Bolted Flanged Joints New Methods & Practices - Proceedings.pdf

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16-17 March 2010 Turbine Hall The CastleGate Melbourne Street Newcastle upon Tyne NE1 2JQ Seminar Proceedings Pressure Systems Group BOLTED FLANGED JOINTS: NEW METHODS AND PRACTICES Improving the world through engineering Photo courtesy of Hydratight

Transcript of Bolted Flanged Joints New Methods & Practices - Proceedings.pdf

16-17 March 2010 Turbine HallThe CastleGateMelbourne StreetNewcastle upon TyneNE1 2JQ

Seminar Proceedings

Pressure Systems Group

BoltedFlanGed JointS:new MethodSand PracticeS

improving the world through engineering

Photo courtesy of Hydratight

© 2010 The Institution of Mechanical Engineers, unless otherwise stated. The copyright in these papers is the property of the Institution of Mechanical Engineers unless otherwise indicated. Apart from any fair dealing for the purpose of private study, research, criticism or review, as permitted under the Copyright, Designs and Patent Act 1988, no part may be reproduced in any form or by any means without permission. Enquiries should be addressed to: Director of Publication and Information Services, The Institution of Mechanical Engineers, 1 Birdcage Walk, London SW1H 9JJ, telephone: 020 7222 7899. The Institution is not responsible for any statement contained in these papers. Data, discussion and conclusions developed by the authors are for information only and are not intended for use without independent substantiating investigation on the part of potential users. Opinions expressed are those of the author and not necessarily those of the Institution of Mechanical Engineers.

Registered Charity Number 206882

CONTENTS 1. Setting the Scene

Robert Noble, Hydratight © Robert Noble 2. Flange Selection

Dr David Nash, University of Strathclyde Simon Earland, Earland Engineering

© Dr David Nash & Simon Earland 3. Gasket Selection

Dr Gavin Smith, Novus Sealing Limited © Dr Gavin Smith, Novus Sealing Limited 4. Material Selection for Industrial Fasteners

Rod Corbett, James Walker Rotabolt © Rod Corbett 5. Traditional Flange Design Methods

Warren Brown, The Equity Engineering Group © Warren Brown 6. Overview of Developments in EN 1591

Manfred Schaaf, AMTEC Services GmbH © Manfred Schaaf 7. Failure Mechanisms of Bolted Joints

- Bolting Aspects Bill Eccles, Bolt Science Limited

© Bolt Science Limited

Seal failure from a gaskets perspective Dene Halkyard, Flexitallic

© Flexitallic Ltd 8. European Emissions Legislation

Dr Brian Ellis, European Sealing Association © Dr Brian S. Ellis

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9. Tension Control, the key to Bolted Flange Reliability

Rod Corbett, James Walker Rotabolt © Rod Corbett Tightening Techniques for Bolted Flanged Joints Tony Scrivens, Hydratight

No content in proceedings 10. Management of the Integrity of Bolted Joints for

Pressurised Systems Robert Noble, Hydratight

© Robert Noble 11. ASME PCC-1 Updates

Warren Brown, The Equity Engineering Group © Warren Brown 12. Qualification of Personnel Competency –

DD CEN/TS 1591-4 John Hoyes, Flexitallic Ltd © J. R. Hoyes of Flexitallic

13. A regulatory perspective on bolted joints at high hazard

sites Iain Paterson, HSE Offshore Division

© Health & Safety Executive 14. Leak Management

Ed Versluis, James Walker Rotabolt © James Walker 15. Case studies

No content in proceedings

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Setting the Scene

Robert Noble, Hydratight © Robert Noble

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Setting the Scene

Robert Noble Technical Services Leader Hydratight

The World is beginning to realise the bolted joint is just as critical as the Welded Joint?

Welded Joint

Coded Welder

Material Control

Documented Procedure

NDT Verification

Hydro-tested

Competent Personnel

Documented Procedure

Hydro-tested

Material Control

Integritytested

Bolted Joint

In Service InspectionRecords Records

Permanent joint Subject to Breakout

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Is the bolted joint a permanent or temporary Joint?

Consider this please:-Consider this question:-

PED applies to permanent joining with permanent joints defined in Article 1 as:

“2.8. 'Permanent joints` means joints which cannot be disconnected except by destructive methods”

The Bolted Flanged joint being capable of disconnection therefore is viewed as temporary! This is an advantage – not a reason for reduced standards of management and control.

The Bolted Joint and the PED:-

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The Permanent Joint and the PED:-3.1.2. Permanent joining• Permanent joints and adjacent zones must be free of any surface orinternal defects detrimental to the safety of the equipment.• The properties of permanent joints must meet the minimum propertiesspecified for the materials to be joined unless other relevant propertyvalues are specifically taken into account in the design calculations.• For pressure equipment, permanent joining of components which contributeto the pressure resistance of equipment and components which aredirectly attached to them must be carried out by suitably qualifiedpersonnel according to suitable operating procedures.• For pressure equipment in categories II, III and IV, operating proceduresand personnel must be approved by a competent third party which, atthe manufacturer's discretion, may be:— a notified body,— a third-party organization recognized by a Member State• To carry out these approvals the third party must perform examinationsand tests as set out in the appropriate harmonized standards or equivalentexaminations and tests or must have them performed.

EQMS no:5144-AC

Standards and concern around Bolted Joints are becoming more prevalent

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• Improved Training and Competence• Improved Design Codes• Improve guidance on determining correct

bolt load.• A trend towards increased bolt load.• Focus on Gasket performance.• Inspection of Bolted Joints.• Improved Management of Bolted Joints

Trends in industry and standards:-

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Flange Selection

Dr David Nash, University of Strathclyde & Simon Earland, Earland Engineering

© Dr David Nash & Simon Earland

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Flange Selection Simon Earland, Earland Engineering Ltd & David Nash, University of Strathclyde INTRODUCTION This paper covers the important features of the main types of flange and indicates some typical uses. Flanges are used for a variety of applications in pressure systems, including piping, valves, nozzles and access openings on vessels and other equipment, and girth flanges on vessels and heat exchangers. Many of these flanges will be standard, “off the shelf” items; others will be custom designed for a specific application. Normally, flanges are specified on the basis of a pressure requirement. Thereafter, other loadings and deflection or leakage requirements, or even welding, installation or access requirements may drive the rationale for flange selection. The intention of this paper is to present an overview of bolted flange types, including both standard and specialist flange designs. STANDARD FLANGES The most common type of flange used for pressure equipment is the standard piping flange. These are supplied in accordance with various national and international standards such as:

• EN 1092 – Flanges and their joints – Circular flanges for pipes, valves, fittings and accessories, PN designated

• EN 1759 – Flanges and their joints – Circular flanges for pipes, valves, fittings and accessories, class designated

• ASME B16.5 – Pipe Flanges and Flanged Fittings: NPS ½ through NPS 24 Metric/Inch Standard

• ASME B16.47 – Large Diameter Steel Flanges: NPS 26 through NPS 60 Metric/Inch Standard

• EN ISO 10423 (ANSI/API Specification 6A) - Petroleum and natural gas industries. Drilling and production equipment. Wellhead and Christmas tree equipment

Flanges to ASME B16.5 are often referred to as “ANSI” flanges because the standard was originally published by ANSI (American National Standards Institute), but it is now published by ASME (American Society of Mechanical Engineers). The European standard EN 1759 is based on the ANSI/ASME standard B16.5, and EN 1092 is based on DIN standard flanges. Flanges are selected according to their nominal size, DN for metric or NPS for inch sizes (also referred to as NB), and their pressure - temperature rating. The main advantages of these standard flanges are:

• Readily available from a range of manufacturers • Design calculations are not normally required • Pressure ratings recognised by the main piping and pressure vessel design codes • Standard dimensions • Wide range of gaskets available in standard sizes

Disadvantages:

• They tend to be overly large and heavy compared with modern designs • Some problems with high seating stress gaskets and low pressure rating flanges

There are two main systems for flange rating, the American system of class designated flanges given in ASME B16.5 and EN 1759, and the European system of PN designated flanges given in EN 1092 Parts 1, 2, 3 and 4. In the oil, gas and petro-chemical industries class designated flanges are generally specified. The same is true in other industries, such as chemicals and pharmaceuticals where the plant is operated by an American based company. For plants operated by European based companies PN designated flanges are often specified.

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Class designated flanges ASME B16.5 and EN 1759 cover sizes from NPS ½ to NPS 24, and flanges are specified by the designations Class 150, Class 300, Class 400, Class 600, Class 900, Class 1500 and Class 2500. These flanges are often referred to as 150 lb (or 150#), 300 lb, etc. The class designations of these flanges correspond to the pressure ratings in psig at elevated temperature, typically 567°F (297°C) for class 150 and 860°F (460°C) for class 300 and above for carbon steel A-105 material. The pressure ratings in psig at ambient temperature are much higher than the class designation. For example, the pressure rating of a Class 150 flange in ASTM A-105 material at ambient temperature is 285 psi, and the rating of a Class 300 flange is 740 psi. The maximum working pressures are tabulated against temperature, and tables are provided for various groups of materials. Standard flange dimensions are also tabulated. In ASME B16.5 tables are provided in both metric units (bars and mm) and US customary units (psig and inches). ASME B16.5 covers a wide range of carbon and alloy steels, stainless steels and nickel alloys. In EN 1759 Part 1 covers steel flanges, Part 3 covers copper alloy flanges and Part 4 covers aluminium alloy flanges. ASME B16.47 covers sizes from NPS 26 to NPS 60, and flanges are specified by the designations Class 75, Class 150, Class 300, Class 400, Class 600 and Class 900. This standard covers two series of flanges – Series A, which were previously known as MSS SP-44; and Series B, which were previously know as API-605. Tables of pressure/ temperature ratings and standard dimensions are provided in US customary units only. The pressure – temperature rating tables are basically the same as those in ASME B16.5 except for the addition of Class 75. EN ISO 10423 is identical to ANSI/API Specification 6A and covers flanges for high pressure applications, such as wellhead and “Christmas tree” equipment used in the oil and gas industry. Three types of flange are covered (all ring joint type):

• Type 6B flanges are available as weld neck, threaded, integral (long weld neck) or blind flanges and the bolting force reacts on the metallic ring gasket.

