Trevor Jee - Flange Stress Analysis Guidance

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E N G I N E E R I N G C O N S U L T A N T S

FLANGE STRESS ANALYSIS

GUIDANCE NOTES

APRIL 2005 TECHNIP 35 R01A.DOC



DOCUMENT CONTROL

REVISION HISTORY

Rev Date Descr ipt ion By Check

A 26/4/05 First issue AGK PACM

DISTRIBUTION

Mr Duncan Warwick

Technip Offshore UK Limited

Enterprise Drive

Westhill

Aberdeen

Aberdeenshire AB32 6TQ

Tel 01224 271703

Fax 01224 271271

E-mail [email protected]

ISSUED BY

Eur Ing Alan Knowles

Trevor Jee Associates on behalf of Jee Ltd

26 Camden Road

Tunbridge Wells

Kent TN1 2PT

Tel 01892 500 775

Fax 01892 544 735

E-mail [email protected]

© Trevor Jee Associates 2005. The moral rights of the author have been asserted. The text of this work may bequoted in any form (written, visual, electronic or audio), up to and inclusive of five hundred (500) words without

express written permission of the publisher. Notice of copyright must appear on either the title, copyright or

acknowledgements page of the work in which it is quoted, as follows: “Quotations from [name of work] are

copyright© 2005 Trevor Jee Associates. Used by permission”.

Technip 35 R01a flange stress analysis guidance.DOC 2/41



CONTENTS

1 INTRODUCTION......................................................................................................................5

1.1 BACKGROUND ...........................................................................................................................5 1.2 OBJECTIVES...............................................................................................................................5

2 CONCLUSIONS.......................................................................................................................6

2.1 ALLOWABLE STRESSES...........................................................................................................6

3 RECOMMENDATIONS............................................................................................................7

3.1 FUTURE WORK ..........................................................................................................................7

4 OVERVIEW OF DESIGN PROCESS.......................................................................................8

4.1 SCOPE.........................................................................................................................................8 4.2 LOADING ON FLANGES.............................................................................................................8 4.3 FLANGE FAILURE MODES ........................................................................................................9 4.4 DESIGN APPROACHES .............................................................................................................9 4.5 SUMMARY OF CODES AND METHODS .................................................................................10 4.6 ASME B16.5

22CLASS SPECIFICATION..................................................................................11

4.7 RECOMMENDED PROCEDURE..............................................................................................12

5 STRESSES USED IN ANALYSIS..........................................................................................14

5.1 MAXIMUM SHEAR STRESS.....................................................................................................14

5.2 VON MISES EQUIVALENT STRESS........................................................................................14 6 DESIGN METHODOLOGIES.................................................................................................17

6.1 ORDER OF CONSIDERATION AND OF DESIGN....................................................................17 6.2 NON-STANDARD FLANGES ....................................................................................................17 6.3 STANDARD WELD-NECK FLANGES.......................................................................................18 6.4 SWIVEL RING FLANGES..........................................................................................................18

7 DISCUSSION OF DESIGN METHODS .................................................................................20

7.1 OVERVIEW................................................................................................................................20 7.2 SIMPLIFIED STRESS ANALYSIS .............................................................................................21 7.3 THERMAL GRADIENT ANALYSIS............................................................................................23 7.4 LOW TEMPERATURE FLANGES.............................................................................................24

7.5 AGGRESSIVE ENVIRONMENTS INCLUDING SOUR SERVICE............................................24 7.6 FINITE ELEMENT ANALYSIS ...................................................................................................25 7.7 COFLEXIP PROGRAMME ........................................................................................................25 7.8 TECHNIP MATHCAD SHEETS .................................................................................................26




8 FLANGE GRADES, BOLTS AND GASKETS .......................................................................27

8.1 FLANGE MATERIAL GRADES .................................................................................................27

8.2 METRIC AND IMPERIAL STUD BOLT SIZES ..........................................................................28 8.3 BOLT TENSIONS ......................................................................................................................28 8.4 GASKETS ..................................................................................................................................30

9 RELEVANT READING...........................................................................................................36

10 REFERENCES.......................................................................................................................39




1 INTRODUCTION

1.1 BACKGROUND

Technip Offshore UK Limited (Technip) have been developing calculation methods

to verify flange and bolt stresses on externally loaded flanges. This has included

writing three MathCAD sheets1, &2

3 for weld neck and swivel ring flanges, which

cover both the PD 5500 4

and ASME VIII 5

codes. These sheets take as the starting

point that a flange has been selected, so the type, material, and pressure/temperature

rating are known. They then check whether that flange can withstand applied

bending and tension.

A software package developed by Coflexip 6 is also used by Technip to design fixed

weld neck and swivel ring flanges to ASME VIII 5 code, using RX, R and BX

gaskets. It again allows bending and tension to be applied at the joint but it ignores

temperature derating. It calculates the bolt torque required during make up and

stresses under operating and test conditions are examined.

Technip have asked Trevor Jee Associates:

1. To write the technical guidance on how Technip designs flanges.

2. To research allowable stresses and bolt loads and to include advice on this in the

technical guidance.

Trevor Jee Associates is an independent company which carries out pipeline

engineering consultancy and associated training for the oil and gas industry.

1.2 OBJECTIVES

The overall objective of this study is to produce a guidance document covering the

calculation methods used to verify flange and bolt stresses for externally loaded

weld neck and swivel ring flanges.

The main tasks are:

■

■ ■

Determine allowable stresses

Write guidance documentManage project




2 CONCLUSIONS

2.1 ALLOWABLE STRESSES

The stresses which are allowed depend upon the method being used.

With the standard flanges of ASME B16.5 22

and API TR 6AF 24

, there is no need to

examine the yield or permitted stress. The approach is to determine a classified

flange for a particular temperature and pressure rating. The flange manufacturer

selects suitable materials.

With ASME VIII 5, the yield and membrane stresses are listed for US steels. These

are derated depending on operating temperature.

With PD 5500 4, the permissible stresses are listed for BS and European steels.

These are derated depending on operating temperature.

With FEA, providing all external forces and moments have been considered for

installation, hydrotest and operational conditions, then it is permissible to assess the

equivalent von Mises stress in the bolts and flange with the corresponding material

yield value suitably de-rated for temperature.




3 RECOMMENDATIONS

3.1 FUTURE WORK

3.1.1 COFLEXIP METHODIt is recommended that a check be made on the existing Coflexip software. This

requires a review of the coding pages of manual.

From a reading of the Coflexip manual, it would appear that the minimum yield

stress rather than the de-rated permissible stress intensity may have been used.

It is recommended that a retrospective review be undertaken to clarify what

permissible stresses were used as input to their flange design. If the minimum yield

stress was used directly, then a list of currently installed flange locations should be

drawn up.

3.1.2 MATHCAD SHEETSIt would be convenient for designers to have assistance in selecting dimensions and

stresses. For this, it is recommended that look up tables be written in Excel to select

standard flange sizes, and to select suitable material stresses. These could be

embedded into MathCAD.

The existing MathCAD sheets are incomplete and unverified.

It is recommended the MathCAD sheets for weldneck and swivel ring flanges be

developed and verified.

3.1.3 THERMAL GRADIENTThe thermal gradient can introduce additional bending stress in the flange faces.

It is recommended that the critical temperature differential causing excessive

additional stresses due to a thermal gradient be identified.




4 OVERVIEW OF DESIGN PROCESS

4.1 SCOPE

This report limits itself to a discussion of subsea flanges for pipelines and

spoolpieces. These are weld-neck and swivel flanges in carbon steel. For subsea

use, it is common to use an iron or soft steel ring gasket located in a groove within

the bolt circle diameter.

Occasionally, corrosion resistant alloys are used for flanges and in this instance,

matching suitable materials are used for these rings. The consequences of thesematerials’ properties are discussed.

