN

2.1

ELLIPTICAL

24

I~M

16

NOZZLE MARK A-D E F

SCHEDULE: SIZE PURPOSE 20" MANWAY 18" INLET 32" OUTLET

MM

Figure 2-Geamelry for Tower Design Example Metric

12-10


\


(

Skirl and BaseRing Design

SKIRT AND BASERING DESIGN Basering Thickness Calculations Thickness of basering under compression: The equalion for the thickness of the basering is Ihe equalion for a simple cantilever beam. The beam is assumed to be supported al the skirt, and loaded with a uniform load caused by the compression ofthe concrete due to the combined weight of the vessel and bending moment on the down-wind I down-earthquake side of the vessel. The equation for the cantilever is found in most of the common vessel design textbooks, including Jawad & FaIT, Structural Analysis and Design of Process Equipment, page 434, formula 12.12:

Where:

(

fc

bearing stress on the concrete

J

cantilever length of basering

J

allowable bending stress of basering (typically 1.5 times Code allowable).

There are two commonly accepted methods of determining the bearing stress on the concrete. The approximate method simply calculates the compressive load on the concrete, assuming that the neutral axis for the vessel is at the centerline. Thus the load per unit area of the concrete is, from Jawad & FaIT equation 12.1, equal to·

Where:

(

W

Weight of vessel (worst case).

M

Bending moment on vessel (worst case).

A

Cross sectional area of basering on foundation

c

Distance from the center of the basering to the edge

J

Moment of inertia of the basering on the foundation

However, when a steel skirt and basering are supported on a concrete foundation, the

behavior of the foundation is similar to that of a reinforced concrete beam. If there is a net

Tall Vertical Towers (

I

12-11



bending moment on the foundation, then the force upward on the bolls must be balanced by the force downward on the concrete.

T 11'---"

'~ •

.

•

. ..

~

(. )

.. ~ 'WL! fll

t

i

I

C

ICI

Ie

(

,

d d ( b)

Figure 3--Neutral Axis Shift Because the concrete and the steel have different elastic moduli, and because the strain in the concrete cross section must be equal to the strain in the basering at any specific location, the neutral axis of the combined bolt/concrete cross section will be shifted in the direction of the concrete. Several authors, including Jawad & FaIT (pages 428 to 433) and Megyesy (Pages 70 to 73) have analyzed this phenomenon. In the program, we have used the formulation of Singh and Soler, Mechanical Design of Heat Exchangers and Pressure Vessel Components, Pages 957 to 959. This formulation seems to be the most readily adaptable to computerization, as there are no tabulated constants. Singh and Soler provide the following description of their method: In this case a neutral axis parallel to the y axis exists. The location ofthe neutral axis is identified by the angle a. The object is to determine the peak concrete pressure p and the angle a.

For narrow base plate rings, an approximate solution may be constructed using numerical iteration. It is assumed that the concrete annulus under the base plate may be trcated as a thin ring of mean diameter c. Assuming the foundation to be linearly elastic, and the base plate to be relatively rigid, Brownnell and Young have developed an approximate solution that can be cast in a form suitable for numerical solution. Let the total tensile stress area of all foundation bolts be A. Within the limits of accuracy sought, it is permissible to replace the bolls by a thin shell of thickness f and mean diameter equal to the bolt circle diameter c, such that t ~ A / pc. We assume that

/2-12




the discrete tensile bolt loads, acting around the ring, are replaced by a line load, varying in intensity with the distance from the neutral plane. Let n bc the ratio of Young's moduli oflhe bolt material to that oflhe concrete; II normally varies bctwcen 10 and 15. Assuming that the concrete can take only comprcssian (non-adhesive surface) and that thc bolts are effective only in tension (untapped holes in base plate), an analysis [similar to that given above) yields the following results: 2W+ P21ca p ~ -:p'-,(7:,-,_-=-:.,)-c a ~ 2(M - Wp,c)

P2 P)/C

2

as -"~-' ex::: c o ( a+n

( Where: I,

width of basering (similar to I in Jawad & Farr's equations above)

c

bolt circle diamctcr

p

four constants based on the neutral axis angle, and defined in Singh & Soler equations 20.3.12 through 20.3.17, not reproduced here. These equations give the required 7 non-linear equations to solve for 7 unknowns, namely p, s, a, and thc r; (i = 1,4) parameters. The simple iteration scheme described below converges rapidly. The iterative solution is started with assumed values of sand p; say So and Po' [the program takes thesc from the approximate analysis it has just performed). Then a is determined via the above equation, and then the dimensionless parameters Tj, T2' T)) and £4 are computed. This enables computation of corrected values ofp and s (say Po' and so'). The next iteration is started with s, andp, where we choose the following: l71 ::;

(

05 ( l70 +

l7n

P, = 05 (PO + pi)

This process is continued until the errors e; and e; at the ith iteration stage are within specified tolerances, (e; = e; = 0.005 is a practical value), where

Actual numerical tests show that thc convergence is uniform and rapid regardless of the starting values of So and Po'

(


12-13

Skin and BaseRing Design


Once the new values of bolt slress and bearing pressure are calculated, the thickness of the basering is calculated again using the same fonnula given above for the approximate method.

Thickness of Basering under Tension On the tensile side, ifthere is no top ring but there are gussets, there is disagreement on how to do the analysis. For example, Brown & Root does not look at the tensile side at all, Megyesy uses a "Table F" to calculate an equivalent bending moment, Dennis R. Moss uses the same approach but gives the table (Page 126-129), and Jawad & FaIT use a "yieldline" theory (Page 435-436). Since Jawad & FaIT is both accepted and explicit, Ihe program uses their Equation 12.13:

3.91F

t= S,

2b,,21~ +

-

d(~ + ..!.-) ,,2/

Where:

= Allowable Stress * Area

F

Boll Load

a

Distance between gussets

b

Width of base plate that is outside of the skirt

I

Dislance from skirt to bolt circle

d

Diameler of bolt hole

Thickness of Top Ring under Tension Ifthere is a top ring or plate, its thickness is calculated using a simple beam fonnula. Taking the plate to be a beam supported between two gussets with a point load in the middle equal to the maximum bolt load, we derive the following equation:

t=lM ,7

Where:

M

Bending moment from Megyesy, beams, case II, fixed beam. 2 F,

e.