• Type 6BX flanges are available as weld neck, integral (long weld neck) or blind flanges. The bolting force can react on the raised face of the flanges when the ring-joint gasket has been properly seated. This prevents damage to the flange or gasket from excessive bolt torque, but is not essential for proper functioning of the flange.

• Segmented flanges have a recessed face, and the bolting force can react on the surface outside the recessed face of the flange when the ring-joint gasket has been properly seated. This prevents damage to the flange or gasket from excessive bolt torque, but is not essential for proper functioning of the flange.

The maximum rated working pressures and size ranges of type 6B, 6BX and segmented flanges are given in Table 1. Table 1 – Rated working pressures and size ranges of flanges to EN ISO 10423 Rated working pressure

Flange size range mm (in)

MPa (psi) Type 6B Type 6BX Segmented 13.8 (2000) 52 to 540 (2 1/16 to 21

¼) 680 to 762 (26 ¾ to 30) -

20.7 (3000) 52 to 527 (2 1/16 to 20 ¾)

680 to 762 (26 ¾ to 30) -

34.5 (5000) 52 to 279 (2 1/16 to 11) 346 to 540 (13 5/8 to 21 ¼)

35 to 103 x 108 (1 3/8 to 4 1/16 x 4 ¼)

69.0 (10000) - 46 to 540 (1 13/16 to 21 ¼)

-

103.5 (15000) - 46 to 476 (1 13/16 to 18 ¾)

-

138.0 (20000) - 46 to 346 (1 13/16 to 13 5/8)

-

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Standard flange dimensions are tabulated in both metric units (mm) and US customary units (inches). Information is given in the standard for evaluating the rated working pressure for elevated temperatures. A new edition of EN ISO 10423 was published in December 2009, but has not yet been issued as a BS EN ISO standard. PN designated flanges EN 1092 covers the pressure designations PN 2.5, PN 6, PN 10, PN 16, PN 25, PN 40, PN 63, PN 100, PN 160, PN 250, PN 320 and PN 400, and sizes from DN 10 up to DN 4000 (for PN 2.5 flanges). The upper size limit reduces for the higher pressure ratings. The PN designation indicates the pressure rating of the flange in bars at ambient temperature. The maximum allowable pressures at other temperatures are obtained from the pressure - temperature rating tables given in the appropriate part of EN 1092. Part 1 covers steel flanges, Part 2 covers cast iron flanges, Part 3 covers copper alloy flanges and Part 4 covers aluminium alloy flanges. Standard flange dimensions are also tabulated. Flange configurations Standard flanges are available in a variety of combinations of type of flange and facing. The types of flange include weld neck, long weld neck, slip-on, socket welding, lapped, threaded and blind. The most commonly used facings are raised face, flat face and ring joint, but other facings such as tongue and groove and O-ring groove are also used. Weld neck - this type of flange has a tapered hub at the back of the flange and is butt welded to the pipe or nozzle neck, as shown if Figure 1. The butt weld can be subjected to volumetric examination (radiography or ultrasonics) to ensure a high integrity joint. This type of flange is widely used in the oil, gas, petro-chemical and power generation industries.

Figure 1 - Weld neck flange Long weld neck - this type of flange is used for nozzles on equipment as an alternative to using thick walled pipe. The nozzle neck is replaced by an extended parallel hub at the back of the flange, as shown in Figure 2.

Figure 2 - Long weld neck flange

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Slip-on - this type of flange fits over the outside of the pipe or nozzle neck and is attached with fillet welds at the back and the face of the flange, as shown in Figure 3. The welds can only be checked by surface examination techniques. This type of flange is not recommended for high temperature applications or cyclic service.

Figure 3 - Slip-on flange Raised face. Figure 4 shows a flange with a raised face for gasket seating. This is the standard facing for use with gaskets which are located inside the bolt circle, and a wide range of gaskets is available.

Figure 4 - Raised face flange Flat face. The face of the flange is flat, as shown in Figure 5, and is used in conjunction with a full face gasket which extends beyond the bolt circle. Relatively soft gasket materials are generally used. This type of facing is best suited to low pressure applications, and has the advantage that the gaps between the flange faces at the inside and outside surfaces can be eliminated where cleanliness is important.

Figure 5 - Flat face flange Ring joint. The face of the flange has a groove for use with a metallic ring type joint, as shown in Figure 6. Ring joint facings are generally used in high pressure and/or high temperature applications.

Figure 6 - Ring joint flange

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COMPACT FLANGED CONNECTIONS (CFC) There are various proprietary flange designs on the market as an alternative to the standard flanges described above, including Taper-Lok, Vector SPO, Desflex and Verax. There are many others. Compact flanges are used in a variety of industries, including oil and gas (onshore, offshore and subsea applications), petro-chemical and power generation. The advantages include:

• Many designs use a reusable seal. • High quality of leak tightness. • Compact design reduces space and weight (up to 70% lighter than the conventional

flanges). • Reduced weight gives substantial cost benefit with expensive materials. • Smaller bolt diameters making assembly and installation easier.

Disadvantages

• Most piping and vessel codes do not give automatic exemption from design calculations. • Can only be joined to another flange of the same type. • Some designs have male and female flanges. • Most designs require flanges to be separated to insert or remove seal.

Norsok L005 The only standard for compact flanges is the Norwegian Norsok L-005, however a committee draft of an ISO standard based on Norsok L-005 has recently been issued for comments. The Norsok standard is based on common principles utilized by VERAX, Vector International AS and Off.N.Galperti SpA. The compact flange described below (and in clause 5 of the Norsok standard) is based on the SPO compact flange developed by Vector International AS. The flange face includes a slightly convex bevel with the highest point, called the heel, adjacent to the bore and a small outer wedge around the outer diameter of the flange. The assembly is made up by tightening the flange bolting which pulls the two connector halves together. Axial forces are exerted on the taper of the metal seal ring and translated into a radial sealing force. As the bolt load is increased the bevel is closed and face to face contact is achieved at the outer wedge. Most of the bolt pre-load is transferred as compressive forces between the flange faces at the bore. The flange design incorporates two independent seals. The first is created by application of seal seating stress at the flange heel. The heel contact will be maintained for pressure values up to 1.8 times the flange rating at room temperature. The main seal is the IX seal ring. The seal ring force is provided by the elastic stored energy in the stressed seal ring. Any leakage at the heel will give internal pressure acting on the seal ring thereby increasing the sealing action. The design aims to prevent exposure to oxygen and other corrosive agents to prevent corrosion of the flange faces, the stressed length of the bolts and the seal ring. When the flange is bolted up the back face of the flange is parallel to the flange face in order to prevent bending of the bolts in the assembled condition. Flanges covered by a class of Clause 5 of Norsok L-005 will stand the maximum rating of the corresponding ASME B16.5 class over the temperature range covered by the Norsok standard. Tables of standard dimensions are provided for sizes in the range DN 16 (NPS ½) to DN 1200 (NPS 48), except for CL 2500 which has an upper limit of DN 600 (NPS 24).

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Table 2 – Pressure class designation and ASME rating ceiling values to ASME B16.5

Pressure class Class abbreviation

Nominal pressure

ASME pressure rating ceiling values

psig barg Class 150 CL 150 PN 20 290 20.0 Class 300 CL 300 PN 50 750 51.7 Class 600 CL 600 PN 110 1500 103.4 Class 900 CL 900 PN 150 2250 155.2 Class 1500 CL 1500 PN 260 3750 258.6 Class 2500 CL 2500 PN 420 6250 431.0