4.2 LOADING ON FLANGES

4.2.1 FLANGE FORCES AND LOAD COMBINATIONSSubsea flanges are subject to:

Internal pressure due to the contained liquid

External hydrostatic pressure

Axial loading due to residual lay tension and temperature effects in pipeline

Bending moments due to thermal effects or misalignment of spool piece

Thermal reduction in strength of flange and bolting material due to the

temperature of the contained liquidThermal stresses due to the temperature gradient between the inside and outside

walls of the flange

Shear across the joint, which can only be resisted by the gasket

It is normal to consider the following conditions as a minimum:

Bolt tightening at installation

Hydrotest

Operational conditions

4.2.2 DESIGN AND ASSEMBLY CONSIDERATIONSDuring initial assembly, there is no hydrostatic pressure differential and generally

zero or very low axial forces and bending moments. Bolts are tensioned to a statedvalue for gasket seating.




The joint is then pressure tested, during which some movement of either the pipeline

or spoolpiece may take place, subjecting the flanges to axial and bending

displacement. Finally, in operation, temperature effects and further expansion mayimpose additional axial and bending movement.

During these subsequent stages, the bolts are not normally adjusted in a subsea

situation – although onshore, this can occur. This means that the initial tension

applied to the bolts needs to be sufficient to maintain sufficient force on the gasket

to ensure a seal without any increased bolt tension resulting in yielding.

Any design process needs to verify as a minimum the three conditions of assembly,

hydrotest and operation.

Allowance should be made when assessing stresses for the reduction in cross-

sectional area at end of life due to corrosion.

4.3 FLANGE FAILURE MODES

4.3.1 INSTALLATION FAILURESFlange leakage can be caused during installation by many conditions, including:

Incorrect flange or groove facing

Incorrect gasket size or material

Dirty or damaged flange faces

Inaccurate or uneven bolt stress – reliance on the torque of the nuts rather than

measuring bolt extension may result in under- or over-tensioned bolts

These can be classified as lack of care in specification or installation.

4.3.2 OPERATIONAL FAILURESFlanged joints can fail by:

Overstressing of the flange material

Yielding of the bolt material

Leakage through the gasket seal interface

High vibration levels in adjacent pipeline and spool – causing repetitive damage

to the gasket and groove or loosening of the bolts

In many cases, a combination of material overstress eventually results in gasket

leakage. The last failure is due to inadequate support of the joint.

Flange failures do not result in rupture of the pipeline itself. However, for

hazardous product lines, any leak (however minor) should be considered as a

failure, and must therefore be taken seriously.

4.4 DESIGN APPROACHES

4.4.1 NON-STANDARD WELD-NECK FLANGESFor non-standard weld-neck f langes, it is necessary to:

Derive using ASME VIII 5 or PD 5500 4 or

Analyse using finite element analysis.




4.4.2 STANDARD WELD-NECK FLANGESFor standard weld-neck flanges, it is possi ble to:

Select from ASME B16.5 22

, API 605 28

or API TR 6AF 24

,

Derive using ASME VIII 5 or PD 5500

4 or


4.4.3 NON-STANDARD SWIVEL RING FLANGESThere are no standard sizes for swivel ring flanges larger than 346 mm (13

5/8in) of

API 17D 27

. It is necessary to:

For small well head flanges, select directly from API 17D 27

or

Derive using ASME VIII 5 or PD 5500

4 or


4.5 SUMMARY OF CODES AND METHODS

Each of the codes and methods has pros and cons:

4.5.1 ASME B16.5 22 Is limited to standard Classified US weld neck flanges only (<24in ND)

Does not allow for applied moments or tension

Cannot be used for high pressure flanges such as API 6A 23

Provides a list of ASTM steel grades for flanges and bolts

4.5.2 API 605 28 Equivalent to above

Is limited to standard US weld neck flanges only (26in to 60in ND)

4.5.3 API 17D 27

Equivalent to above for weld-neck and tapered swivel ring flanges for subseawellheads

Is limited to 5 000 and 10 000 class rating up to 346 mm (135/8in) ND

4.5.4 API TR 6AF 24 Is limited to standard API flanges only

Concentrates on small diameter at higher working pressures

Only method to explicitly allow for thermal gradient stresses

Uses a graphical method based on results of FEA. This is difficult to adapt into

a convenient spreadsheet for design over a range of temperatures and pressures

Reported that six of the standard flanges did not work! They leaked or yielded.

4.5.5 ASME VIII 5

Provides general approach to design of non-standard flangesUses yield and membrane stresses, derated depending upon operating

temperature

Provides tables for US steels

Latest version is in metric but with imperial bolts

4.5.6 PD 5500 4 Provides general approach to design of non-standard flanges

Uses design stresses*, derated depending upon operating temperature

(*equivalent to ASME VIII 5 yield and membrane stresses)

Provides tables for UK and European steels

Does not give formulae for derivation of F, V and f (though same as

ASME VIII 5

so the latter may be used)Requires reduction in stress for larger diameters (>1 m) See Clause 3.8.3.4.2.




4.5.7 COFLEXIP PROGRAMMECalculates non-standar d weldneck and swivel ring flanges

Based on ASME VIII 5

methodSeems to use yield strength rather than the correct allowable temperature-

derated values

The swivel ring deviates slightly from ASME VIII 5 loose ring analysis

4.5.8 EN 1092 31, EN 1591 32 & 33 & EN 1759 34 Not yet fully released

Provides standard dimensions of Classified metric DIN type and imperial-

converted ASME flanges

Method of analysing non-standard flanges and gaskets

4.5.9 FINITE ELEMENT ANALYSISCan calculate non-standard weldneck and swivel ring flanges

Can allow for all applied moments, thermal and tensile effectsCompares von Mises equivalent stress with temperature de-rated yield

Needs a substantial investment of time and effort to set up and run the model

It is the only method available if the temperature gradient effect needs to be

included

Can be used to design all flange shapes including the tapered swivel rings of

API 17D 27

4.6 ASME B16.5 22 CLASS SPECIFICATION

4.6.1 PROCEDURE AND LIMITATIONSMost flanges are specified to a class by the pipeline engineer rather than fully

designed by him. The manufacturer can then select suitable material and supply a

flange to that class; a flange that has been certified as adequate for the particular

service by the regulatory body.

ASME B16.5 22

covers standard flanges in carbon steel for use with pipelines. It

includes blank (blind) flanges, threaded, lapped and slip-on designs as well as the

weld-neck flanges that are being considered in this document. It does not cover

swivel ring flanges.

The latest version of the document is in SI units with imperial unit conversions in

Annex F.

The dimensions of the flanges are provided for a number of classes: 150, 300, 400,

600, 900, 1500, 2500. It is common to see these written as 150# etc, and they are

sometimes mistakenly referred to as 150 lb etc. Whereas there is somewhat of a

correlation between working pressure and Class (in pounds) at around 441°C

(825°F) for the more common materials, this is not a consistent for all materials and

flange Classes (especially Class 150 flanges).

The standard flanges are sized by consideration of the material, temperature and

pressure, in order to determine the Class. No consideration is made of moment or

tension from the pipeline, nor is there any allowance for thermal gradient through

the flange.





To accommodate these other stresses, it is common to go up one class for subsea

spools, where large moments or thermal loads may occur. That is to say, if

Class 900 has been calculated, then Class 1500 would be specified.

4.7 RECOMMENDED PROCEDURE

Figure 1 shows a simplified flow diagram of this recommended approach.

Initially, all internal pressures, temperatures and pipeline bore should be

determined.

Select a standard flange if possible as the basis for the design. However, it should

be noted that ASME B16.5 22

only specifies standard weld neck flanges up to

610 mm (24in) nominal diameter. Larger flanges must be to API 605 28

.