80 FI

Bolt Load ~ Allowable Stress time Area

s

Allowable stress, 1.5

z

Seclion Modulus, from Megyesy, Properties of Seclions

* plate allowable

z Wt

12-14

( Do2 _D,2. _d.)

Width of Section




Required thickness of gussets in tension: Iflhere are gussets, they must be analyzed for both tension and compression. The stress fonnuJa in tension is just the force over the area, where the force is taken to be the allowable bolt stress times the boll area, and the area of the gusset is the thickness of the gusset times one half the width of the gusset (because gussets normally taper). . Required thickness of gussets in compression: In compression (as a column) we must iteratively calculate the required thickness. Taking the actual thickness as the starting point, we perfonn the calculation in AISC 1.5.1.3. The radius of gyration for the gusset is taken as 0.289 t per Megyesy (Fifth edition, Page 404). The actual compression is calculated as described above, then compared to the allowed compression per AISC. The thickness is then modified and another calculation performed until the actual and allowed compressions are within one half of one percent of each other.

Basering Design Selections

(

Selection of Number of Bolls: This selection is made on the basis of Megyesy's table in Pressure Vessel Handbook (Table C, Page 67 in the fifth edition). Above the diameter shown, the selection is made to keep the anchor boll spacing at about 24 in. Calculation of Load per Boll: This calculation is made per Jawad& Farr, Equation 12.3: W

2M

p= - - - N NR

Where:

W N

Number of bolls

R

Radius of boll circle

M

Bending moment

Weight of vessel

Calculation of Required Area for Each Bolt This is just the load per boll divided by the allowable stress.

Selection of the Bolt Size

(

The program has a table of boll areas, and selects smallest boll with area greater than the area calculated above.

Selection of Preliminary Basering Geometry The table of bolt areas also contains the required clearances in order to successfully tighten the selected boll (wrench clearances and edge clearances). The program selects a preliminarybasering geometry based on these clearances. Values selected at this point arc the boll circle, basering outside diameter, and basering inside diameter.

Analysis of Preliminary Basering Geometry Using the methods described above for the analysis section, the program determines the approximate compressive stress in the concrete for the preliminary geometry.

Selection of Final Basering Geometry If the compressive stress calculated above is acceptable, then the preliminary geometry becomes the final geometry. lfnot, then the boll circle and bascring diameters arc scaled


12-15


Skirt and BttseRing Design

up to the point where the compressive stress will be acceptable. These become the final basering geometry values.

Analysis of Basering Thicknesses The analysis then continues through the thickness calculation described above, determining required thicknesses for the basering, top ring, and gussets.

Skirt Thickness Calculations Basic Skirt Thickness The required thickness of the skirt under tension and compression loads is determined using the same formula used for the compressive stress in the concrete, except using the thickness of the skirt: W

Me

A

I

lZ = - - - -

,

Where: W

Weight of vessel (worst case)

M

Bending moment on vessel (worst case)

A

Cross sectional area of skirt

c

Distance from the center of the basering to the skirt (radius of skirt)

1

Moment of inertia of the skirt cross section

In tension this actual stress is simply compared to the allowable stress, and the required thickness can be calculated directly by solving the formula for t. In compression, the allowable stress must be calculated from the ASME Code, per paragraph UG-23, where the geometry factor is calculated from the skirt thickness and radius, and the materials factor is found in the Code external pressure charts. As with all external pressure chart calculations, this is an iterative procedure. A thickness is selected, the actual stress is calculated, the allowable stress is determined, and the original thickness is adjusted so that the allowable stress approaches the actual stress.

Stress in Skirt due to Gussets or Top Ring If there are gussets or gussets and a top ring included in the base plate geometry, there is an additional load in the skirt. Jawad & Farr have analyzed this load and determined that the stress in the skirt due to the boll load on the base plate is calculated as follows:

Where: F

Total load in one bolt ~ load on one gusset

b

Width of the gusset at the base thickness of the skirt

12-16



h


height of the gusset

Jawad & Farr note that this stress should be combined wilh the axial stress due to weight and bending moment, and should then be less lhan three times the allowable stress. They thus categorize this stress as secondary bending. The program performs the calculation of this stress, and then repeats the iterative procedure described above to determine the required thickness of the skirt at the lop of the basering. Generally the skirt/head/shell weld is a non-inspectable weld detail. Frequently the joint efficiency oflhis weld is taken to be as Iowa 0.45, sometimes 0.55 (See UW-12). Thus the skirt thickness may very well be governed by the loads at the top, even though the bending moment is higher at the bottom.

r

TTA

•I

I

HG

-L

DS

I

II I

I

-rl.

DC

DC

! ..

DO

Ih-'~--BND

I I TGA

1-=

l-

I

I

I I

I

I

Figure 4--Geometry for Baserings and Bolt Chairs

(


12-17

Skirl and BaseRing Design


(

12-18



Skirl and BnseRing Design


Baserings: Use the Basering program to analyze the support for tower shown on Page 198. You will need the following additional information: Basering ID ~ 118 in. (2997 mm.), OD~128 in. (3251 mm.), Thickness ~ I in. (25.4 mm), SA-516 70 at 100°F (37.7°C) Top Ring Thickness ~ I in. (25.4 mm), SA-193 B7 on a 124 in. (3150 mm) Bolt Circle Compressive strength of concrete, 3,000 psi (20.68 N/mm 2) Gussets, I in.(25.4 mm) thick by 12 in. (304.8 mm) tall, 2 per bolt. (locate 3 in. (76.2 mm) to cach side of bolts) Gussets arc SA-516 70 at 100°F (37.7°C), Elastic Modulus ~ 29,000,000 (19947 N/mm 2).