Taper-Lok® The Taper-Lok® Weld Neck Assembly is a compact flange comprised of a male flange, a female flange, a seal ring, and a complete set of studs and nuts. Taper-Lok® is a registered trade mark of Taper-Lok Corporation. The design is made up of two converging angles based on the wedge principle. The male nose is a 20° angle cone, and the female contains a 10° pocket. The Taper-Lok® seal ring, with comparable angles, sits in between the flange components and acts as a “door stop” by creating a wedge. The tapered seal ring geometry design ensures a significant length of the sealing surfaces as contact forces are generated between both the male and female components; this geometry is what gives all Taper-Lok® flanges a self-energizing and pressure-energizing seal. Taper-Lok® flanges require lower bolt loads than standard connections. The seal ring is generally made of the same material as the flange and is reusable. Standard Taper-Lok® connection sizes range from 1/2" to 83" with varying wall thicknesses, sealing pressures up to 40,000 psi, and temperatures ranging from -350º to 1600º F. Variations of the basic weld neck design are available for blind flanges, long weld neck flanges, heat exchanger closures, swivel flanges and other applications. Desflex The Desflex compact flange is manufactured by Destec Engineering Ltd and uses a ‘D’ type metal-to-metal seal which is flush with the bore of the flange. The flange stresses during assembly are controlled by limiting the flange rotation via a small gap at the outer edge of the flange. The flanges are more resistant to external bending, and excessive bolt tightening cannot overstress the flange. Desflex flanges are available in sizes from 1” NPS up to 40” NPS, and pressure rating classes 300, 600, 900, 1500 and 2500. Destec provide their own pressure rating tables that are based on the stress analysis methodology in ASME VIII Division 1, Appendix 2. Desflex flanges are available as weld neck, blind and swivel flanges. Verax The concept of the Verax compact flange (VCF) originated as far back as the early 1950s. The VCF does not principally use seal rings or a gasket, although these can be added if required. This means that normal installation and assembly of equipment can be easier as components should slip into place. Since there is no gasket present, the assembly operates in a ‘static mode’. Verax specify that the bolts should be tightened to 80% of the yield strength, so once assembled and tightened, the bolt loads remain steady and do not change over time when the pressure is applied. This is not the case with a gasketted joint. Verax claim that the VCF reduces corrosion in the assembly as neither the flange faces nor the loaded part of the bolts are exposed to the internal media or external environment. As there is full metal-to-metal contact, interface corrosion is eliminated. The VCF system performs well on the failure mode evaluation analysis, and risk of leakage is minimised with this approach. Annual monitoring of the VCF system is not required and VCF systems comply with the 4 year schedule in accordance with US-EPA legislation. The VCF must be handled with care and be assembled correctly. Most VCF joints have a greater number of smaller bolts than standard flanges. This gives more uniform bolt load around the circumference and better feel for the operator, but takes more time.

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In addition, the mating faces must be scratch free. Some minor scratches are permitted, but since this face is the primary seal, good operator training and installation procedure must be adopted. PROPRIETARY CLAMP CONNECTORS Clamp connectors consist of a pair of hubs for that are welded to the ends of the pipe (similar to a flange), and a seal ring; but the normal flange bolts are replaced by a clamp set, which can be rotated around the hubs to suit the most practical position. There are several designs available, including Grayloc, Taper-Lok, Vector Techlok and Destec. Clamp connectors are used in a variety of industries, including oil and gas (onshore, offshore and subsea applications), petro-chemical and power generation. The advantages include:

• Many designs use a reusable seal. • High quality of leak tightness. • Smaller and lighter than conventional flanges. • Only four bolts to tighten, making maintenance simpler and quicker. • No periodic retightening of the bolts is required when the connector is in service.

Disadvantages

• Most piping and vessel codes do not give automatic exemption from design calculations. • Can only be joined to another flange of the same type. • Some designs have male and female flanges.

Compared to a standard flange, clamp connectors are significantly lighter and smaller. There are only four bolts to tighten, making maintenance considerably simpler and quicker. No periodic retightening of the bolts is required when the connector is in service. Grayloc® The Grayloc connector has three basic components – the metal seal ring, the two hubs and the clamp assembly. The metal seal ring achieves a self-energised and pressure-energised bore seal that will hold vacuum or external pressures. The hubs are welded to the ends of the pipe, and as they are drawn together by the clamp assembly the seal ring lips deflect against the inner sealing surface of the hub, forming a self-energising seal. The two piece clamp assembly is the primary pressure retaining component, not the bolting. The clamp carries all the internal pressure loads as well as axial and bending loads transmitted by the pipe. Grayloc is a registered trade mark of Oceaneering International Inc. Taper-Lok® The Taper-Lok Clamp Connector is similar to the Grayloc connector, but utilises the tapered sealing ring as fitted to the Taper-Lok compact flange. Vector Techlok The Vector Techlok Clamp Connector is similar to the Grayloc connector, and utilises a self-energised and pressure-energised metal seal ring at the bore of the flange. Destec G-Range The Destec G-Range clamp connector is also similar to the Grayloc connector, and utilises a self-energised and pressure-energised metal seal ring at the bore of the flange. CUSTOM DESIGNED FLANGES Custom designed flanges are used when the diameter does not match that of a standard flange, or when a better optimised design is required. For example, standard ASME B16.5 flanges generally have a fairly small number of large bolts, rather than a larger number of smaller bolts. This increases the bolt circle diameter and flange outside diameter, which in turn increase the bending moment in the flange and hence the flange thickness. The end result is a flange that is considerably heavier than an optimised design. When expensive alloy materials are being used this will have significant cost implications.

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Flange design methods are given in most pressure vessel design codes, such as EN 13445, ASME VIII and PD 5500. Most of these are based on what is generally known as the “Taylor Forge Method”. Alternative design methods are given in the EN 1591 series of standard. These design methods will be covered by other presentations at this seminar. Custom designed flanges are commonly used for the girth flanges in shell and tube heat exchangers, vessels and other pressure equipment where there is a requirement for sections to be removable. The advantages of using a custom designed flange are:

• Can be designed for the specific design conditions. • Designed for specific flange, bolting and gasket materials. • Usually smaller and lighter than a standard flange.

Disadvantages:

• Design calculations must be performed. • Longer delivery time compared with a standard “off the shelf” flange. • Total cost may be greater than a standard flange.

QUICK RELEASE OPENINGS Many bolted flanged joints stay in service for long periods (several years) without being dismantled. Others, such as access openings, may be dismantled and reassembled on a regular basis, and this will affect the type of flange selected. One option is to use a design similar to a traditional bolted flange, but with swing bolts or quick release clamps instead of conventional through bolting.

Figure 7 - Quick release clamp For access openings various types of quick release manways are available. These are generally significantly lighter than a standard blind flange, and with fewer bolts. All these openings still require the loosening of a number of bolts in order to gain access. If more rapid access is required there are several proprietary quick release openings on the market, including those offered by GD Engineering, Perry Equipment Corporation, Pipeline Engineering and T D Williamson. These are not strictly bolted flanged connections, but they serve the same purpose. These are usually in the form of a hinged door with some form of quick acting locking mechanism instead of bolts. Various safety features are incorporated to ensure that the door cannot be opened while the equipment is pressurised.

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Typical quick opening closure applications include: • Pipeline pig traps • Filters • Coalescers • Strainers • Separators • Meter skid systems • Hydrocyclones

The main advantages are:

• Rapid access • Reusable seal • Safety interlocks

Disadvantages

• High cost compared to a standard flange CONCLUSIONS The standard flange has served the pressures systems industry reasonably well for over 80 years. However, due to increasingly more demanding operational requirements, various manufacturers and industries have adjusted, improved and even redesigned the bolted flange over time. The main issues of strength, deflection, leakage, weight and cost remain, and users must be fully aware of the design basis and operational limits of each system.

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Gasket Selection Dr Gavin Smith, Novus Sealing Limited

© Dr Gavin Smith, Novus Sealing Limited

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Gasket SelectionDr Gavin Smith, Technical Director

Novus Sealing Limited

To ensure safe operation of a bolted flange connection the gasket must be:

•Correctly SELECTED

•Of the right QUALITY

•Properly ASSEMBLED

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•The gasket should be suitable for the design or operating conditions:

•The process fluid at the operating temperature •The operating temperature•The operating pressure

Fluid Temperature Pressure

•There is a wealth of data from both gasket manufacturers and plant history on the compatibility of gasket materials with process fluids

•However, despite all this knowledge problems do occur

•A good (or bad) example are the numerous failures of nitrile elastomer seals that occurred on the change from Diesel to Bio‐Diesel

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Failure of Nitrile Gaskets in Bio‐Diesel

•Gaskets may seal well initially but can fail over time at temperature

• Creep and Stress Relaxation•A gasket may seal well initially but over time will lose load which may result in flange leakage 

•Oxidation •Graphite will oxidise at elevated temperature at a rate determined by the temperature, the oxygen concentration and the quality

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Oxidation of Graphite in a Spiral Wound Gasket

•The resistance of a gasket material to the internal pressure is related to its ability to withstand the load applied. 

•Gasket Stress is the key parameter: Defined as the as the total applied bolt load divided by the compressed area of the gasket

•Gasket stress defines the load bearing characteristics of the gasket and is used to calculate the torque applied to the bolts during assembly. 

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Every gasket has a minimum and maximum stress

1 2 3 4 5 116 7 8 9 10 12

Stud Number

Gasket  Stress

Min

Max

Non‐metallic gaskets have a low minimum and low maximum stress

1 2 3 4 5 116 7 8 9 10 12

Stud Number

Gasket  Stress

Min

Max

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Metallic gaskets have a high minimum and high maximum stress

1 2 3 4 5 116 7 8 9 10 12

Stud Number

Gasket  Stress

Min

Max

2” Flange Size by Pressure Class

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Setting the target stress

1 2 3 4 5 116 7 8 9 10 12

Stud Number

Gasket  Stress

Min

Max

Target Stress

1 2 3 4 5 116 7 8 9 10 12

Stud Number

Gasket  Stress

Max

Target Stress

Temperature effects will reduce the stress on the gasket significantly (all gasket relax!)