For standard flanges not subject to external moments or tension stresses, it is

possible to look up their requirements in ASME B16.5 22

tables for a suitable class.

There are no standards for the dimensions of swivel ring flanges other than the

limited selection of small wellhead designs of API 17D 27. However the swivel ring

flange must mate with the weld-neck flange and bolt sizes, so some standardisation

is necessary.

Use ASME VIII 5 or PD 5500

4 for design of non-standard weld necks or swivel ring

flanges. The preference on which of these codes is used depends upon the materials

being supplied and where the flanges will be installed. The former code should be

used for American steel and US waters; the latter code is better for European steel.

If the calculated stresses are too high compared with the de-rated maximum shear

stresses (Secton 5.1) then it will be necessary to revisit the initial size of flange and

reanalyse.

If there is a severe thermal gradient (which could cause additional bending stress in

the flange), then it may be possible to minimise this effect using insulation.

However, if this is not possible then it may be necessary to check the stresses using

Finite Element Analysis (FEA).

Only use FEA as a last resort. It is the only method that can capture all effects but

does involve a substantial design effort. Check the equivalent stress (Section 5.2)against yield. If this is too high then the flange dimensions need to be adjusted.

A fundamental consideration with all the above design procedures is the selection of

the correct bolts and gaskets. Guidance is given in Section 8.

It is recommended that the design should follow the three stages – find a similar

standard size of flange; develop using ASME VIII 5 or PD 5500

4; then check using

FEA if necessary.



Technip 35 R01a flange stress analysis guidance.DOC

Figure 1 Flow chart for flange design

flangetype?

nominaldiameter

flange dimensions known bore,determine temperature and pressure of product

weld-neck

26in to2in to 24in

use sizes from ASME B16.5 or

API RP6

swivel ring

match bolt diameterand positions of the

mating standard flange

look-uptable inExcel

flangestresses

noyes

sel AS

externalmomt orforce?

flangestresses

US/UKsteel?

US UK

design to ASME VIII design to PD 5500

severe

thermalradient?

MathCAD

insu-late?

yes no

yesend

noselect from API TR6 AF2or undertake FEA

graphsor

FEA



5 STRESSES USED IN ANALYSIS

5.1 MAXIMUM SHEAR STRESS

Both ASME VIII 5 and PD 5500

4 make use of the maximum shear stress. This is

compared with a basic allowable stress equal to the lesser of2/3 the yield or

1/3 the

ultimate.

The latter limit prevents rupture. It is an essential limit when following the method

set out in these two codes.

When calculations were undertaken by hand, this method was easier to use than the

equivalent stress method.

Maximum shear stress theory is slightly less accurate but is always conservative

compared to von Mises for ductile materials. However, von Mises method should

not be used for brittle materials that might be subject to rupture.

A full explanation of the ASME VIII 5 stress definitions and combinations is

described in Appendix 4-1 and Figure 4-130.1. See also Section 7.2 for a discussion

of the stress intensity design approach.

5.2 VON MISES EQUIVALENT STRESSFinite Element Analysis (FEA) commonly determines the equivalent stress from the

component stresses using von Mises equivalent stress theory. It then compares this

to an allowable stress equal to2/3 the minimum specified yield.

It is common to modify this allowable stress by a design case factor based on the

probability of occurrence of a particular load com bination. This may be greater or

lesser than 1.0. For example, API RP 2A-WSD7 permits an increase in allowable

stress (from 60% to 80% of yield) for stresses due to combinations of environmental

loads.

Von Mises method has been shown to more accurately predict the onset of yield forductile materials than the older method.




The three principal stresses should be calculated at all critical points of the flange.

At locations with axisymmetric geometry such as found in plain pipe, the principal

stresses will usually be in the axial, hoop and radial directions. (For non-axisymmetric geometry, the directions may be different.)

API 8 have defined that the principal stress components should be classified as one

of the following:

Any normal or shear stress that is necessary to have static

equilibrium of the imposed forces and moments. A primary

stress is not self-limiting. Thus, if a primary stress subsequently

exceeds the yield strength, either failure or gross structural

yielding will occur.

Membrane p is the average value across the thickness of asolid section excluding the effects of

discontinuities and stress concentrations. For

example, the general primary membrane stress in

a pipe loaded in pure tension is the tension

divided by the cross-sectional area. σ p may

include global bending as in the case of a simple

pipe loaded by a bending moment.

Primary

Bending b is the portion of primary stress proportional to

the distance from the centroid of a cross section,

excluding the effects of discontinuities and stress

concentrations.

Secondary q is any normal or shear stress that develops as a result of

material restraint. This type of stress is self limiting, which

means that local yielding can relieve the conditions that cause the

stress, and a single application of load will not cause failure.

Table 1 API definition of stresses

Combining the stresses at each cross-section is commonly done using von Mises

yield criterion as defined by the following equation:

( ) ( ) ( )

2

213

232

221 σ−σ+σ−σ+σ−σ

=σe

Where:

σe = von Mises equivalent stress

σ1, σ2, σ3 = principal stresses

5.2.1 API 6AF2 26 This adopts the same approach as ASME VIII

5 Division 2 Appendix 4. In Section

4.3, it recommends design stresses be limited to:

ST = 0.83 · SY and

Sm = (2/3) · SY




Where:

Sm = design stress intensity at rated working pressure

ST = maximum allowable general primary membrane stress intensity athydrostatic test pressure

SY = material minimum specified yield stress

Alternatively, it permits combining triaxial stresses based on hydrostatic test

pressures and limiting them to:

SE = SY

Where:

SE = maximum allowable equivalent stress at the most highly stressed distance

into the wall, computed by the distortion energy theory method

SY = material minimum specified yield stress

The distortion energy theory method is more commonly known as the von Mises

equivalent stress method.




6 DESIGN METHODOLOGIES

6.1 ORDER OF CONSIDERATION AND OF DESIGN

We will discuss non-standard flanges first because the design of these is a general

case. Standard flanges can then be understood as a special case.

However, it is recognised that common design practice is to start with a standard

flange and then check that it is suitable for any applied moments and forces.

6.2 NON-STANDARD FLANGES6.2.1 RECOMMENDED STEPS FOR DESIGNFollowing the simplified flow diagram of Figure 1, the following steps should be

undertaken for the design of non-standard weld-neck flanges:

Determine the diameter, temperature, pressure, moments and axial tension

Determine flange dimensions by selecting a standard flange of the correct bore


4 to pick suitable material and to calculate the

flange stresses. Use could be made of a MathCAD sheet or the Coflexip

programme 6

Increase flange dimensions if stresses exceed the maximum shear stress

(Section 5.1) and repeat analysis

Insulate to counter any severe thermal gradient or carry out a full FEA using theflange sizes as a basis.

6.2.2 ASME VIII 5 AND PD 5500 4 These codes consider in detail the pressure, moment and loads on the flange. They

provide a method of sizing weld-neck and the other flange types considered in

ASME B16.5 22

. But in addition, it provides a method of sizing swivel ring flanges

by adapting the lap joint flange design.

However, the method makes some assumptions:

For make-up there is no moment or tension in the system, only bolt tension.

There is no allowance for shear or thermal gradient stresses during operation.

6.2.3 COFLEXIP PROGRAMME 6 The Coflexip programme

6 generally follows the method described in ASME VIII

5

and PD 5500 4 to design non-standard flanges.




However, the appendix at the end of the guidance notes seems to indicate the use of

material yield stress rather than the temperature de-rated permissible stress. It is

important that the correct material property be used – namely the basic allowableshear stress. Refer to Section 5.1.

6.3 STANDARD WELD-NECK FLANGES

6.3.1 DESIGN USING STANDARD FLANGESStandard dimensions have been determined for weld-neck type flanges in US and

European pipe diameter.