Skirt is 0.625 in. (15.9 mm) thick, 121.875 in.(3096 mm) OD, SA 516 70 at 100°F (37.7"C) From PYElite, get the weights and moments for the tower: (thcse have been increased to make the problem more interesting) Dead weight ~ 200,000 Ib (839672 N) Operating weight ~ 250,000 Ib (1.I1e 6 N) Test weight = 400,000 Ib (1.77e6 N) Operating Moment = 1,000,000 ft Ib (1.335e 9 Nmm2) Test Moment ~ 1,000,000 ft Ib (1.335e9 Nmm2) Notes:

Questions: Are thc thicknesses chosen adequatc?


12-19



COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------------------- Page 76 Basering Analysis: PVD&A Item: 1 9:51a Sep 21,2000

Input Echo, Sasering Item

1,

Analyze or Design the Sasering Design Temperature for sasering Thickness of Sase ring Sasering Material Sasering Operating Allowable Stress Sasering Ambient Allowable Stress Inside Diameter of sasering Outside Diameter of sasering Nominal Diameter of Bolts Bolt Material Bolt Operating Allowable Stress Salt Ambient Allowable Stress Number of Bolts Diameter of Bolt Circle Nominal Compressive Stress of Concrete Thickness of Gusset Plates Temperature for Gusset plates Average Width of Gusset Plates Material for Gussets Gusset Plate Elastic Modulus Gusset Plate Yield Stress Height of Gussets Distance from Bolts to Gussets Number of Gussets per bolt Thickness of TOp Ring/Plate Radial Width of the Top Ring/Plate

Description: PVD&A AORD RNGTMP

TBA BASOPE SASAMB

Dr DO

Dead Weight of Vessel Operating Weight of Vessel Test Weight of Vessel Operating Moment on sasering Test Moment on Sasering

F in. psi psi in. in.

SAHAMB NGIV DC

psi psi

FPC

3000.0

psi

TGA

CG

1.0000 100.0000 2.0000 8A-516 70 29000000.0 38000.00 12.0000 3.0000

NG

2

TTA

1.2500 4.0000

in. in.

0.0000

in.

SA

GUSTMP AVGWDT E

SY HG

TOPWTH

TS SKTTEM Stress Stress at Base at Bottom Head Weld

100.00 1.0000 8A-516 70 20000.00 20000.00 118.0000 128.0000 0.8750 81\-193 87 25000.00 25000.00 20 124.0000

BND

External Corrosion Allowance Skirt Thickness Skirt Temperature Skirt Material Skirt Operating Allowable Skirt Ambient Allowable Outside Diameter of Skirt Outside Diameter of Skirt Joint Efficiency of Skirt

AM

SKTOPE SKTAMB DS SKTHED ARCJNT

DW ROW TIl

ROM TM

0.6250 100.0000 SA-516 70 20000.00 20000.00 121. 8750 121.8750 0.7070 200000.0 250000.0 400000.0 1000000.0 1000000.0

in.

in.

in. F

in. psi psi in.

in.

in. p

psi psi in.

in.

lb. lb. lb. ft.lb. ft.lb.

RESULTS FOR BASERING ANALYSIS : ANALYZE OPTION

12-20


\

Pressure Vessel Design ami Analysis - Seminar Noles

Skirl and BnseRing Design

COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName : Seminar --------------------------------------- Page 77 Basering Analysis: PVD&A Item: 1 9:51a sep 21,2000 Calculation of Load per Bolt, Dead Weight Condition: W/Bolt ({ 4 • M/De ) - W ) / RN per Jawad Farr, Eq. 12.3 W/Bolt ({ 4 * 0.120E+08 / 124.000) - 200000 ) / 20 W/Bolt 9354.8389 lb. Required Area for Area Available in Area Available in Bolt Stress Based

Each Bolt, Based on Max Load a single Bolt all the Bolts on Approximate Analysis

Concrete Contact Area of Base Ring Concrete Contact Section Modulus of Base Ring

(

0.3742 0.4190 8.3800 22326.58

in in in psi

1932.08 57185.03

Calculation of Concrete Load, Test Weight Condition: SC ((ABT*SA+W)/CA) + M/CZ per Jawad & Farr Eq. 12.1 SC {( 8.3800· 25000+ 400000)/ 1932.08) + 0.120E+08/ 57185.03 SC 525.31 psi Calculation of sasering Thickness, (Simplified): TB RW * SQRT( 3 • SC / S ) + CA per Jawad & Farr Eq. 12.12 TB 3.0625 * SQRT{ 3 • 525 / 30000 ) + 0.0000 TB 0.7019 in. Results of Neutral Axis shift Calculation: Searing Pressure on Concrete Stress in Bolt

412.50 0.00

psi psi

Calculation of Sasering Thickness, (N.A. Shift): TBNA RW * SQRT{ 3 • SCNA / S) + CA per Jawad & Farr Eg. 12.12 TBNA 3.0625 * SQRT( 3 • 412 / 30000 ) + 0.0000 TBNA 0.6220 in. Required Thickness of Top Ring/Plate in Tension: (Calculated as a fixed beam per Megyesy) FT (SA*ABSS) , Bolt Allowable Stress * Area RM (FT*2.0*CG)/B.0. Bending Moment (1.5*BASOPE), Allowable stress· 1.5 SS WT (TOPWTH BND), Width of Section T T T

SQRT ( 6 • RM / ( SB * WT » + CA SQRT( 6 .7856 / ( 30000 • 3.1250 )) + 0.0000 0.7091 in.

Required Thickness of Gusset in Tension: T = ( SA*ABSS )/( NG*S*( AVGWDT ») + CA Required thickness based on average cross-section Actual thickness as entered by user

0.1309 1.0000

in. in.

Required Thickness of Gusset in Compression, per AISC E2-1 9th Ed. 1. Allowed Compression at Given Thickness: Factor KI/r Per E2-1 83.0450 122.7360 Factor Cc Per E2-1 All. Buckling Str. per E2-1 15572.10 psi

(


12-21

Skirt ami BaseRing Design


COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------------------- Page 78 Sasering Analysis: PVD&A Item: 1 9:51a Sep 21,2000 Act. Buckling Str. at Given Thickness

2618.75

psi

0.4931

in.