Service Stress

Relaxation

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1 2 3 4 5 116 7 8 9 10 12

Stud Number

Gasket  Stress

Max

Target Stress

Solution is to set stud loads high and select a gasket with high resistance to relaxation

Service Stress

Relaxation

Min

Gasket Selection for Heat Exchangers

• Gasket Load Loss from Relaxation

• Inability to Tolerate Relative Movement Between the Flanges

The two main reasons flange connections on heat exchangers leak are: 

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Differential Expansion in a Heat Exchanger

Differential Radial Expansion Of Channel and Shell Flanges, Relative To The Tubesheet, Over 21 Days

-0.01

-0.009

-0.008

-0.007

-0.006

-0.005

-0.004

-0.003

-0.002

-0.001

0

0.001

0.002

21:0

8

11:3

8

2:08

16:3

8

7:08

21:3

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12:0

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2:38

17:0

8

7:38

22:0

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8:08

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3:38

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8

8:38

23:0

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4:08

18:3

8

9:08

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14:0

8

4:38

19:0

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9:38

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14:3

8

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8

"X" Axis Shows Time With Data Taken Every 30 Minutes

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Channel Flange Shell Flange

Startup Following A Plant Shutdown

Exchanger Channel Side Flow Stalls

Second Restart of Exchanger

Normal Operation For This Exchanger

Effect of Differential Expansion

•Differential Expansion leads to differential movement between the mating flanges

•Flange movement results in shearing of the gasket or leads to slippage at the gasket / flange interface

•Double Jacketed gaskets are unable to tolerate this movement between the flanges. 

33

Failure of a Double Jacketed Gasket

Graphite faced gaskets are the best solution for heat exchanger applications

•Corrugated Metal Gasket

•Camprofile Gasket

•Spiral Wound Gasket

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Quality

•Once the gasket has been correctly selected it must be manufactured to the highest quality

•Unfortunately, failures do to poor quality gaskets remain a problem

•As an example, lets have a look at the Camprofile….

The sealing integrity of a Camprofile relies upon precise standards of machining

35

There are three methods of manufacture

• Bend and weld pre‐profiled strip

• Bend and weld strip and profile

• Laser cut rings and lathe profile

Failure due to poor quality weld

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Failure due to poor quality weld

Bend and Weld construction. Poor!

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Bend and Weld ConstructionReally Poor!!

Welds must be machined down to the same height as the metal core

Failure Point –thinner material and no serrations

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The best solution is no welds

All graphite looks the same, but looks can be deceiving!!!

Low Quality Graphite High Quality Graphite

Basic oxidation test at 600°C, 4 hours

39

But ash content does not guarantee oxidation rate

0

0.5

1

1.5

2

2.5

3

3.5

4

4.5

5

Sample A Sample B Sample C Sample D Sample E

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Ash

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Ash Content Weight Loss

The reliability of the flanged joint depends on competent control of the joint making process

Well lubricated studs and nuts

Controlled Tightening

Trained technicians

40

Conclusions

•�A gasket is a relatively low cost item but it is critical to thesafe operation of any plant. 

•To ensure safe operation a gasket must be:

•Correctly Selected•Of the right quality•Properly Assembled

•Use your gasket provider. They have a wealth of data and experience that can ensure a leak free, safe plant.

Contact Details

Dr Gavin SmithTechnical Director

Novus Sealing LimitedTel: 07785247202

email: [email protected]

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42

Material Selection for Industrial Fasteners

Rod Corbett, James Walker Rotabolt © Rod Corbett

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Material Selection for Industrial Fasteners Rod Corbett, Managing Director, JamesWalker Rotabolt 1 Introduction Material selection for fasteners depends on the service environment, the load carrying requirement and the cost of a joint for an expected service life. The main selection categories are :- Tensile and fatigue strength Elevated and cryogenic temperatures Corrosion resistance One or more of these categories has to be analysed for a particular application before the fastener material is selected. 2 Tensile and Fatigue strength The main considerations for bolt strength selection in any environment are:- Achievement of design bolt tension/joint compression/gasket seating stress. Assured joint reliability Increased fatigue life Service load carrying capability in tensile and shear. Reduced equipment build costs – fewer, smaller bolts for the same service. Generally, tensile strength is the most important consideration in fastener selection. The bolt has one objective – to deliver a known level of bolt tension and subsequent equal and opposite compression in the flanged joint, within a safe elastic strength margin. This level of tension when achieved delivers assured joint reliability – no leaks from pressure containment, no fatigue failure, no self loosening, and no structural slip. 2.1 Hardenability ISO 898 10.9 grade is widely used for high performance structural usage e.g. cranes. This strength level combines high strength with good ductility. Higher strength grades such as 12.9 and 14.9 have decreasing ductility and increased susceptibility to brittle failure so 10.9 is an optimum choice in difficult environments. For strength of 1040 MPa minimum, fasteners can be manufactured from low alloy steels. The main alloying elements Chromium, Nickel and Molybdenum enhance the mechanical properties of the steel. Their main effect on carbon steel is to increasing hardenability. This must no be confused with hardness which is dependant mainly on carbon content. The higher the carbon content, the higher the hardness, potentially. Other alloying elements that are normally present in steel also affect hardenability.

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The list of hardenability agents are as follows: Carbon Carbon is a strong aid to hardenability. It is advisable to control carbon contents. However, excessive carbon decreases forge-ability, causes embrittlement problems in heat treatment as well as room temperature/low temperature applications. Manganese

Strongly increases hardenability. However, it reduces forgeability. Chromium Molybdenum Boron }

Most effective hardenability agents

When present in a small amounts (0.001%) it has a pronounced effect, particularly in lower carbon steels.

Their effect on hardenability is demonstrated in the presentation slides covering tensile strength. You will note that boron is the most powerful agent. The use of plain carbon boron steels are not recommended for use in large diameter, high strength, and high performance applications such as large diameter slew ring bolting. Boron should only be used to boost hardenability of low alloy steels maybe for larger diameter fasteners however it is important to consider the operating environment especially with respect to stress corrosion cracking and elevated temperatures. Note that for larger diameters, steels with greater alloy content are needed. For strength levels greater than 1220 MPa, 826M40 or SAE 4340 should be selected. Finally, although different materials have the same tensile strength, their fatigue properties can be different. 2.2 Ideal material for high strength Ideal pre-requisites for candidate materials to be used in the manufacture of high strength, high performance fasteners could be: - The tempering temperature of at least 480C allows the material to be used in fasteners that will satisfy several different markets. A threaded fastener is really a component containing a series of notches. The criterion for high notch strength is that the material’s notched tensile strength (at kt = 6) must be equal or greater than the materials smooth bar tensile strength. The material must have sufficient ductility at high strength levels and minimum 7% elongation is indicative of this. Hardenability through a 2 inch diameter enables the manufacture of fasteners throughout a size range, and subsequently satisfies. Be hot and cold forgeable Materials are resistant to environmental embrittlement. Although stress levels have a big say in susceptibility, increasing alloy content, especially molybdenum also increases resistance.

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Good thread fatigue resistance as the most common form of bolt failure is fatigue. 2.3 Application of high strength bolting We have mentioned that increased tensile strength enables the use of fewer and smaller diameter fasteners, resulting in weight reduction. This is of prime importance in the aerospace industry but is important in other industries as reduced weight means reduced cost (smaller bolt quantities, reduced diameters, smaller number of bolt holes, less tightening cycles). This design concept as also been used in the design of latest technology wellhead equipment where traditional 8 or 12 bolt flanges have been reduced to four or six bolts of the same diameter. Whilst alloy steels and super alloys are capable of developing much higher strength levels, their strength to weight ratio is not as good as titanium. This advantage along with its excellent corrosion resistance would seem attractive to the offshore industry but as yet hasn’t been used extensively. 2.3.1 True Strength of bolting Medium carbon low alloy steels are used for high strength bolting that is used in a wide range of environmental conditions ranging from the benign to the hostile. Alloy steel bolting is relatively low cost but has a limited service life. The increasing demand in most industries for longer service life with reduced maintenance costs has led to the assessment and use of non-ferrous alloys which have inherently superior environmental resistance. One design pre-requisite for the candidate alloy is that it has similar mechanical strength and properties to the alloy steel. Many of the new alloy developments are produced with similar UTS and 0.2% proof stress values to their steel counterparts. 0.2% proof stress is a traditional bench mark for a bolt’s yield strength. The following schematic shows typical stress strain curves for alloy steels and non-ferrous alloys. They may have similar 0.2% proof stress values but their true elastic limit is significantly different. With many alloys designed for use in hostile environment it is a fact that their elastic strength capacity is significantly less than their medium carbon low alloy steel counterparts. The effective strength reduction can be as much as 30-35% below the 0.2% proof stress value compared with a nominal 12-15% with alloy steels. Indeed the British Steel Advisory centre has recommended that engineers use elastic strength assessments based on 60% of the specified 0.2% proof stress for austenitic stainless alloys.

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Incremental load extensions tests, carried out on M22 all thread and double ended studs manufactured from the relevant alloys, revealed the following true elastic limits for various materials. Alloy condition 0.2% proof true elastic Stress N/mm2 limit N/mm2 Monel k500 GR MA 18 655 475 Ferralium 225 fully heat treat 649 441 Marinel 21A ------------- 720 487 Titanium Beta C solution treated 834 679 Titanium 6AL-4V ELI solution treated 830 593 Austenitic Stainless Work hardened 600 360 These substantial reductions in effective fastener strength are very significant in energy industries, especially when you consider how most of these fasteners are installed. Hydraulic tensioning is the most common tightening method for larger diameter studs. Because of the need to compensate for load transfer relaxation the studs have to be hydraulically overloaded appreciably in excess of the design tension target. Initial fastener diameter selection will be based on the latter and the materials specified 0.2% proof stress. With most of the non-ferrous bolting alloys having up to 30% lower effective elastic strength compared to the specification bench mark the prospect of yielding or even breaking bolts on installation is significant. 2.3.2 Causes of low elastic strength in alloy steels Whilst the above schematic indicates a relatively high elastic limit for alloy steels, recent findings suggest that larger diameter bolting can have much lower elastic strength than that suggested by its certified 0.2% proof stress. Bolting made from alloys such as EN24, EN25, EN26 and larger diameter B7 are showing between 25-40% deficiencies compared to their certified proof street values. This is especially so for stud bolting that is made simply by thread forming bar stock that is already supplied in a heat treated condition that matches the required finished fastener mechanical properties; rather than heat treating the finished fastener to the specific fastener mechanical properties. There are a number of reasons that could explain this low elastic strength. Bar stock, such as EN24’V’, is produced for the manufacture of any metallic component, not just bolts; quite often the properties are at the bottom of the tensile range. Also the manufacturing process leaves residual stresses in the bolt and when further process deformation occurs, say during bar straightening etc., the effect is to lower a bolts yield or flow stress. This is illustrated in the following schematic.