With the standard flanges of ASME B16.5 22

and API TR 6AF 24

, there is no need to

examine the yield or permitted stress. The approach is to determine a classified

flange for a particular temperature and pressure rating. The flange manufacturer

selects suitable materials.

However, the tables assume that there are no applied moments or axial forces acting

on the flange. Nevertheless, it is possible to select a standard flange from the tables

and then check to ensure it will withstand the applied forces and moments in

operation. This would then assist with size selection for a non-standard flange

design.

6.3.2 RECOMMENDED STEPS FOR DESIGNFollowing the simplified flow diagram of Figure 1, the following steps should be

undertaken for the design of standard weld-neck flanges:

Determine the diameter, temperature and pressure

Determine the Classified flange dimensions by selecting a standard flange of thecorrect bore using ASME B16.5

22 , API 605

28 , API TR 6AF

24 or equivalent

European standard.

If designing to API TR 6AF 24

, then the effect of thermal gradient has been

included. Otherwise, insulate to counter any severe thermal gradient.

6.3.3 WELD-NECK FLANGE DIMENSIONS AND PRESSURESThe dimensions for standard weld-neck flanges up to 24in diameter are given in

Tables 5 and 8 through 22 of ASME 16.5 22. The dimensions are given in mm apart

from the stud bolts, which are in inches.

API 6A 23

gives standard flange dimensions for higher operating pressures.

A new combined code is being produced which specifies both the US and modified

DIN standards. This is EN 1092 31

and EN 1759 34

.

6.4 SWIVEL RING FLANGES

6.4.1 DESIGN USING SWIVEL RING FLANGESThere are no standards for the dimensions of swivel ring flanges other than the

limited ratings and sizes smaller than 346 mm (135/8in) of API 17D 27. Instead,

these are usually proprietary items but must be designed to match the bolt diameter

and spacing and the gasket size for that of the matching weld neck flange. They are

similar in principle to lap joint flanges and can be designed to ASME VIII 5 or

PD 5500 4. The swivel inner hub profile and full thickness of the swivel outer cangive the flange the same strength and external load capabilities as a weld neck

flange.




The figure to the left of Figure 2 shows how the flange hub is commonly socketed

into a recess in the ring. The codes, however, only provide design methods for the

simple ring-type, lap joint flange shown to the right. The additional steel of thesocketed ring helps stiffen it to resist ‘toroidal’ bending of the flange faces.

Figure 2 Socketed and simple ring type swivel ring flanges

The shape of flanges of API 17D 27

is similar to that on the left, but the tapered

ring/hub interface is at an angle of 25° rather than vertical. If this style of flange

were to be required at a non-standard size or rating, then full finite element analysis

(FEA) would be required.

6.4.2 RECOMMENDED STEPS FOR DESIGNFollowing the simplified flow diagram of Figure 1, the following steps should be

undertaken for the design of swivel ring flanges:

Determine the diameter, temperature, pressure, moments and axial tension

Design the mating weld-neck flange dimensions


4 to pick suitable material and to calculate the

flange stresses. Use could be made of a MathCAD sheet or the Coflexip

programme 6

Increase flange dimensions if stresses exceed the maximum shear stress

(Section 5.1) and repeat analysis

Insulate to counter any severe thermal gradient or carry out a full FEA using the

flange sizes as a basis.




7 DISCUSSION OF DESIGN METHODS

7.1 OVERVIEW

7.1.1 WELD NECK FLANGESPipeline flanges 610 mm (24in) nominal diameter or less for hydrocarbon transport

are commonly sized according to the standard imperial dimensions set out in

ASME 16.5 22

. For each diameter, there are a number of classes to match the

pressure and temperature rating of the pipeline. Originally, the flanges were for

land-lines but the same dimensions can be used with external overpressure for

subsea use.

However, no account is made by this method for the stresses induced by a thermal

gradient.

DIN produced an equivalent set of standard flanges for metric pipe diameters. Both

are now listed in the new EN 1092 31

and EN 1759 34

.

API has produced ranges of flanges for high pressure hydrocarbon and wellhead

use. Additional work has recently been undertaken by API to make allowance for

thermal gradient. This was undertaken using finite element analysis and published

as sets of graphs.

7.1.2 LARGER FLANGE DIAMETERSSome subsea pipelines used for hydrocarbons are larger than the standard diameters

of flange given above. API 605 28

gives dimensions of flanges from 660 mm to

1524 mm (26in to 60in).

Alternatively, two design codes, ASME VIII 5 and PD 5500 4 may be used for non-

standard weld-neck flanges. Both use a similar approach to determine the stresses.

7.1.3 SWIVEL RING FLANGESThere are no standard sizes for swivel ring flanges other than the limited selection

from API 17D 27

. However, for subsea use, it is necessary to be able to marry up the

alignment of the boltholes on either side of the flange connection without imparting

torque into the pipelines.

The approach of ASME VIII 5 and PD 5500

4 can be used for swivel ring flanges.




7.2 SIMPLIFIED STRESS ANALYSIS

Both ASME VIII 5 and PD 5500 4 use an analysis method developed in the 1930s. It

was based on the Taylor Forge method, further developed by Walters et al 9.

Additional information on the method is given by Singh 10

. It simplifies the

problem to a 2 dimensional balance of bolt, pressure and gasket reaction forces

using plain shells.

Bolt load

Gasket reaction

Pressureload

Flangehub

Taperedhub Shell

Figure 3 Simplified axial forces in flange

The flange tends to rotate due to the moment from the combination of axial loads of

the bolts, gasket and pressure. Refer to Figure 3. To this must be added the rotation

due to the internal pressure, which is more pronounced in flanges without a tapered

hub section. Refer to Figure 4.

Figure 4 Radial deflection of shell and flange with internal pressure load

The bolts are deemed to be a point load as if they were spread in a thin continuous

membrane around the flange at the bolt centreline. Other axial forces are deemed to

act at the centreline of the gasket or shell (see Figure 3 for definition). To this is

added the internal pressure during operations.

An initial value of bolt force must be selected to account for makeup. This can then

be checked for overstressing (or slackening) in operational use.

The method can incorporate tensile forces and applied external moments. But no

allowance is made for thermal gradient stresses.

7.2.1 ASME VIII 5 MAXIMUM SHEAR STRESSASME VIII

5 uses the concept of stress intensity, which is defined as twice the

maximum shear stress; which in turn is equal to the algebraic difference between the

maximum and minimum principal stresses:

SI = 2 τmax = σ1 - σ3




The stresses permitted by ASME VIII 5 are given in Tables 1A and 1B of Section II,

Part D Material Requirements Tables for ferrous and non-ferrous alloys. The

permitted operational tensile membrane and yield stresses (Sm and Sy) for eachmaterial are reduced with increasing operating temperature. Values for a full

selection of carbon, low and high alloy steels and nickel alloys are given. The same

section also gives guidance for quenched and tempered steel and clad pipe.

Section 3-350 of ASME VIII 5 provides the maximum allowable flange and nozzle

stresses (Sf and Sn) are compared with the maximum longitudinal, radial and

tangential (SH, SR , and ST) stresses, defined in Section 3-340 (a). Note that for

weld-neck and swivel ring flanges, the flange and nozzle are normally from the

same material; but the thickness or final treatment may cause the allowable stresses

to differ slightly. In other types of welded flange, they may be made from two

different materials.

The code recommends that design stresses be limited to:

SH ≤ the lesser of 1.5 Sf or 2.5 Sn

SR ≤ Sf

ST ≤ Sf

(SH + SR )/2 ≤ Sf

(SH + ST)/2 ≤ Sf

7.2.2 HYDROTEST Note that in the US, it is common to pressure test at 1.25 times the MAOP at the

flange (API RP 1110 11

, ASME B31.4 12

and B31.8 13

) rather than the 1.5 times

specified elsewhere in the world (including the UK under the old BS 8010 14

) and

for some in-company procedures.