168.4202 122.7360 5264.57 5310.98

psi psi

SUMMARY OF BASERING THICKNESS CALCULATIONS Required Basering Thickness (simplfied) Required Basering Thickness (N.A. Shift) Actual Basering Thickness as entered by user

0.7019 0.6220 1. 0000

in. in. in.

Required Top Ring/Plate Thickness as a Fixed-Beam Actual Top Ring Thickness as entered by user

0.7091 1.2500

in. in.

Required Gusset thickness, + CA Actual Gusset Thickness as entered by user

0.4931 1. 0000

in. in.

Required Gusset thickness, + CA 2. Allowed Compression at Calculated Thickness: Factor KI/r Per E2-1 Factor Cc Per E2-1 All. Buckling Str. per E2-2 Act. Buckling Str. at Calculated Thickness

TENSILE STRESS CALCULATIONS FOR SKIRT AT TOP HEAD: S Mil PI*R"2*T I - F/ (2*PPR*T) Operating Dead Load 60.9375 Skirt Rad. Given by User 60.9375 0.6250 Skirt Thkn. Given by User 0.6250 Bndg. Mom. Given by User 1000000.0 1000000.0 Axial Force Given by User 250000.0 200000.0 601.1 Actual Stress in Skirt 810.1 Allowed Stress in Skirt 11312.0 11312.0 0

THICKNESS CALCULATION FOR SKIRT: Required Thickness 0.0332 Actual Thickness as Given 0.6250

N/A in. 0.6250 in.

0.0448 0.6250

COMPRESSIVE STRESS CALCULATIONS AT BASE OF SKIRT: S = M/IPIrR"2*T) + F/(2*PIrRrT) Operating Dead Load Skirt Rad. Given by User 60.9375 60.9375 Skirt Thkn. Given by User 0.6250 0.6250 Bndg. Mom. Given by User 1000000.0 1000000.0 Axial Force Given by User 250000.0 200000.0 Actual Stress in Skirt 2690.5 2481. 6 Allowed Stress in Skirt 13426.0 13426.0

Test Load 60.9375 in. 0.6250 in. 1000000,0 ft.lb. 400000.0 lb. 3317.4 psi 20139.0 psi

THICKNESS CALCULATION FOR SKIRT AXIAL COMPRESSION: Required Thickness 0.2378 0.2284 Actual Thickness as Given 0.6250 0.6250

0.2156 in. 0.6250 in.

SUMMARY OF SKIRT THICKNESS: Req. Thickness, Tension Req. Thickness, Camp. Actual Thickness as Given

Test Load N/A in. 0.2156 in. 0.6250 in.


12-22

Test Load 60.9375 in. 0.6250 in. 1000000.0 ft. lb. 400000.0 lb. -25.7 psi 16968.0 psi

Operating 0.0332 0.2378 0.6250

Dead Load 0.0448 0.2284 0.6250

{cl 1989-2000 by COADE Engineering Software


Pressure Vessel Design and Analysis· Seminar Noles

Chapter 13: Vessel Legs, Support Lugs, and Lifting Lugs These design caleulations are done with simple hand methods. It is useful to review the origin of the calculations.

VESSEL LEGS The loading for vessel legs is the weight of the vessel, including operating weight or weight of water for hydrotest, plus the horizontal shear and overturning moment due to wind, earthquake, or external loads. A vessel must have at least three legs. A vessel on two legs develops unacceptably high bending stresses in the legs under even small horizontal loadings. For the column furthest away from the neutral bending axis, the axial force in the column due to these loads, per Jawad & Farr 12.3, is W N

2M NR

p= - - - - -

Where: W

N R M

Note

Weight of vessel and contents Number oflegs (columns) Radius of column circle Bending moment due to wind or earthqu~ke loads

Because the bending moment is only effectively carried by legs that are quite far removed from the neutral axis, the total number of legs is divided by two in this equation.

To generate the stress in the leg, simply divide the force by the cross sectional area of the· leg. For the column closest to the neutral axis, the axial force does not contain the bending moment term. However, the axial load may be increased by cross-bracing of the legs. Cross bracing resolves the shear at the top of the legs into an axial load. The shear at the top of one leg is the horizontal load divided by the number oflegs which support the shear load. As implemented in Codecale, the number ofeffective legs for shear is the integer part of the number of legs divided by two, rounded up. For example, if you have 4 legs, the number oflegs effective for shear is two, but if you have three legs the number of effective legs is still two.

Vessel Legs, Support Lugs, and Lifting Lugs

13-1


Vessel Legs

The axial force can be computed from the horizontal force by resolving the force vectors. First, the horizontal force, assumed to be tangential to the vessel, is resolved into a force

into the vessel and a force in the plane from one leg to the next. The angle between the two legs is alpha, and the force X in that plane is H X=-sin it

Where: H

Horizontal force at top of one leg

a

Angle between tangent and next leg

a

=

180N -360 2N

This horizontal force in the plane of the legs is further resolved into an axial force in the cross bracing and an axial force in the leg (F): X

F=tantl'

Where: X

Horizontal force in plane of legs

P

Angle between top and bollom of legs ,.vDIN atan---

P

L

D

Diameter of leg circle

L

Length oflegs

When you have many legs, or if the legs are quite tall, the angle of the cross bracing will be high, and even a small horizontal load will require a large vertical component in the brace and leg to resist it. The leg closest to the neutral axis carries the weight of the vessel and the additional axial compression caused by the cross bracing:

w

p=---p N

Vessel legs are generally a simple structural shape - an I beam, a channel, an angle, or a structural tube. I-beams and channels have a strong and a weak orientation. If the beam is allached such that the tangent to the vessel is parallel to the beams strong axis use the strong axis beam properties. Otherwise use the weak axis properties. ' For angles, iflhe angle is allached with one leg welded to the vessel or one flat welded 10 the vessel, use the strong axis properties. If the angle is attached with both legs welded 10 the vessel (like an upside down V), use the properties of the angle in a diagonal direction.

13-2



Support Lugs

These leg loads are compared to the allowable leg comprcssion per the AISC Stcel Construction Manual, paragraph 1.5.1.3: KI

Wmn-
F. =

5 3( ~) (~r -+------

3

gC,

8C;

Wmre: C =

,

(

l,,i'E F

K

>

Effective length factor Length

,.