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There are consequences for in-service performance. It is difficult to know if a bolt has been yielded on tightening. If it does occur, the following results:- Reduced joint clamp loads Bolt is highly stressed, with likely increased hardness. This maybe significant in terms of resistance to environmental embrittlement. Both situations are potentially detrimental to the performance of the bolted joint.

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3 Corrosion resistance Selection here not only depends on required strength but on the service environment too. Ordinary alloy steel fasteners may be perfectly satisfactory in certain applications and environments where merely protected by a surface coating. They can be more cost effective than corrosion resistance materials. However, let us concentrate on severe conditions where the fastener must have a long life in a hostile environment. 3.1 The most common group of materials is stainless steels. As the name implies, these steels are more resistant to rusting and staining than plain carbon and lower alloy steels. The superior corrosion resistance is brought about by the addition of chromium. These are the four basic types:- 3.1.1 Austenitic stainless E.g. 18% chromium -12% nickel This type of material cannot be strengthened by heat treatment. Any strength this stainless steel has comes from cold work or deformation during its production cycle e.g. during raw material bar rolling or cold forging and thread rolling. As indicated previously, their true elastic limit is significantly lower than the specification stated 0.2% Proof Stress. 3.1.2 Martensitic These steels are hardenable by heat treatment in the same way as carbon alloy steels. Strength levels similar to low alloy steels can be achieved up to 1200 MPa subject to section size. It is worth noting however that these steels are also prone to lower than expected elastic limits compared to the specification stated 0.2 % Proof stress value. 3.1.3 Precipitation hardening Precipitation hardening is a heat treatment process similar to hardening and tempering with low alloy steels. This heat treatment can develop strength levels as high as 1500 MPa with alloys such as PH13-8Mo and A286. In the offshore industry the derivative for A286 is B17 or A453 660 grade primarily used for sour gas applications or where higher strength is needed compared to austenitic stainless steel. 3.1.4 Ferritic stainless steels There is no demand for fasteners made from this material, they cannot be heat treated and tend to be very notch sensitive and have very poor creep strength. The material is used predominantly in acid handling applications. The nominal compositions for stainless steels seem similar. However, the alloy contents in the composition matrix determine whether the stainless steel is austenitic or martensitic etc. Nickel and Nickel equivalent elements, such as manganese, promote austenite. Chromium and chromium equivalent elements such as molybdenum, promote martensite and ferrite type stainless steels.

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3.2 Cupro Nickels and High Nickel alloys One family of corrosion resistant materials used in the offshore industry are Cupro Nickels. As the name suggests the main elements in the alloys are Copper and Nickel. Monel K 500, Marinel and more recently Nibron have all been used but as with the stainless steels their true strength is significantly below that suggested by the 0.2% Proof Stress stated in relevant specifications. Many applications can be accommodated with fasteners made from aforementioned materials. However, once the environment severity increases even further, materials such as Inco 718 and Hastelloy have to be used. For the most severe, extreme cases e.g. sour gas environments at the bottom of the oceanic oil wells, proprietary alloys called Multiphase will provide a fastener with the optimum solution. This alloy is a nickel cobalt quaternary, available in two compositions, and has ultra high strength 1800 MPa and fatigue resistance. It is also immune to stress corrosion cracking and hydrogen embrittlement. 3.3 Environmental Embrittlement A very common environmental failure mechanism is stress corrosion cracking (SCC). A combination of stress, susceptibility and a corrosive environment causes stress corrosion cracking. Initial pitting of the metal surface takes place and leads to a stress concentration. The effect is cumulative and, in a highly stressed joint, it can lead to very sudden failure. Both trans and intergranular attack of the metal takes place in SCC but the failure is generally characterised by a brittle intergranular fracture. The amount of corrosion involved can be very small but its effect can be catastrophic. SCC can be avoided through material selection based on the following factors:- Keep the material stress below a critical threshold level for that alloy. Use a stress corrosion cracking free alloy e.g. Multiphase, Inco 718. Protect the fastener from corrosion e.g. surface coat alloy steels.

A typical application of SCC prevention is on offshore pedestal cranes where most slew ring bolting/boom bolting is now 10.9 strength grade. At 12.9 grade the maximum specification alloy hardness exceeds the threshold for many medium carbon low alloy steels so they become susceptible. The same stress threshold concept exists for other embrittlement failures such as hydrogen embrittlement (HE). Whilst generally there is no corrosion in this type of failure, the failure mode is virtually identical to SCC. Hydrogen diffuses into small voids near to the surface of the metal, and embrittles the lattice structure, thereby lowering the threshold stress level for brittle failure. Possible sources for this hydrogen are:

1. Using alternative cleaning methods e.g. the use of dry cleaning methods such as aluminium blasting instead of an acid pickle.

2. Ensure post plating baking procedures are carried out to drive out any hydrogen that has diffused into the fastener during plating;

3. At the joint design stage, ensure compatibility of joint materials.

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4 Elevated temperature applications. The creep and rupture strength of steel can be greatly improved by the addition of alloying elements. Molybdenum greatly increases creep and rupture strength. Tungsten and vanadium have a similar effect. Chromium has a negative effect on heat resistance, but one needs chromium present for oxidation resistance. Cobalt increases the hot tensile strength and temper resistance. These elements therefore, are very important in selecting materials for elevated temperature fasteners. The following groups of materials are used for elevated temperature applications. Alloy steels (up to 10% alloy content) Austenitic stainless steels Precipitation hardening stainless steels. All these are iron based materials and are generally used from 350C to 550C. Alloys based on Iron, Nickel, Chromium and Molybdenum, for example B17/660 grade, are used at temperatures up to 650C. Nickel based alloys e.g. Nimonic, Waspaloy and Inco 718 are used where operating temperatures range from 650C to 850C. With temperatures in this region, creep, oxidation and hot strength are major problems. Materials selected for these applications therefore, contain sufficient quantities of Nickel, molybdenum and cobalt. For extreme temperatures greater than 1000C, refractory materials based on tantalum have to be used. For the final environment we will cover on material selection, we will go from the extreme of very high temperature to the opposite of low temperature or cryogenic application. 5 Cryogenic Applications Selection for cryogenic applications is dependant mainly on the crystallographic structure of the candidate material. Alloys with the body centre cubic structures lose ductility at lower temperatures and tend to have a threshold temperature below which they go brittle. Materials selected for cryogenic applications tend to have faced centred cubic structures.

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Typical selection use for low temperature applications range from :- Iron based – A320 L7(BCC structure), Austenitic stainless B8, (FCC structure) A453 660/B17; Nickel based alloys include Inco 718 and Nimonic 80. Many materials have a limited temperature range usage but the above illustrates that selection for elevated temperatures features the same alloys for cryogenic applications. Inco 718, PH13–8MO, A286, Wasploy and B17 have the advantage of a wide temperature range. They all show excellent strength and ductility at low temperatures, and retain tensile strength at their maximum utilisation temperature. In offshore and energy sectors, alloys such as B8/B17 are used for high and low temperature service.

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6 Material Specifications 6.1 Structural Bolting Most of the high performance structural bolting in industry is specified to a strength level and is manufactured from medium carbon low alloy steels. In the main, specifications such as 1SO 898, BS3692 allow the bolt manufacturer to make the material selection for the strength grade required unless otherwise stated by the customer. Provided he meets some basic minimum alloying element requirement, he is supposed to have the experience and knowledge to make the selection and thereby guarantee the mechanical property specification e.g. 8.8 or 10.9 etc. The same can be said for SAE J429. A traditional imperial material specification is BS1768 which also categorises by strength grade. The DIN Euronorm specification tends to be different and categorises by material with the material strength grade being determined by the bar stock or bolt diameter in that material. 6.1.1 Strength. The all embracing standard is covered by the B7 designation. The ultimate tensile strength of B7 is regarded as high in energy industry and is a bench mark for all other types of environmental alloy bolting such as duplex stainless and cupro-nickel. Compared to structural steels, specifications such as ISO 898 10.9 grade, B7 strength is significantly lower however the lower hardness tends to be below threshold levels for environmental embrittlement such as SCC and hydrogen. Comparative table B& versus 8.8 versus 10.9 B7 is covered by A193 and the similar BS 4882. The material alloy is a medium carbon low alloy steel containing nominally 1% Chromium and 0.25% Molybdenum. Both specifications are typified using a constant composition over the full bolt diameter range. This means for larger diameter bolts, the B7 alloy has insufficient hardenability to provide constant tensile strength across this range – Table . where the highest entry level B& strength is required on larger diameters, designers often call for A540 B24. The alloy here is more commonly known as SAE 4340. It is capable of much higher strengths than B7 and its chromium, molybdenum and high nickel content of 2% creates deep hardenability enabling high strength and ductility at the largest bolt diameters. The A193 and BS 4882 specs extend to an elevated temperature capability because of the alloy composition. BS 4882 stretches out the use of B7 to approx 450 C and then by adding Vanadium to create B16, the V resists tempering effects pushing its allowable design usage to 525C. One could argue this is the material’s absolute limit so great care on service longevity and replacement strategy has to be taken along with assured control on installed design bolt tension objectives if it is to provide a cost effective bolted joint at these maximum temperatures. The B7 designation is mirrored by L7 for low temperature usage. Mechanically and composition wise they are identical, the only difference being L7 has a low temperature charpy test requirement. This qualifies it to be used at temperatures of the order of minus 100C. As with B7, larger diameters have lower tensile strength because of hardenability constraints. Once again where there is a requirement for the larger diameters to have the highest specification strength, L43 (4340) alloy must be used.