The ratio of ST/Sm = 0.83/(2/3) = 1.245. This means that where the test pressure is

the higher value, the flange membrane stress may reach yield.

7.2.3 PD 5500 4 NOMINAL DESIGN STRESSThis British Standard (formerly BS 5500) closely follows the approach of

ASME VIII 5. However, in Tables K.1-5 through K.1-7 of Appendix K, it provides

the minimum tensile, the minimum yield and the nominal design strength values

(R m, R e and f N) for UK grades of steel at a range of temperatures. This also gives

guidance on the maximum design lifetimes for steel running at very hot

temperatures (>350°C or more). Such temperatures are normally outwith the design

condition of subsea pipeline flanges.

The method used by PD 5500 4 for deriving minimum yield/proof stresses for

elevated temperatures is that described in BS 3920 15

, now superseded by

BS EN 10314 16

.

Section 2.3 of the code shows the derivation of nominal design strength. In the

following, variables are:

f E = nominal design strength

R m = minimum tensile strength specified at room temperature

R e = minimum specified yield strength at room temperature

R e(T) = minimum specified yield strength at elevated operating temperature




For carbon, carbon manganese and low alloy steels, the following strengths apply:

Up to 50°C, f E = R e/1.5 or R m/2.35, whichever is the lower value

With a known value of elevated temperature R e(T), for 150°C and above,f E = R e(T)/1.5 or R m/2.35, whichever is the lower value

Without a known value of elevated temperature R e(T), for 150°C and above,

f E = R e(T)/1.6 or R m/2.35, whichever is the lower value

Between 50°C and 150°C, based on linear interpolation of the above

For austenitic stainless steels, the following strengths apply:

Up to 50°C, f E = R e/1.5 or R m/2.5, whichever is the lower value

With a known value of elevated temperature R e(T), for 150°C and above,


Without a known value of elevated temperature R e(T), for 150°C and above,


Between 50°C and 150°C, based on linear interpolation of the above

7.3 THERMAL GRADIENT ANALYSIS

API TR 6AF1 25

and 6AF2 26

consider thermal forces due to the bending induced in

flanges when the product is hot and the exterior is cold.

They assessed the full range of standard flanges for leakage due to lack of pressure

on the gasket or failure of the flange or bolt. In addition to thermal gradient,

temperature derating combined with internal pressure and moments were applied.

Tables were derived for the full range from finite element studies.

Figure 5 Finite element mesh used by API TR 6AF 24

The FE mesh in Figure 5 shows a simple flange with no taper to the hub. The

element size is reduced at the stress concentration point between the flange and the

shell. It seems to be common to model the pipeline thickness in two or three

elements. It is important to model a quadrant of flange in 3D because of the effect

of the applied moment and to examine bending in the space between the boltholes.

At the holes themselves, the elements are adjusted to form a ring around the bolts.




However, in this model, the gasket has not been examined in detail. Compare this

with the FE mesh shown in Figure 7, which models the groove and gasket in detail.

Figure 6 shows a typical thermal gradient (in °F) and the resulting stresses in the

flange obtained from the analysis. If the outside of the flange is insulated, the

thermal gradient is more even (the outside is almost as warm as the inside) and there

is little additional stress.

Thus, it may be possible to ignore the effects of thermal gradient stresses if the

flange is well insulated.

However, the bolts need to be de-rated to take account of the hotter operating

temperature.

Figure 6 Thermal gradient (in °F) and flange stress plots

7.4 LOW TEMPERATURE FLANGES

Most concern is given to de-rating of material stresses for high temperatures due to

the product. However, strength can also reduce with extreme low temperatures.

A guidance document 17 is given for flanges made from ASTM A105 material

subjected to low temperatures down to -29°C. Such flanges may be at risk from

brittle failure. The document is valid for standard weld neck flanges conforming to

dimensions given in ASME B16.5 22

and also non-standard flanges with a nominal

thickness range of 32 mm to 102 mm.

7.5 AGGRESSIVE ENVIRONMENTS INCLUDING SOUR SERVICE

Where the aggressive nature of the product demands additional corrosion resistance

API recommends design to NACE MR 01 7518

and the use of type 6B flanges.

It is possible to replace the carbon steel flanges with CRA material or to use cladsteel. The latter should be returned up around on the inside of the groove for the

gasket. ASME VIII 5 gives guidance on clad pipe in Section AM-220.




7.6 FINITE ELEMENT ANALYSIS

A proprietary FEA service offered by Welding Units (www.welding-units.co.uk) for

flange design does model the gasket and adjusts the mesh locally in the hub to a

smaller size around the groove. Refer to Figure 7.

Figure 7 FE mesh used by Welding Units

Typical cell sizes are shown in Figure 5 and Figure 7. It is necessary to analyse a

quarter of the flange in 3D to allow for externally applied moments as shown in

Figure 5. For extreme conditions, it is usual to assume the worst position of the bolt

aligned on the quarter axis.

All conditions need assessing including the stress for thermal gradient (see examplein Figure 6), which causes toroidal bending in the flange.

Checks should be made for bolt, gasket, flange membrane stress and body stress

under all load combinations. It is possible to include plasticity in design for

secondary stresses.

7.7 COFLEXIP PROGRAMME

Some guidance on materials is provided in Appendix C. Information is provided for

stainless steel flanges because these are commonly used with flexible pipelines and

risers. However, no values are given for temperature de-rating. Also, the values

given are for material yield strengths rather than the correct figures for de-ratedshear stress intensity required by ASME VIII

5 and PD 5500

4.




The programme simplifies the design of swivel ring flanges. It calculates the ring

stresses following Section 3-340 (b) rather than (a). That is: it assumes the

longitudinal and radial stresses in the ring to be zero. This then means that thetangential stress has no reduction from the radial stress moment contribution. It also

assumes a simple ring rather than the socketed ring shown in Figure 2.

There is a minor discrepancy in the manual for one value used to determine the

value for Y. The ASME VIII 5 code on page 248 Figure 3-340.1 uses a constant of

5.71690 whereas the manual quotes 5.717690. Since these are very close, it is not

possible to determine if this is a typographic error, or if the programme itself uses

the wrong value.

7.8 TECHNIP MATHCAD SHEETS

7.8.1 INPUT DESIGN STRESSESIt is recommended that design stresses be input from the tables rather than scaled

from the ultimate stresses. These need to be de-rated for operating temperature.

7.8.2 EXTERNALLY APPLIED MOMENTBecause the ASME VIII

5 and PD 5500

4 methods both simplify the bolts to a single

point load, there is no need to find the moment of inertia of the bolt plane.

7.8.3 DERIVATION OF HUB CONSTANTSTwo of these sheets use the formulae from ASME VIII

5 to derive values for F, V

and f. In PD 5500 4, the formulae are not provided, necessitating the use of the

graphs. However, the graphs in the two codes are identical and are derived from the

same underlying theory.

For the ‘ASME VIII RTJ Flange_Master.mcd’ a hidden section on page 5 is used to

derive values of F, V and f. However, values for F and V then seem to have been

misread from the graphs and re-defined as new inputs. Also, the value for the

longitudinal stress due to the applied bending moment (SH1) has been omitted from

page 6.

The note on page 252 of ASME VIII 5 Division 2 Table 3-340.1 referring to integral

flanges states that certain values shall be given to F, V and f if there is no taper to

the nozzle (that is, if g1 = go). However, this hidden section would derive incorrect

values for F and V in this condition.