Least radius of gyration

E

Young's Modulus of column material

F,

Yicld strength of column material

If you cross brace the legs, there will be no bending at the bottom. OthelWise the bending at the bottom will be equal to the length ofthe leg times the shear at the top of the leg closest to the neutral axis. The combination of bending and compressive stress is compared to the allowable per AISC 1.6.1 (unity check).

SUPPORT LUGS The lug support point should be set as close to the vesscl OD as possible, to reduce bending moments on the lugs and the vessel. Most support lug calculations also assume a minimum of three lugs, so that the lugs do not have to carry a bending moment. However, some vessels are supported on two wide lugs that can carry transverse bending. The force on one lug is the weight of the vessel divided by the number oflugs, plus the bending moment at the lug support point divided by the radius to the lug and the number of effective lugs. This is thc same fonnula shown above for the force on a leg far away from the neutral axis of the bending moment. Most support lugs have two gussets, equally spaced about a bolt hole. The distance between the gussets is used to calculate the bending in thc support plate.

(


t3-J


Lining Lugs

The bending stress in the bottom plate of a vessel support lug is calculated using a simple beam formula for a beam (the plate) on two simple supports (the gussets): M

17=-

Z

Wmre:

2

F

= W,lf,l 6 = Force on one lug

D gp

=Dist.ance between gussets

Wpi

= Width of supportlugboLlotn plaLe

f pi

= Thicknessof support plaLe

z

Bending stress in Lhe bottom plate ofa vessel support lug is compared to the AISC allowable bending stress. The stress in the gusset is one half the lug force divided by the gusset area. This compression is compared Lo the AISC allowable compression. If the gussets arc angled (as they would be iffor cases with no top plate), ealeulate the mean gusset widlh and usc that value in the gusset compression calculations.

Compressive stresses in support lug gussets arc compared to AISC compressive allowable, which is caleulaled using the same formulas given above for the legs as columns. The loads on the vessel shell should also be checked. These can be checked using the WRC- I 07 analysis technique.

LIFTING LUGS There are two orientations of lifting lugs: flat to the vessel and perpendicular to the vessel. Flat lugs arc generally welded below the top head scam and extend far enough above the scam for the lifting cables to clear the head and its nozzles. Perpendicular lugs (cars) are used to clear some obstruction at or ncar the top head (such as a l;>ody flange) by moving the support point away from the vessel shell. They arc also used as tailing lugs. Lifting lugs are normally subject to forces that arc "horizontal" in the natural coordinate system of the vessel during the early part of the lift, and then to purely vertical forces. For lifting lugs that are flat to the vessel shell, these tangential forces are usually not a problem. However, for lifting lugs that are perpendicular to the shell (ears), these forces can easily bend the lug, and should be avoided.

(

Forces normal to (away from) the shell should be avoided on lifting lugs that arc flat to the shell, since these will tend to peel the lug off the shell. In other words, don't pull on the lug in its weak direction. The most significant stresses in the lug usually occur in the welds which attach the lug to the vessel. You need to know the minimum leg width and the location of these welds. Then you need to calculate the cross sectional area and the shear stress on the welds that attach the lug to the vessel. The load for the direct shearing stress may calculated as the

13-4

Vessel Legs, Support Lugs, and Lifting Lugs (

Lining Lugs


square root oflhe squares of all Ihe loads on Ihe lug (W, Nand T). This is only slightly cOIlservativc. WeldArea

=(2 L\Uld

+

BWeld )( Twdi )

Wmre: LWld = Length of weld on sides of support B wl1 =

Widlh of weI d on boltom of support

TWld "" Thickness of welds

Load =

lw' + N' + T'

Wmre: W = WeighlLo.d

N

=Horizontal force nonn at to the vessel

T = Horizon.tal force tangent to the vessel

(

In addition to the direct shear stress, the shear stress at the ends of the welds due to the bending on the lifting lug is must be calculated for each load. First, calculate the section modulus of the weld group. Then calculate the shear stresses: Bending Stress Due to Normal Force:

Bending Stress Due to Weight:

Bending Stress Due to Tangential Force:

Z"

Height of Lifting Lug Length of weld on sides of support Offset vessel OD to center of hole Section Modulus, Longitudinal Direction

Z/c

Section Modulus, Circumferehtial Direction

Hlif L,'dd

O'if

Tensile, bending and shear loads are combined to calculate the stresses in vessel lifting lugs. The stresses in the lugs arc compared to ASME allowable for base material shear and fillet weld shear. Most lifting lugs have a circular hole located in a semi-circular arc of meta I. Shear across this section may control the lug design. The bending load on a flat lifting lug is caused by the force N that you entered normal to the vessel. The bending load on perpendicular lifting lugs is caused by the combination of


13-5

Lifting Lugs


Ihe vessel weight (W) and Ihe force you entered in Ihe tangential direclion. Nole Ihallhis combinalion calculates the slress only allhe highest slressed corner of the weld group. _

/

OFFSET OF UFllNG lOO

~

~IU F'f:RF'ENDlC

LUG

flATUFTllm

_-__'LU'G\

ARC

Of'S£[

RADIUS

o SlDEFU..1ET WELD

LEGs----+-<.1 CROSS BRAaNG-+-I-<~ LEG ORlENTATJON;

__=:::- WEAK

-SlRONG.

Figure 1--Geometry for Vessel Legs, Support Lugs, and Lifting Lugs

13-6


I.

Pressure Vessel Design find Analysis - Seminar Noles

Lining Lugs

Pressure Vessel Desigu and Analysis Seminar Component Design Problem

Vessel Legs Design vessel support legs for the following vessel: The design pressure is 200 psi (1.37 N/mm2), design temperature is 300°F 149°C. The diameter of the vessel is 42 in. (1067 mm) The height of the vessel itself is about 14.3 ft. (4359 mm) The weight of the vessel is 30,000 lbs. (133451 N) The wind load is 40 psf (0.002 N/mm 2) The centroid of the vessel area is 10 ft (3048 mm) above grade. The vessel has 4 legs Each leg is a W6x25 beam, oriented with the axis in the strong direction. The leg material is SA-285, C The vessel legs are 5 ft (720 mm) tall and are not cross-braced.