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The increased toughness at through hardened strength from the higher nickel is especially effective at lower temperatures found in LNG operations for example. A320 also has some other strange material options including a plain carbon steel with added Boron for hardenability. Having no experience of such a requirement, the author can only summise, it is an economy option for high volume, small diameter bolting on a process site; 6.1.2 Environmental selection - Lower Bolt Strength. Where medium carbon low alloy steel fasteners are required to operate in corrosion environments they need to be resistant to embrittlement mechanisms such as stress corrosion and hydrogen. Immunity can be achieved by reducing the strength/hardness of the fastener below a threshold value below which the mechanism will not initiate. In terms of hydrogen embrittlement and general stress corrosion cracking the standard B7 strength hardness of the chromium molybdenum alloy is low enough to ensure these types of failure will not occur. However certain, hostile environments are such that the strength level has to be reduced to an even lower threshold. Operating environments with high sulphur/hydrogen sulphide present are such an example; the failure mechanism in these environments is sulphide stress cracking. A193 and BS4882 designate the lower strength B7M as the selection grade for such an environment. 6.1.3 Elevated / High temperature/ cryogenic applications. Where operating temperatures exceed the absolute maximums for medium carbon low alloy steels, the use of austenitic and precipitation hardening steel alloys must be used for resistance to heat , creep and oxidisation and maintain installed bolt tension/joint compression. Austenitic stainless steels are designated B8. There are two versions, one high strength A193 B8 class 2 or BS4882 B8X; the other low strength A193 B8 class I or BS4882 B8 not ‘X’ categorised. The lower strength B8, is in the carbide solution treated condition and has a constant low tensile strength across the full size range. Because austenitic stainless steels cannot be heat treated to increase strength, higher strength requirements must come from the cold working and subsequent deformation induced during fastener manufacture. As with larger diameter carbon steel bolts having through hardening constraints for a certain alloy composition, the effect of the cold work/deforming forces go from maximum at bolt surface layers and steadily reduce the closer you get to the bolt cross section core. On larger diameters the effect of higher strength surface layers diminishes in terms of overall tensile strength of the total bolt cross section. BS 4882 illustrates clearly the rapid drop off in tensile strength, particularly the 0.2% proof stress strength of B8X on bolt diameters in excess of 19mm diameter. This is especially significant for flange bolting where metallic/semi metallic gaskets are used. These gaskets require generally higher seating stresses and subsequent design bolt stress to seal. Reducing proof stress and true elastic limit potentially 30% below these tabulated values makes strength selection of B8 crucial. The situation becomes even more complex if hydraulic tensioner tightening is being considered for installation. The hydraulic overload that has to be applied to compensate for relaxation losses reduces the safety margin on usable elastic strength or even disqualifies this methodology as bolt yield could be exceeded.

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Where gasket seating stress and true elastic limit is an issue, precipitation hardening steels such as BS4882 B17 need to be considered. This alloy can boost its strength thru’ heat treatment so is a natural selection option for higher performance gasketted flanged joints. The similar ASTM designation is A450 660 grade. These materials also have a higher temperature capability up to 650/675. For even higher bolt temperatures, high nickel super alloys such as Nimonic 80 and Inco 718 provide high strength with creep and oxidisation resistance in these severe environments. The BS4882 categorisation for Nimonic bolting, is B80. For cryogenic applications beyond the capability of medium carbon alloy steels material selection mirrors that for high temperatures. Alloy selection is the same and the same limiting strength factors apply in terms of providing the required level of elastic strength enabling the bolt to deliver the design bolt tension that assures bolted joint reliability/zero leak performance. Service temperatures down to minus 200 – 250C are within these materials’ ranges for good strength and ductility/toughness. 7 Summary Material selection for any bolted application in terms of mechanical properties, operating environment and service life is straightforward. The complications start when budgeted cost does not correlate with technical/service specification. The Offshore industry is notorious for stating extended service life but then being totally unrealistic in the money it is prepared to spend on the fastener selected to achieve the requirement. Often one ends up with coated alloy steel bolts being used, rather than an inherent corrosion resistant bolt, against a 25 year life expectancy in the splash zone. Similar lack of realism occurs on petrochemical bolting exposed to high service temperatures and extended periods between planned outages. Often medium carbon low alloy steel is the final selection when precipitation hardening stainless bolts should have been used. Generally lower cost alloy steels can be used in the more hostile environments but planned maintenance / change out times will be shorter and more frequent. It’s all down to cost and subsequent in service risk.

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58

Traditional Flange Design

Methods Warren Brown, The Equity Engineering Group

© Warren Brown

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60

Traditional Flange Design Methods

Warren Brown, Ph.D., P.Eng. Principal Engineer, The Equity Engineering Group Shaker Heights, Ohio, USA Email: [email protected] Introduction

Early research in design and analysis of bolted joints was conducted in the 1920’s to 1940’s in

Germany, the UK and the USA. The findings of this early work led to flanged joint design rules being introduced by the American Society of Mechanical Engineers (ASME) in the 1940’s. The design method has remained largely unchanged since that time. Other international methods of design have been introduced recently, most notably the CEN EN-1591 method, however the ASME method remains the most widely accepted and most popular method of flange joint design. The method has given very good service across a wide variety of applications, and the fact that it has remained largely unchanged is testimony to its effectiveness. However, the method is not without issue and, while flange design issues represent a relatively small portion of the leakage that occurs in practice, there continues to be a number of failures associated with poor design.

In the engineering field, one often expects the latest methods to be the best and, in fact, that

they will eventually render the older, more traditional methods obsolete. However, in the case of flange design, the traditional ASME method still holds significant advantage over newer methods and, with minor alterations and improvements, the method can be modified to ensure an extremely high integrity joint design. Traditional ASME Method versus Other Codes

A comparison of current code design methods (noting that other additional international codes,

such as BS 55000 or AS1210, the Australian Pressure Vessel Code, are largely based on one of the listed methods) was performed for Welding Research Council bulletin 514 “Flange Design: Status of Present Rules”. The table outlines some of the key differences and similarities in the methods. One of the most common comparisons is between the EN1591 method, based on the TGL 32903/13, and the ASME method. Such comparisons usually highlighting the lack of a mechanical interaction calculation in the ASME method and the advantages that calculation offers when performed using the EN1591 method. However, the comparisons generally neglect one of the other key differences between the methods, which is the treatment of hub to flange and shell to hub interaction. For the most common flange design, weld neck flanges, the ASME method is based on the calculation of the shell, hub and flange ring as connected series, whereas the EN1591 method does not account for the shell restraint and approximates the effect of the hub with an equivalent increase of the flange ring moment of inertia. While this method is undoubtedly an acceptable and proven flange design method, this approximation means that some of the advanced methods that can now be applied to flange design, such as mechanical interaction, thermal effects and determining flange strength limits will not be as accurate or even possible to perform with the current EN1591 method.

In fact, recent work into determining the acceptable limits for assembly bolt load, to avoid damage to the flange, have shown that the inclusion of the effects of the hub and shell (both in terms of rigidity and stress locations) is essential. In addition, many of the advantages of the EN1591 method, such as the inclusion of mechanical interaction, can be relatively easily added to the ASME design method.

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Unfortunately, any such improvements to the ASME method are not likely to be included in the code updates in the near future, so it is advisable to step outside the standard design and analysis practices in order to improve on the traditional ASME design method. The following sections will outline the major areas of improvement that are required, whether there is a plan for inclusion of them in the ASME code eventually and what can be done in the interim to improve the existing method. Required Improvements to the Current ASME Design Method Inadequate Gasket Design Basis

One of the most significant areas of improvement that has received the bulk of the focus, in terms of research, over the past 20 years is the need to better determine operating limits for the gasket and apply that to flange design. The research effort commenced with the realization that there was no reliable standard method of determining the values of “m” and “y” for new gasket types that are not presently listed in Appendix 2. In addition to the need to determine the minimum stress required to seal the joint and the minimum seating stress for a given gasket, in flange design it would also be a significant advantage to know the bounds of application that are acceptable for a given gasket. These bounds include such aspects as the maximum permissible gasket stress (versus temperature) and the maximum permissible flange rotation (also versus temperature). Unfortunately, in spite of the level of research into these topics, there is presently no standard ASME or ASTM test methods that can be adopted for improving the ASME code and many of the current international test methods that have already been adopted do not adequately address the requirement or have inherent problems that make their application questionable.