In the ‘12in PD 5500 sheet.mcd’ 1 sheet, despite the values of h/ho and g1/go being

correct, it would appear that the value for V on sheet 7 of 8 has been misread from

the graph. It is a factor of 10 too high. Subsequent variables are then incorrect.

7.8.4 COMPARISON OF CALCULATED WITH PERMISSIBLE VALUESOn the final line of ‘ASME VIII RTJ Flange_Master.mcd’ , the sheet derives stress

limits which are both more than and less than 1.0. It is recommended that this line

be an automatic check rather than an input.




8 FLANGE GRADES, BOLTS AND GASKETS

8.1 FLANGE MATERIAL GRADES

8.1.1 APIAPI 6A

23 classifies suitable materials in Tables 5.1 and 5.2 by pressure. For

flanges, only three grades are given, reproduced here in Table 2. They do not

correspond to named steels, merely grades.

API materialdesignation

0.2% Minimumyield strength

Minimumtensile strength

45K 248 MPa (45 ksi) 483 MPa (70 ksi)

60K 414 MPa (60 ksi) 586 MPa (85 ksi)

75K 517 MPa (75 ksi) 655 MPa (95 ksi)

Table 2 API material properties for flanges

The allowable steel grade for each API pressure rating is reproduced in Table 3.

13.8 20.7 34.5 69.0 103.4 138.0Pressurerating

MPa (kip) (2) (3) (5) (10) (15) (20)

Material 45K 45K 45K 60K, 75K 75K 75K

Table 3 API material applications for weld neck flanges

8.1.2 US STEELSTables 1A and 1B of Section II, Part D of ASME VIII 5 gives specified minimum

yield and tensile along with appropriate design strength intensity values, Sm for

elevated temperatures to 482°C (900°F) or more. The list includes carbon and low

alloy steels, high alloy steels, quenched and tempered steels, nickel and nickel

alloys.

PD 5500 4 enquiry case 5500/91 of September 2003 permits the use of ASTM and

API materials. The response tabulates de-rated temperature values in metric units.




8.1.3 UK AND EUROPEAN STEELSTable K.1-6 of PD 5500

4 gives a ppropriate design strengths for steel forgings

following the method of BS 1503 19

. These include values for carbon manganesesteels and alloy, martensitic and austenitic stainless steels at elevated temperatures

to 480°C or more. However, BS 1503 19 has now been replaced by BS EN 1022220.

8.2 METRIC AND IMPERIAL STUD BOLT SIZES

It is common to find imperial flanges but with metric sizes of bolts. The

preliminary standard prEN 1759-1 34

Annex F gives guidance as to the equivalents

up to M39.

Imperialdiameter

(inch)

Pitch(threads

per inch)

Metricequivalent

(mm)

Pitch(mm)

1/2 13 UNC M14 2.0

5/8 11 UNC M18 2.5

3/4 10 UNC M20 2.5

7/8 9 UNC M24 3.0

1 8 UNC M27 3.0

11/8 8 UNC M30 3.5

11/4 8 UNC M33 3.5

13/8 8 UNC M36 4.0

11/2 8 UNC M39 4.0

15/8 8 UNC M42 4.5

13/4 8 UNC M45 4.5

17/8 8 UNC M48 5.0

2 8 UNC M52 5.0

21/4 8 UNC M56 5.5

21/2 8 UNC M64 6.0

23/4 8 UNC M72 6.0

3 8 UNC M76 6.0

Table 4 Suggested metric bolt sizes

These are similar to the Technip recommendations given in Appendix C of the

Coflexip work procedure 6, and reproduced here in Table 4. The only difference

between the two is for5/8in bolts, which the former recommends as M16 and

Coflexip recommend as M18.

Sizes smaller than M14 (½in) should not be used. Occasionally larger sizes than

listed – such as M90 (3½in) – may need to be used but these are hard to handle by

divers.

8.3 BOLT TENSIONS

Bolt selection is an important consideration in the design of flanges. Although not

part of the present scope, some information is supplied here as guidance.




8.3.1 ACCEPTABLE LOADSBolts are generally tensioned to approximately 50% of yield. See discussion of bolt

tensions in API TR 6AF2 21

.

During operation, this is allowed to increase to 83% or reduce to 37.5%, depending

upon the combination of loads. With increased temperature and internal pressure

alone, the load in bolts tends to decrease. Eventually, the gasket will cease to seal.

However, for subsea flanges, it is not common to further adjust the bolt tensions

after the make-up operation, so the initial tension must maintain the seal during the

subsequent hydrotesting and operation load combinations. Bolt re-adjustment

occasionally does take place with landline flanges, in particular with large

diameters.

8.3.2 TORQUEIt is normal to specify a make-up tension for the bolts. The contractor can then

pretension to this level prior to closing down the nuts.

Alternatively, a torque wrench can be used to achieve this amount of pretension.

However, the amount of torque that is applied is heavily dependent upon the

lubrication applied to the nuts, washers and stud bolts.

A procedure of tightening bolts generally follows a four stage make-up:

The bolts are brought to 30% of their required torque following a pattern (12, 6,

3, 9; 1, 7, 4, 10; 2, 8, 5 and 11 o’clock for a twelve bolt flange)

The same pattern brings the bolts to 60% of the specified torque

The full torque is applied following the same orderFinally, full torque is repeated.

It should be noted that most codes have been written assuming that bolts will be

torqued up using a wrench, or in some cases, simply using a spanner (with no check

on their torque).

Nevertheless, PD 5500 4 recognises that where controlled tensioning of bolts is

undertaken (the normal subsea practice), then the values in the table may be

increased by 20%. During testing, the design stress may be multiplied by 150%.

8.3.3 BOLT MATERIALSPermitted stresses in bolting materials are given in ASME VIII 5 Table 3 Section II

Part D. These provide values for low and high alloy steels and nickel alloys,

respectively. The allowable bolt stress is reduced depending upon the temperature

and diameter of the bolt. Note that the temperature is not necessarily the same as

that of the flange or product.

PD 5500 4 (in Table 3.8-1) provides recommended design stresses for UK grades of

steel bolts over a range of temperatures. The root areas for common sizes of both

metric and imperial bolts are given in Table 3.8-2.

Commonly, uncoated ASTM Type B7M or L7M studs are used with 2HM nuts.

8.3.4 CATHODIC PROTECTION

It is possible to coat the bolts with PTFE but care should be taken to ensure theymake electrical contact for cathodic protection.




Bolt yield strengths greater than around 700 MPa may suffer from hydrogen

embrittlement. For higher strength bolts, Type L7 or B7 with 2H nuts can be used.

8.4 GASKETS

8.4.1 MAKEUP AND OPERATIONAL STRESSCare needs to be taken when loading the bolts to ensure that there is sufficient

tension to guarantee no leakage after makeup and yet prevent over crushing of the

gasket during operation. Refer to Figure 8 for the following discussion.

Figure 8 Simplified comparison of gasket deflection and stress

When the bolts are first tightened up, the gasket follows a non-recoverable path

shown above between points 1 and 2. During this stage, the gasket is deformed to

match the flange faces to make a seal. The point at which the gasket provides the

minimum effective seal is known as the gasket seating stress.

The region marked 2-4 is the useful sealing range of the gasket. The value of the

seating stress can be found in the associated codes for the particular gasket.

However, the crushing limit must usually be obtained from gasket manufacturers

because it is not specified in the codes. When the gasket is compressed beyond its

crushing limit, some form of breakdown usually occurs, varying according to the

type of gasket.

If the gasket is tightened to some value between points 2 and 4 and then the gasket

is unloaded (by internal pressure or bolt loosening) then reloaded, it will follow a

path such as points 3-5-3, rejoining the original curve.