(

Notes:

( Questions: What is the result of the AISC unity check?



13-7

Lining Lugs


COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName: Seminar --------------------------------------Page 66 Leg & Lug Analysis 0102 Legs Item; 1 9:51a Sep 21,2000

Input Echo, Leg & Lug Item

Description: 0102 Legs

Design Internal Pressure Design Temperature for Attachment Vessel Outside Diameter Tangent to Tangent Length of Vessel Height of Bottom Tangent Above Grade

TEMP OD TANTAN

W W

Operating weight of Vessel (vertical load ) Erection Weight of Vessel (Lifting Analysis) Force Coefficient Additional Area Importance Factor Wind Velocity Exposure Catagory Use ASCE 7-95 Wind Code

Cf

200.00 300.00 42.0000 14 .3000 5.0000 30000.00 30000.00

I V

0.500 2880.00 1.050 110.000

Expcat

C

FF FH

0.00 0.00

•

Number of Legs Length of Legs Effective Leg End Condition Factor Material for Legs Yield Stress of Leg Material AIse Member Designation Leg Orientation to vessel Axis Are the Legs Cross-Braced

ORIENT XB

5.0000 1.00 SA-285 C 26500.00 W6X25 Strong NO

Occasional Load Factor (AISe A5.2)

OCCFAC

1 .33

AREA PWIND

13962.21 38.02 12.15

RESULTS FOR LEGS : Section Properties for the selected Member Cross Sectional Area for W6X25 Radius of Gyration ( strong axis ) Section Modulus (strong axis ) Overturning Moment at top of Legs Weight Load at top of one Leg Additional force in Leg due to Bracing

F

in. ft. ft.

lb. lb.

in mile/hr

No

Additional Horizontal Force on Vessel Location of Horizontal Force above Base Point

COMPUTED pARAMETERS: Effective Wind Area of Vessel Wind Pressure on Vessel ( ASCE #7 or User Location of Centroid above Base Point

psig

NLEG LLEG K

WH

lb. ft.

ft.

psi

in psf ft.

Description:D102 Legs 7.340 2.700 16.700 13179.9 7500.0 0.0

in in. in. **3 ft.lb.

lb. lb.

Shear at top of one Leg Vleg (V + FF '" ( Imax / Itot ) Vleg (1843.3 + 0.0 ) '" ( 53.5 / 140.93 ) Vleg 699.86 lb. Axial Compression, Leg futhest from N.A.

13-8


l

Lining Lugs


COADE Engineering Software PVElite 4.00 Licensee: COADE Inc., Local White Lock FileName : Seminar ------------~-------------------------- Page 67 Leg & Lug Analysis DI02 Legs Item; 1 9:51a Sep 21,2000 Sma Sma Sma

«(W/Nleg) + (Mleg/{Nlegm*Rn) I I . 0000) / (Aleg I 1. 0000») «( 30000/4) + ( 13179 /( 2 .., 2.02 I) .., 1.00001/ ( 7.340 .., 1. 0000) ) 1466.92 psi

Allowable Compo for the Selected Leg (KL/r <·Cc ) Sa (1- (kl/r) "2/ (2*Cc"2») *Fy / { 5/3+3"'(KI/r)/(8*Cc)-(KI/r~3)/(8*Cc~31 Sa (1-( 39.47 )"2/(2 ... 146.97'''2 I) ... 26500 / ( 5/3+3*( 39.47 1/(8* 146.97 )-( 39.47"3)/(8* 146.97"3) Sa 19249.06 psi Bending at the Bottom of the Leg closest to the N.A. 12.0000 / Smdsa ) S {Vleg * Rlngth S (699.86 * 5.00 • 0.0000 / 16.70 ) S 2514.48 psi Sb Sb Sb

(0.6 * Fy * Occfac ) (0.6 * 26500 * 1.33 ) 21147.00 psi

AISC Sc Sc Sc

Unity Check ( must be < or = to 1.00 ) (Sma/Sa)+(0.85·S)/«1-Sma/spex)*Sb) ( 1466/ 19249 )+( 0.85 .., 2514.476 1/«( 1 - 1466/ 127463 ) * 21147 ) 0.1785

Additional Results Forces/Moments for WRC107 Analysis: Weight only Max. Shear/leg Radial Load Max. Shear/leg Circumferential Shear Load / Leg Longitudinal Shear Moment/ Nlegs Longitudinal Moment Max. Shear/leg* Dist. Circumferential Moment Forces/Moments for WRC107 Analysis: Occasional Radial Load Max. Shear/leg Circumferential Shear Max. Shear/leg Longitudinal Shear Load / Leg Longitudinal Moment Moment/ Nlegs Circumferential Moment Max. Shear/leg· Dist. The PV Elite Program,

o. o. o.

lb. lb. lb. it.lb. ft.lb.

..700. 700. 10767. 2862. ;186.

lb . lb. lb. ft.lb. ft.lb.

7500. 1994.

(cl 1989-2000 by COADE Engineering Software

(

(


13-9

Lining Lugs


COAOE Engineering Software PVElite 4.00 Licensee: COAOE Inc., Local White Lock FileName: Seminar ----------------------------------- Page 68 Leg & Lug Analysis 0102 Lugs Item: 2 9:51a Sep 21,2000

Input Echo, Leg & Lug Item

2,

Description: 0102 Lugs

Design Internal Pressure Design Temperature for Attachment Vessel Outside Diameter

TEMP 00

Operating Weight of Vessel (vertical load ) Erection Weight of Vessel (Lifting Analysis)

W W

200.00 300.00 42.0000

0.500 2880.00 34.490

in psf

Additional Horizontal Force on Vessel Location of Horizontal Force above Base Point

FF FH

3000.00 0.00

N T

WLIF TLIF DLIP RLIF HLIF OLIF TWELD LWELD BWELD

3000.00 15000.00 SA-516 70 33600.00 Flat 6.0000 1.0000 1. 5000 3.0000 15.0000 0.5000 0.5250 8.0000 6.0000

OCCFAC

1.33

AREA PWIND WH

13962.21 34.49 12.15

COMPUTED PARAMETERS: Effective Wind Area of Vessel Wind Pressure on Vessel ( ASCE #7 or User Location of Centroid above Base Point

in. lb. lb.