In addition to the above mentioned improvements, the present “m” value used in the code accounts for the required gasket sealing stress during operation and part of the reduction in gasket load caused by pressure (which is why it is higher for stiffer gaskets). In most cases, the simple ASME code method using the “m” value will result in a conservative treatment of the effects of mechanical interaction in reducing the bolt load as pressure is applied. However, in some cases, and in particular for large diameter joints with stiffer gaskets, the simple method currently employed does not adequately cover the effects of mechanical interaction in reducing the gasket load over and above the amount of the hydrostatic end force. Joints with a larger diameter and stiffer gaskets will typically see a reduction in bolt load once pressure is applied and this means that the total gasket load lost is the sum of the hydrostatic end force and the bolt load loss. For those joints, there is risk that the current ASME method will provide a joint design that is prone to leakage. Disconnect Between Design and Operation

One of leading causes of joint leakage in the field is an inadequate initial assembly bolt load. In many cases, this can be directly traced back to disconnect between the bolt load used for flange design and the bolt load that must be applied in practice to achieve a leak free joint (typically in excess of double the design load). Many well meaning engineers have fallen into the trap of thinking that, because “code limits should never be exceeded”, the assembly bolt load should be limited to the code design bolt load. This invariably provides an excellent training lesson for the engineer in question when practically every joint leaks on start-up. There is no reason why the bolt load used for flange design must be so low, other than to meet the current expected norms of pressure vessel design for material stress limits.

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Inadequate Joint Design for Integrity Even though the majority of joints designed to the ASME code operate without incident, there

are categories of joints that have repeatedly proven to be susceptible to leakage due to design issues. These include large diameter, low pressure joints, refinery flanges with ≤19mm  (≤¾ inch) diameter bolts and higher pressure flanges with larger bolts (≥75mm, ≥3 inches). There is also a class of flanges that is more difficult to define, where the current design practices (including the now mandatory flange rotation limit) result in a flange that meets code design but will plastically deform and take on permanent flange rotation set at relatively low bolt loads (often at a load corresponding to 50% of bolt yield or less).

In the case of low pressure joints, one of the issues is that these fall into the category where the simple treatment of elastic interaction in the code is non-conservative and, on top of that, they often have inadequate bolting and flange strength to deliver sufficient gasket seating stress. Similarly, in the case of ≤¾ inch bolted flanges, there is most often insufficient bolt area available to provide adequate gasket stress for seating and/or operational considerations (resulting in the need to assemble the joints to in excess of 70% of bolt yield to achieve adequate gasket assembly stress levels). The larger diameter bolt flanges have the opposite problem; they have so much bolt area available that the gasket stress often exceeds twice the yield of the flange material, resulting in deformation of the flange, over-compression of the gasket, inadequate elastic rebound and subsequent leakage. The fix in this case is relatively simple; increase the gasket width to obtain a lower initial assembly stress. Unfortunately there is little that can be done for a flange that will yield prior to sufficient bolt load being applied to the gasket. One successful fix has been to strengthen the flange with additional backing-rings applied between the nuts and existing flanges, but these rings are relatively a poor solution, requiring an extremely thick backing ring to make any appreciable difference, when compared to identifying the issue and making the integral flange ring thicker at the design stage prior to fabrication. Inadequate Joint Design at Temperature

A common cause of joint leakage for high temperature, larger diameter flanges is the effect of thermal transients during operation. The bolt load can increase or decrease during operation due to variations in process conditions and such changes can cause joint leakage if sufficiently high. In addition, flanges that are relatively thin compared to the shell that they are attached to (flange thickness less than five times the shell thickness) are likely to lose a significant amount of bolt load during the initial stages of any high temperature process start-up, due to the shell forcing flange ring rotation. The effects of temperature can be accounted for, both at the design stage and later in the operation stage when selecting the appropriate assembly bolt load, by increasing the assembly bolt load over and above the stress required to seal the gasket, the expected relaxation stress and the loss in bolt load due to pressure and temperature. However, to allow for this at the design stage, it is necessary to determine the transient temperatures of the joint components and apply those temperatures to a mechanical interaction analysis to establish the associated change in bolt stress.

In addition to the transient effects of temperature on bolt load, there is also long term relaxation of the joint components. Due to micro-plasticity, relaxation of stress levels in materials occurs at temperatures much lower than normally expected for creep (above only 200°C (400°F) in carbon steel, for example). This means that, if not accounted for, there may be inadequate bolt load remaining to seal the joint for the expected life of the joint. Once again however, if the expected amount of component relaxation is known, then it is possible to select an assembly bolt stress level that will ensure sufficient long-term gasket stress exists during operation and thereby avoid joint leakage. Alternatively, it is possible to adjust the flange and/or bolt geometry and materials at the design stage to reduce the expected amount of creep/relaxation that will occur.

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Miscellaneous Improvements While the following items cause fewer leakage issues, they are relatively easily addressed at the

design stage, and therefore warrant inclusion in this section. In the present ASME VIII, Div. 1, Appendix 2 design method, there is no procedure outlined to address the effects of external bending moments or external forces during operation on the integrity of the joint. Once again, if this operational loading is quantified at the design stage, it is possible to strengthen the flange and select an appropriate assembly bolt load to ensure that leakage will not occur.

For lower pressure joints, and especially those with very thin gaskets, there is presently no limit in the Appendix 2 design method for flange bolt hole spacing. There are limits in other codes, such as ASME III and TEMA, but at present it is possible to design a flange that meets the ASME VIII code, but has bolt spacing that will result in regions of insufficient gasket stress between bolts, which may lead to leakage.

For slip-on flanges that are designed to the ASME code, there is a clause that allows them to be assessed as either integral (shell restrains the hub and the hub is assumed to taper over the hub height, like a weld neck flange) or loose (shell is not connected to the hub). Obviously the real case is neither of these and, in fact, using either of the methods can result in much higher stress levels at the shell to hub junction than for similarly designed weld neck flanges. Additionally, the flange rotation (and therefore mechanical interaction if calculated) will not be accurate due to the poor representation of the hub and/or the connection of the hub to the shell. Proposed ASME Code Revisions

As can be seen in the updated Table 1, the latest version of Appendix BFJ (the intended update

to ASME VIII, Div. 1 Appendix 2) includes most of the additional design improvements listed above in one form or another. The work is still at an early stage in many cases and requires some clarification and improvement prior to implementation, but at least the intent is there to make the improvements. Unfortunately, the fact that the basis for Appendix BFJ is leakage based design, means that there is little likelihood that it will be approved for publication in the near future and therefore the other improvements are being held from publication as a consequence. There is a significant amount of trepidation regarding the use of leakage based flange design among the ASME code community. Industry experience with the leakage based method present in Appendix BFJ is the converse of experience with the existing ASME code flange design method; one was rapidly installed and has remained relatively unchanged for over sixty years, while the other has been around for almost twenty years and has yet to gain any measurable acceptance within industry. The reasons for the lack of acceptance of the method are numerous, but unfortunately there has been little progress in addressing the issues, which undoubtedly points to significant underlying problems. Even the currently proposed path forward, to include the appendix as an optional non-mandatory requirement to the code, which would only be performed as a secondary check to the existing Appendix 2 design, is still unlikely to meet with success. Therefore, in the near term, designers and end-users will need to look to some of the following non-code methods outlined in order to improve ASME code joint integrity at the design stage. Non-Code Improvements to the ASME Design Method Inadequate Gasket Design Basis

Unfortunately, the first item off the list is one where there really is no good standardized solution to the problem. Individuals have had much success with the implementation of relatively simple gasket stress limits (a required seating stress, a minimum stress required during operation and a maximum permissible gasket stress & rotation). However the method of establishing these stress limits is non-standard and is usually a combination of both laboratory test results and field experience (as in Brown [1], for example).

64

The Pressure Vessel Research Council – Sealing Reliability Council is presently attempting to bring together current laboratory methods and end-user experience to establish suitable standard procedures for determining these values, however as of present none exist.

Equations to include the effects of mechanical interaction on bolt load have been available since just after the release of the present code method (Wesstrom, et. al. [2]). By incorporating the equations outlined in that paper, or one of the many subsequent papers written by others using this method, it is possible to accurately determine the effect of applied pressure on bolt load and, therefore, on residual gasket stress during operation. Disconnect Between Design and Operation

The issue of a design bolt load that is significantly less than the bolt load required to seal the joint is being addressed by post construction documents such as ASME PCC-1 Appendix O “Assembly Bolt Load Selection” [3], however it is good practice to think in terms of the actual assembly load when designing the joint. For example, the original ASME code method did not include assessment of the tangential stress at the hub to shell junction, because at the design bolt loads typically used, this stress is always smaller than the other regions (Waters, et. al. [4]). However, if the flanges are analyzed at normal assembly bolt load levels, then this stress can be significant and is one of the indicators of an inadequate flange design. In addition, when gasket stress limits are established by test, then these must be compared to actual expected bolt stress levels, rather than design stress levels. Therefore, it is generally necessary to either adjust the acceptable code bolt load and limits or perform a separate assessment after the code design assessment to account for component limits based on actual bolt load. Inadequate Joint Design for Integrity

The issues with large diameter low pressure joints are partially resolved by performing the aforementioned mechanical interaction analysis. The remainder of the issues for those joints and also for the excessively small and excessively large diameter bolt size joints are resolved by specifying an acceptable ratio of bolt to gasket area that must be met for flange design. For example, a suitable area ratio range for common graphite based gaskets used in refining (spiral wound with inner & outer rings, kamprofile and corrugated) is a gasket area divided by bolt area ratio of between 2.0 and 1.2, resulting in an assembly gasket stress of between 170 MPa and 240 MPa (25 ksi and 35 ksi) for an assembly bolt stress of 345 MPa (50 ksi). The gasket area should be based on full width for the perimeter portion and half width for the pass partition portion.