Crushin L imit

Make-u Stress

Seatin Stress

GasketStress

Gasket Deflection1

2

3 4

5




8.4.2 GASKET MATERIALFor subsea flanges, there are three classes of gasket material listed in ASME VIII

5,

Table 3-320.1. Gaskets are selected to be softer and more flexible than the materialof the flanges in order that they seat properly, ensuring a good seal.

The suggested design values for m and y given in the code and reproduced below in

Table 5 are described as ‘generally proving satisfactory in service’. A similar table

is given in PD 5500 4 (Table 3.8-4).

Ring jointmaterial

Gasketfactor, m

Minimum designseating stress, y

Iron or

soft steel

5.50 124 MPa

(18.0 ksi)

Monel or4% to 6% chrome

6.00 150 MPa(21.8 psi)

Stainless steels 6.50 180 MPa

(26.0 ksi)

Table 5 Recommended gasket factors, m and y

Design code API 6A 23

provides guidance on the Rockwell hardness for different

gasket materials. This is reproduced in Table 6.

Ring gasket material Maximum hardness

Soft iron HRB 56Carbon and low alloy steel HRB 68

Stainless steel HRB 83

Corrosion resistant alloys To manufacturer’s specification

Table 6 Recommended gasket hardness

For land-based facilities (where it is possible to re-seat flanges during maintenance

operations), it is common to use other gasket materials and shapes. For offshore

use, the approach normally demands a metal-to-metal seal and an octagonal or oval

ring joint fitted into a groove in the face of the flange.

In the event of an external fire, metal gaskets are able to resist temperatures similarto that of the flange material. Leakage of hydrocarbon product due to gasket failure

would result in the line contents feeding the fire. Use of non-flammable metal

gaskets helps prevent this.

8.4.3 API GASKET SHAPESFour cross-sections of gasket are permitted with appropriately shaped grooves.

These are the BX, RX and R (octagonal and oval) from API 6A 23

, which are

reproduced in the figures below. API 17D27

has similar shaped grooves and gasket

types.

Even when non-standard flanges are to be designed, it is recommended that the

groove dimensions and tolerances should follow these standards.




OD

C

H

A

R R

R R

ODT

23°

D

Type BXGasket

1/16"

45° max

/32 R

N

G

E

23° 23°

Breaksharpcorner

Type BXGroove

Figure 9 API Type BX pressure-energised gasket and grooveAPI 6A

23 provides standard dimensions for Type BX gaskets, which are given up to

a nominal size of 726 mm corresponding to a ring diameter of 852.75 mm. In the

above figure, tolerances are as listed in Table 9:

Dimension Descrip tion Tolerances (mm)

A* Width of ring +0.20, -0

C Width of flat +0.15, -0

D Hole size ±0.5

E Depth of groove +0.5, -0

G OD of groove +0.10, -0

H* Height of ring +0.20, -0

N Width of groove +0.10, -0

OD Outside diameter of ring +0.0, -0.15

ODT Outside diameter of flat ±0.05

R Radius of ring See note

23° Angle ±¼°

Table 7 Type BX tolerances

* A plus tolerance of 0.20 mm for width A and height H is permitted, provided thevariation in width or height of any ring does not exceed 0.10 mm throughout its

entire circumference.

Note: Radius R shall be 8% to 12% of the gasket height H.

One pressure-passage hole is required per gasket on the centreline.

The API Type SBX pressure-energised ring gaskets from API 17D Table 906.1 are

of a similar profile as type BX but with slight differences in tolerances. Sizes up to

640 mm OD are specified. They also have a pair of pressure holes linking to the

inside of the flange to prevent pressure lock when the connection is made

underwater. Raised flange faces are assumed to touch with the Type BX gasketonly.




OD

C

H

A

23°

R1D

23°

Type RXGasket

R2

F

P

E

23°

Type RXGroove

Figure 10 API Type RX pressure-energised gasket and groove

API 6A 23

presents standard sizes for Type RX gaskets up to an outside diameter of

ring of 600.87 mm. In the above figure, tolerances are as listed in Table 8:


A* Width of ring +0.20, -0

C Width of flat +0.15, -0

D Height of chamfer +0, -0.8


F Width of groove ±0.20

H* Height of ring +0.2, -0

OD Outside diameter of ring +0.5, -0

P Average pitch diameter of groove ±0.13

R 1 Radius in ring ±0.5

R 2 Radius in groove max

23° Angle ±½°

Table 8 Type RX tolerances

* A plus tolerance of 0.20 mm for width A and height H is permitted, provided the

variation in width or height of any ring does not exceed 0.10 mm throughout its

entire circumference.

Note: The pressure passage hole illustrated in the RX ring cross-section is for rings

RX-82 through RX-91 only. The centreline of the hole shall be located at the mid-

point of dimension C. The hole diameter shall be 1.5 mm for rings RX-82 through

RX-85, rings RX-86 and RX-87; and 3.0 mm for rings RX-88 through RX-91.

The API Type SRX pressure-energised ring gaskets from API 17D Table 906.2 are

of a similar profile as type RX but with slight differences in tolerances. Sizes up to

140 mm OD are specified. They also have a pair of pressure holes linking to theinside of the flange to prevent pressure lock when the connection is made

underwater.




P

A

B

Type ROval

E R2

F

P 23°

Type RGroove

P

C

H

A

23°

R1

Type ROctagonal

23°

Figure 11 API Type R gaskets and grooveAPI 6A

23 presents standard sizes for Type R gaskets up to a pitch diameter of

groove and ring of 584.20 mm. In the above figures, tolerances are as listed in

Table 9:


A Width of ring ±0.20

B Height of oval ring ±0.5

C Width of flat on octagonal ring ±0.2


F Width of groove ±0.20

H Height of octagonal ring ±0.5

Average pitch diameter of ring ±0.18P

Average pitch diameter of groove ±0.13

R 1 Radius in ring ±0.5

R 2 Radius in groove max

23° Angle ±½°

Table 9 Type R tolerances

For pipeline diameters requiring gaskets larger than given in the tables, theappropriate tolerances need to be adopted.

8.4.4 SELF-ENERGISING GASKETSGaskets Type BX and RX are deemed to be self-energising. Note that for these, the

outside diameter of the gasket is specified. Compare this with Type R octagonal or

oval gaskets where the mean diameter is specified.

For self-energising gaskets, the pressure must be considered to act on the face of the

flange out to the outside of the gasket groove.

8.4.5 PRE-TESTABLE GASKETSAlthough it is possible to adjust the bolts at the hydrostatic leak test stage, it is

normal to tension them up only once during assembly, to a level that will need nofurther adjustment.




Figure 12 KaMOS® gasket in groove

Gaskets are available that can prove the leak tightness of the joint at assembly stage.

One such proprietary product is the patented KaMOS® system shown in Figure 12.

Details of this gasket are available at http://www.karmsund.no/eng_index.htm.

bly stage.

One such proprietary product is the patented KaMOS

® system shown in Figure 12.

Details of this gasket are available at http://www.karmsund.no/eng_index.htm.

These seal tests work by pressurising the gaps at the foot of the gasket grooves

through holes in the octagonal gaskets, as shown in Figure 13. Any lack of seal

between faces of the grooves and the gasket will then indicate adjustment is needed.

These seal tests work by pressurising the gaps at the foot of the gasket grooves

through holes in the octagonal gaskets, as shown in Figure 13. Any lack of seal

between faces of the grooves and the gasket will then indicate adjustment is needed.

Seals

Figure 13 KaMOS® testing of sealFigure 13 KaMOS

® testing of seal

Other similar systems are available, such as from Grayloc, Vector and Oilstates. It

should be noted, however, that not all clients accept this approach to testing.

Other similar systems are available, such as from Grayloc, Vector and Oilstates. It

should be noted, however, that not all clients accept this approach to testing.




9 RELEVANT READING

This section lists the main sources of data regarding flanges and their design, along

with a brief discussion of their content and relevance.