Cf

Occasional Load Factor (AISC A5.2)

F

30000.00 30000.00

Force Coefficient Additional Area User defined Wind Pressure Use ASCE 7-95 Wind Code

Horizontal Force Normal to the Vessel Horizontal Force Tangent to the vessel Lifting Lug Material Lifting Lug Yield Stress Lifting Lug Orientation to Vessel Width of Lifting Lug Thickness of Lifting Lug Diameter of Hole in Lifting Lug Radius of Semi-Circular Arc of Lifting Lug Height of Lug from bottom to Center of Hole Offset from Vessel 00 to Center of Hole Minimum thickness of Fillet Weld around Lug Length of weld along sides of Lifting Lug Length of Weld along Bottom of Lifting Lug

psig

No

YIELD

lb. ft.

lb. lb. psi in. in. in. in. in. in. in. in. in.

"in psi it.

RESULTS FOR LIFTING LUGS : Description:D102 Lugs Held Group Inertia in the Longitudinal Direction 86.50 io**4 Weld Group Centroid distance in the Long. Direction 5.16 in. Weld Group Inertia in the Circumferential Direction 108.41 in I *4 Weld Group Centroid Distance in the Circ. Direction 3.53 in. Primary Shear Stress in the Welds due to Shear Loads: A A Ssll SQRT(W 2+T 2+N A 2)/(2*Lweld+Bweld)*Tweld) A Ssll SQRT( 30000 2+ 15000 A 2+ 3000 A 2)/((2* 8.0+ 6.0)* 0.5250) Ssll 2915.5B psi Shear Stress in the Welds due to Bending Loads ; Shlf = (N*IHlif-Lweld/2»*YLL/ILL+fW*OLIF*YLL/ILL)+(T*OLIP*YLC/ILC)

13-10

Vessel Legs, Support Lugs, and Lining Lugs

(


Lifting Lugs

COADE Engineering Software PVElite 4.00 Licensee; COADE Inc., Local White Lock FileName; Seminar --------------------------------------- Page 69 Leg & Lug Analysis 0102 Lugs Item; 2 9:51a Sep 21,2000 Shlf

~

Sblf

~

3000*( 15.000- 8.000/2)* 5.162/ 86.496 + 30000* 0.500* 5.162/ 86.496) + ( 15000* 0.500* 3.525/ 108.405) 3108.76 psi

Total Shear Stress vs. allowable Shear for Combined Loads St (Ssll + sbU ) St (2915.577 + 3108.760 ) St 6024.34 psi Sta Sta Sta

(0.4 * Yield * Occfac ) AISC Shear All. (.4 * 33600 * 1.33 ) 17875.20 psi

Secondary Shear Stress in the Welds due to Shear Loads: Unit Weld Section Modulus ( Uwsm ) (2*LWELD+WLIF)A 3 / 12 - LWELD A 2(LWELD+WLIF\)A 2 /(2*LWELD+WLIF) ~ 317.15 in .... 3 Loads on Welds due to Torsional Moment Fth T * (Hlif-(Lweld-Cent) * (Bweld/2J/Uwsm Fth 1715.56 lb./in. Ftv Ftv

T * ( Hlif-(Lweld-Cent») * Cent / Uwsm 2911.24 lb./in.

Fsv Fsv

T /

( 2 * Lweld + Wlif ) 681.82 lb./in.

Resultant Load on Weld Group Fr Sqrt( Fth"'2 + ( Ftv+Fsv )A 2 ) Fr = 3981.61 lb./in. Resultant Secondary Weld Stress Fws Fr / Tweld Fws = 7584.02 psi Allowable Resultant Secondary Weld Stress Psa ( . 4 " Yield * Occfac ) Psa = 17875.20 psi shear Shs Shs shs

Stress in Lug above Hole vs. Allowable Base Metal Shear A SQRT( W 2 + N A2 + T A2 ) / Sha SQRT( 30000 A2 + 3000"'2 + 15000 A 2 ) / 4.500 1483.31 psi

Sas Sas Sas

(0.4 * Yield * Occfac ) Shear Allowable (0.4 * 33600 * 1.33 ) 17875.20 psi

Pin Hole Bearing Stress Vs. Allowable Bearing Stress A pbs Sqrt( W 2 + NA 2 + T A 2 )/( Tlif * Dlif ) pbs = Sqrt( 30000"'2 + 3000"'2 + 15000 A 2 )/( 1.000 * 1.500


13-11

Lifting Lugs



FileName: Seminar --------------------------------------- Page 70 Leg & Lug Analysis 0102 Lugs Item: 2 9:51a Sep 21,2000 Pbs

22449.94 psi

Pba Pba Pba

I 0.75 I 0.75

.•

'field 33600 25200.00 psi

AISC Bearing All.

Bending Stress in Lug at Weld Vs. Allowable Stress A Fbs N*(HLIF-LWELD)/(WLIF*TLIF 2/6) Fbs 3000 *( 15.000 - 8.000 )!( 6.000 ,., I.DOO A 2 / 6) Fbs 21000.00 psi Fba Fba Fba

( 0.4 • Yield Occfac ) Shear Allowable ( 0.4 .., 33600 .., 1. 33 ) 17875.20 psi



(

13-12

Vessel·Legs, Support Lugs, and Lifting Lugs


Bibliography

Brief Bibliography of Pressure Vessel Texts and Standards This bibliography describes several of the commonly available texts and standards used by the author of CodeCalc to develop and support the program. This list will help you to identify resources you may need to effectively design or analyze pressure vessels: ANSI Standard A58.1 - 1982, Building Code Requirements for Minimum Design Loads in Buildings and Other Structures, American national Standards Institute, New York,1982. This standard provides the most commonly used design technique for cDlculating wind loads and earthquake loads on structures, including pressure vessels.

ANSI Standard Bl6.5, Pipe Flanges and Flanged Fittings, American National Standards Institute, New York. This is 1he standard for 'standard' flanges up to 24 inches in diameter. Provides flange geomelries and allowable pressures for the various classes oftlanges (150, 300, 400, 600, 900, 1500. and 2500) made from a variety of materials and over a wide range of temperatures.

ASME Boiler and Pressure Vessel Code, SECTION VIII, Division I, Rules for Construction of Pressure Vessels. July 1989. American Society of Mechanical Engineers, 345 East 47th Street. New York, N.Y., 10017. This is 'lhe Code'. The INTERNAL, EXTERNAL, NOZZLE and CONICAL programs are based exclusively on this document. The FLOHEAD and FLANGE programs are based primarjly on Ihis document. All of the allowable stresses used by these programs are also taken from this document.

ASME Code for Pressure Piping, B31, Chemical Plant and Petroleum Refinery Piping ANSIIASME B31.3, American Society of Mechanical Engineers, 345 East 47th Street. New York, N.Y., 10017. This is the piping code for refineries and chemical plants. The PIPE&PAD program is based on this code. In addition, this document has good tables of elastic modulus and coefficient of thermal expansion for many classes of materials.

Bednar, H.H., Pressure Vessel Design Handbook, Van Nostrand-Reinhold Co., Princeton, H.J., 1981.

Bibliography

Drief Bibliography of Pressure Vessel Texis and Standards


Bednar provides good calculation techniques for tall process towers and fair coverage ofa variety of other pressure vessel design problems.

Brownell, L.E. and Young, E.H., Process Equipment Design, John Wiley, New York, 1959. This is a classic reference on process equipment design, and contains many useful calculation techniques. However, many copies of Ihis book are in an unreviscd format Ihat contains errors in tables and formulas. We recommend comparing a given technique to some of the other texts before using it.

Farr, J.R. and Jawad, M.H., Structural Analysis and Design of Process Equipment, John Wiley & Sons, New York, 1984. This is lhe best recent book on pressure vessel design and analysis. The book covers a wide scope of design techniques, and presents the rational and use of the ASME Code techniques beller than any other pressure vessel textbook. (J.R. FaIT is on many of tile ASME Code commiltees). The book also provides a good balance of theory, practice, and example problems. Highly recommended.

Harvey, J. F. Theory and Design of Modern Pressure Vessels, 2nd Edition, Van Nostrand-Reinhold, Princeton, N.J. Harvey provides a basic overview of pressure component design, but little information on supports or other peripherals to the vessel. His sections on thick walled pressure vessels, aUlofretiage, and thermal stress are especilllly useful.

Manual of STEEL CONSTRUCTION, American Institute of Steel Construction, Chicago, III. Eighth Edition, 1980. This standard provides tables ofall beam sections and structural tubing, and provides calculations for allowable compression on columns, allowable stress on beams, and allowable combined loads. These are vel)' useful for support leg, gusset plate, and even flat plate cldculations. Later editions are also available.

Megyesy, E.F., Pressure Vessel Handbook,Pressure Vessel Handbook Publishing, Inc., Tulsa OK, 74135

( This is another very widely used book with a good combination of easy- to-usc formulas, examples, and tables of data, including pipe sizes and schedules, flange dimensions and weights for components.

Modern Flange Design. Bulletin 503, 7th Edition, Gulf and Western Taylor-Bonney Division, Southfield, Michigan. This is the best known bulletin on design of flanges, and includes all of the f1rmge calculation sheets commonly used for flange design. It also contains a good pmctical discussion offlange design and bolting, and a good table of bolt dimensions.

Moss, Dennis R., Pressure Vessel Design Manual, Gulf Publishing Company, HOlls-

2

Bibliography


Brier Bibliography of Pressure Vessel Texis
ton, TX, 1987. This relatively new book provides many different calculation procedures, and some calculation sheets, for most of file common teclmiques in pressure vessel design. A major weakness of the book is a lack of example problems illustrating the use oftlie leehniques. Some ofthe calculations for support lugs,lining lugs, and vessel legs in the LEG&LUG program are from this book.

Roark, R.J., and Young, W.C., Formulas for Stress and Strain, 5th Edition, McGraww Hill, New York, 1795. (Later editions also available). This well known reference book provides an abundance of [annulas for detennining the stresses in structural components. The book provides tables for beams, plates, shells, and many other types of components under many pmcticalloading cOlldifiollS.

Standards of the Tubular Exchanger Manufacturers Association, Seventh Edition, 1988. Tubular Exchanger Manufacturers Association. 25 North Broadway. Tarrytown, N.Y.,10591. This is Ihe standard used for tubcsheets (in the TUBESHT program) and channel covers (in the FLANGE program). This document also contains excellent tables of elastic modulus, coefficient of Ihennal expansion, thennal conductivity, and the bolt tables used in the FLANGE, TUBESHT, and FLOHEAD programs.

Shigley, J. E., Mechanical Engineering Design, McGraw Hill, New York, 1972. This is a 'standard' textbook for mechanical engineering students. It provides many basic fommlns for beams, plates, shells, bolting, screws, springs, shafts, and olher mechanical components you may need to analyze from time to time.

Wiclunan, K.R, Hopper, A.G, and Mershon, J. L., "Local Stresses in Spherical and Cylindrical Shells due to External Loadings," WRC Bulletin 107, Welding Research Council, New York, 1965 (revisions through 1979). This is the 'WRC-I07' technique whieh is widely used to detennine stresses in shells due to loads on nozzles and attachments. This analysis is implemented in the CYLNOZ and SPHNOZ progmms. Though widely used, the results of this analysis are !lot especially accurate.

Zick, L.P., "Stresses in large Horizontal Cylindrical Pressure Vessels on Two Saddle Supports," in Pressure Vessel and Piping Design Collected Papers 1927-1959, American Society of Mechanical Engineers, New York, 1960. This is the 'Zick' analysis, used in the HORIZVES program. The Ziek analysis is very widely used to calculate stresses in horizontal vessels.

Bibliography

3

Brief Bibliography of Pressure Vessel Texts and Standards


)

(

4

Bibliography

Pvelite Seminar Notes

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