ASME B16.5 22

– this publication provides a simplified method for determining the

class of flange needed. It considers pressures and temperature derating for ten

different flange material groups. No allowance is made for bending moments,

tension or temperature. Dimensions for standard flanges up to 24in are tabulated.

ASME VIII 5 – this is the main source of design for American flanges. The whole

code covers the design of pressure vessels. Flange design is covered in Division 2.

PD 5500 4

– this is the main UK source for the design of flanges. Again, like the

ASME VIII 5 code, it covers the design of pressure vessels. The method for the

design of flanges follows a similar underlying code to that of ASME VIII 5

Division 2. Section 3.8 is specific to bolted flanged connections and sixteen

standard working forms are provided to help with flange designs of various types.

Weld-neck flanges are Form 1 and swivel ring flanges can be determined by

adapting Form 5.

API 6A 23

this provides permissible stresses and standard sizes for flanges and

gaskets based on ASME VIII 5 methods, specifically for wellhead equipment.

Pressures classes considered in API 6A 23

are 2, 3, 5, 10, 15 and 20 ksi with nominal

sizes from 71/16in to 21¼in, though the larger diameters do not cover the maximum

pressure classes.

API TR 6AF 24

– this was developed from earlier work and provided an improved

method of flange design as compared with traditional, more conservative methods,

such as designing for pressure loads only. This covered the determination of rating

charts for API 6B and API 6BX flanges but did not consider the effects of

temperatures greater than 121°C.




The effect of temperature on load capacity affects:

the yield strength (and thus the allowable stresses) of the flange and bolt

materials

thermal stresses caused by thermal gradients across the flange and between the

bolts and the flange. This can cause bending of the flange faces.

API TR 6AF1 25 – this study has examined the effect of elevated temperatures on

inside bore of standard weld-neck flanges with the outside at 0°C and subject to

wind-induced forced cooling. The elevated temperatures are 176.6°C and 343.3°C.

The analyses were undertaken using simple ANSYS axisymetric models and four

different materials combinations.

API TR 6AF2 26

– this study extends the work of API TR 6AF1 25

. It considers the

effect of applied moments and tension combined with the temperature profilethrough the flange. All seventy of the API 6A

23 flanges were modelled as a 3D

finite element mesh.

For this work, the SESAM FE package was used. These flange and bolt material

combinations are:

Carbon and low alloy steel flanges with two types of bolt

Martensitic, ferritic and precipitation hardening stainless steel flanges

Austenitic and duplex stainless steel flanges

Suitable derating factors are derived. Graphs are presented for each flange diameter

and material combination to enable the limiting loads to be determined.

It should be noted that gasket leakage tended to be the governing case for the

smaller flanges and for higher pressure operations. This may be due to conservative

assumptions made in the analysis for gaskets. Stresses in flanges tended to have

little reserve beyond the gasket leakage level. The bolt stresses did not govern for

any case. They were typically within ⅔ of their yield strength due to make-up,

pressure, tension and bending moment loads when they are made-up to around half

their yield.

With insulated flanges, the temperature profile is more constant across the

thickness. This results in a higher capacity of the flanges to resist the other loads.

API 17D 27 this provides sizes for weld neck and swivel ring flanges suitable for

subsea wellhead equipment. The present study excludes these designs. Tapered

swivel ring 346 mm (135/8in) flanges at Class 5000 and 10 000 are specified for

wellheads.

API 605 28

this provides sizes for standard flanges between 26in and 60in.

ISO 7005-1 29

– the ISO code covering both DIN and ASME flanges.

BS 1560-3.1 30

– This gives the dimensions of steel flanges for the petr oleum

industry. Standard weld neck flanges are specified. It equates to ANSI B16.5 22

.




BS EN 1092-1 31

– this is based on ISO 7005 29

but only specifies dimensions of

DIN type flanges at temperatures and pressures from PN 2.5 to PN 100 and

diameters DN 10 to DN 4000. It includes a range of different types of flangesincluding weld-neck, which it designates as Type 11. Note that these flanges have

slightly changed from the older DIN pressure rating. This code also considers

swivel ring flanges (combination Type 34 short weld neck with Type 03 loose plate

flange).

BS EN 1591-1 32

and BS EN 1591-2 33

these provide a detailed design method for

flange and gasket selection. Blank and weld-neck flanges are included. Any size of

flange can be designed using this code.

prEN 1759-1 34

– the companion to EN 1092-1 31

covering ASME flanges. The

prefix pr indicates it is not fully released.






15 BS 3920 : 1973 Derivation and verification of elevated temperature properties

for steel products for pressure purposes – superseded by BS EN 10314

16 BS EN 10314 : 2002 Method for the derivation of minimum values of proof

strength of steel at elevated temperatures

17 Un-numbered document by the Belgian Chemical Risks Directorate, Technical

Inspectorate, Administration of Labour Safety, Federal Ministry of Employment

and Labour, July 2002 – Recommendations for pipe flanges made in forged steel

complying with ASTM A 105

18 NACE MR 01 75 Part 1 Rev 1 2001 (ISO 15156-1) – Petroleum and natural gas

industries – Materials for use in H2S-containing environments in oil and gas

production – General principles for selection of cracking-resistant materials

19 BS 1503 : 1989 Specification for steel forgings for pressure purposes – now

superseded by BS EN 10222

20 BS EN 10222 Steel forgings for pressure purposes Part 1 : 1998 generalrequirements for open die forgings; Part 2 : 2000 ferritic and martensitic steels

with specified elevated temperature properties; Part -3 : 1999 Nickel steels with

specified low temperature properties; Part 4 : 1999 Weldable fine-grain steels

with high proof strength; and Part 5 : 2000 Martensitic, austenitic and austenitic-

ferritic stainless steels

21 API TR 6AF2 Second edition : 1999 by the American Petroleum Institute –

Technical Report on Capabilities of API Flanges under Combination of Loading

– Phase II

22 ASME 16.5 : 2003 by The American Society of Mechanical Engineers – Pipe

Flanges and Flanged Fittings NPS ½ through NPS 24 Metric/Inch Standard

(Formerly published by the American National Standards Institute as

ANSI 16.5.)23 API 6A nineteenth edition, July 2004, by the American Petroleum Institute –

Specification for wellhead and christmas tree equipment

24 API Bulletin 6AF First edition : 1989 by the American Petroleum Institute –

Capabilities of API Flanges under Combination of Load


Technical Report on Temperature derating on API Flanges under Combinations

of Loading


Technical Report on Capabilities of API Flanges under Combination of Loading

– Phase II

27 API 17D (also known as ISO 13628-6) with two supplements to June 1996, by

the American Petroleum Institute – Specification for subsea wellhead andchristmas tree equipment

28 API 605 revision March 88, by the American Petroleum Institute – Large

diameter carbon steel flanges (nominal pipe sizes 26in through 60in, Classes 75,

150, 300, 400, 600, and 900)

29 ISO 7005 Metallic flanges

30 BS 1560-3.1 : 1989 Circular flanges for pipes valves and fittings (Class

designated) – Part 3: Steel, cast iron and copper alloy flanges – Section 3.2

Specification for steel flanges

31 BS EN 1092-1 : 2001 Flanges and their joints – circular flanges for pipes, valves,

fittings and accessories, PN designated – Part 1 Steel flanges – incorporating

Corrigendum N° 1

32 BS EN 1591-1 : 2001 Flanges and their joints – design rules for gasketed circular

flange connections – Part 1 Calculation method




33 BS EN 1591-2 : 2001 Flanges and their joints – design rules for gasketed circular

flange connections – Part 2 Gasket parameters

34 prEN 1759 : 2000 – Flanges and their joints – circular flanges for pipes, valves,

fittings and accessories, Class designated

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