In addition to controlling the relative bolt and gasket area ratios, it is good practice to ensure that the flange is not the weak component in the joint. This ensures that it will not be possible to damage the flange during assembly by applying excessive bolt load and it also enables the full range of bolt stress to be used to seal the joint if it is required. Recent work in determining the maximum acceptable load that a flange will take, which formed the basis of ASME PCC-1 Appendix O, has established elastic assessment limits that give an indication of when the flange ring will undergo gross plastic deformation (have permanent rotational deformation). The work is summarized in a series of ASME PVP conference papers (Brown et. al. [5] to Brown [7]), however the limits used in the papers changed with time as the method developed and so an overall summary of the development is also planned to be published as Welding Research Council Bulletin 528. Using the equations and limits outlined in the papers, it is possible to determine both the flange strength and the location of the flange weakness, which can be used as a limit during design for the ensuring the flange is capable of taking, say, >80% of bolt yield. The method can also be used as a post-construction calculation for the upper limit on assembly bolt load for flanges designed without this minimum strength requirement.

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Inadequate Joint Design at Temperature The transient joint component temperatures and associated severity of mechanical effects of

temperature can be assessed, where appropriate, using the methods outlined in Brown [8]. However, this is generally only necessary where the temperature exceeds 200°C (400°F) for flanges ≤1500 mm (≤60 inches) in diameter and where the temperature exceeds 150°C (300°F) for larger diameter flanges. Additionally, assessment should be performed where significant difference in thermal expansion properties exist between the flange, shell or bolts.

A recent ASME-LLC project (Brown [9]) examined the long term characteristics of high temperature flange design. The conclusions of the project were that it was relatively simple to incorporate the effects of material creep/relaxation into the ASME code design process; however the material properties presently available for doing so are generally inadequate for wholesale inclusion of the method into the design process. What is required prior to the inclusion of these effects is that, at least for common materials, the relaxation characteristics of flange and bolt materials are established by extensive testing in a controlled environment. However, there is sufficient data available that the techniques outlined in the project report can be applied in a limited fashion. For example, an “order of magnitude” assessment of the effect on joint life to leakage of, say, higher initial bolt loads, or retightening the bolts during operation, or different magnitudes of bending moments on the joint is possible.

There has been significant research into the extent of gasket relaxation that may be expected, however only a small portion of it can be applied in practice. The existing standard tests do not provide long term relaxation results and so are not suitable for determining the amount of relaxation for a significant portion of the gasket types being employed in practice (graphite based gaskets, for example). Once again, individuals are finding success with using simple percentage relaxation values determined from laboratory tests and field experience that serve to, at the least, account for the majority of the effect of relaxation (Brown [1]). Miscellaneous Improvements

The Koves method was recently introduced into the 2007 version of ASME VIII, Div. 2 to account for the effect of bending moments on flange operation, and these equations can also be used when analyzing Div. 1 flanges.

There are a number of existing codes and references that offer guidance on maximum acceptable bolt spacing. These methods and a simple analytical equation for determining a design limit are outlined in Koves [10].

An extension of the original Waters [4] design method to include integral flanges with straight hubs (slip-on flanges that are welded to the shell) is outlined in Brown [11]. The paper provides alternate flange factors that may be used in the standard ASME code flange design method to accurately determine stress levels and flange rotation. Conclusions

It is the opinion of the author that the Traditional ASME design method remains the best option for designing and analyzing flanges. It is a sound foundation, based on a comprehensive assessment of the joint behavior that creates an excellent platform from which to improve in order to obtain leak free joint design.

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Table 1 – Comparison of Flange Design Methods (updated WRC Bulletin 514, Table 1)

Aspect of Flanged Joint Design

ASME VIII, div. 1, App 2

ASME VIII, div. 2, New Rules 1

ASME Append. BFJ 2

EN13445-3:2002, Sect. 11

EN1591:2001 3

Flange Design Basis Taylor-Forge

Taylor-Forge

Taylor-Forge

Taylor-Forge

TGL 32903/13

Includes effect of joint mechanical interaction Partial 4 Partial 4 Yes Partial 4 Yes Flange Stress Check Yes Yes Yes Yes Partial 5 Flange Rotation Check Yes Yes Yes Yes Check on Lap Joint Stub Shear Stress Yes Yes Yes Check on Lap Joint Stub Bearing Stress Yes Check on Lap Joint Stub Bending Stress Yes Yes Design of Seal Welded Joints Yes Gasket Loads based on “m” & “y” Yes Yes Yes Gasket Loads based on leakage Yes Yes Flange Allowable Stress Basis ST/3.5,

Sy/1.5 ST/2.4, Sy/1.5

TBD6 ST/2.4, 7 Sy/1.5

ST/2.4, Sy/1.5

Austenitic Allowable Stress Increase Allowed No Yes 8 TBD6 No Yes (?) Bolt Allowable Stress Basis ST/4, 9

Sy/1.5 ST/4, Sy/1.5

TBD6 ST/4, Sy/3 10

ST/2.4, Sy/1.5

Gasket Effective Width Basis Simplified Simplified Simplified Simplified Calculate Includes Gasket Creep/Relaxation Partial Partial 11 Includes Flange & Bolt Creep Included Effects of Temperature Partial12 Partial 12 Includes External Moments & Forces Yes Yes Yes Maximum Allowable Gasket Stress Yes Yes Maximum Spacing Between Bolts Yes Yes Effect of bolt holes on flange rigidity Yes Operational Flange Rotation Limits Adjusts for Assembly Accuracy Yes Yes Nubbins Prohibited Yes Table of standard bolt stress areas Yes

1 Based on the revision 7 of the document (current as of 1st January, 2006) 2 Based on the draft document dated February 15, 2006. Updates based on Dec 2009 document are detailed in red. 3 This is also the basis of EN13445-3 Appendix G and some listed aspects (flange stress limits for example) are taken from this appendix, as they are not specified in EN-1591. 4 It can be argued that the factor “m” accounts for the effects of mechanical interaction (ref. Brown [6]). 5 The stress check in EN1591 includes only a check of the circumferential stresses and flange is allowed to have plastic deformation (ref. EN1591, 1.3.4 b). The other methods include radial and tangential stress checks and use an elastic stress check. 6 The exact stress limits for BFJ are still a point of discussion 7 Note that due to experience with problems at the higher allowable stresses in large diameter joints, the allowable is reduced by a factor of 0.75 for ≥ 2000mm (78in.) diameter flanges. For diameters between 1000mm and 2000mm (39in. and 78in.) this reduction factor is taken to linearly vary from 0.75 to 1.0. 8 The basis for allowing higher allowable stresses for austenitic stainless is that the flange rotation limits should eliminate concerns regarding overly flexible flanges when designed to higher allowable. 9 Actually, the ASME II, Part D tables list allowable stresses for common materials that are closer to ST/5. 10 Note that the yield value for austenitic bolts is taken at an elongation of 1.0%, rather than 0.2%. 11 The effects of short term relaxation only are included in the present revision of EN13555. 12 Includes only axial expansion and does not detail how to determine temperature.

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References [1] Brown W., Ryan S., McKenzie, W., 2007, “Obtaining Leak-Free Bolted Joint Operation By Returning to Basics” National Petroleum Refiners Association Conference, Houston, Texas

[2] Wesstrom, D.B., Bergh, S.E., 1951, “Effect of Internal Pressure on Stresses and Strains in Bolted-Flange Connections”, Transactions of ASME, 73, n.5, pp 508-568, ASME, NY, USA

[3] ASME PCC-1 “Guidelines for Pressure Boundary Bolted Joint Assembly”, 2010, ASME NY, USA

[4] Waters, E.O., Rossheim, D.B., Wesstrom, D.B., Williams, F.S.G., 1949, “Development of General Formulas For Bolted Flanges”, Taylor-Forge & Pipe Works, Southfield, Michigan, Reprinted by the PVRC in 1979.

[5] Brown, W., Reeves, D., 2006, “Considerations for Selecting the Optimum Bolt Assembly Stress For Piping Flanges”, Proceedings of the ASME PVP 2006, ASME, Vancouver, Canada, PVP2006-ICPVT11-93094

[6] Brown, W., Reeves, D.., 2007, “An Update on Selecting the Optimum Bolt Assembly Stress For Piping Flanges”, Proceedings of the ASME PVP 2007, ASME, San Antonio, Texas, PVP2007-26649

[7] Brown, W., 2008, “Selecting the Optimum Bolt Assembly Stress: Influence of Flange Material on Flange Load Limit”, ASME PVP Conference, Chicago, IL, PVP2008-61709

[8] Brown, W., 2006, “Analysis of the Effects of Temperature on Bolted Joints”, Welding Research Council Bulletin 510

[9] Brown, W., 2010, “High Temperature Flange Design”, ASME-LLC, Project #3036, ASME, NY

[10] Koves, W.J., 2007, “Flange Joint Bolt Spacing Requirements”, Proceedings of the ASME PVP 2007, ASME, San Antonio, Texas, PVP2007-26089

[11] Brown, W., 2008, “Selecting the Optimum Bolt Assembly Stress – Flange Limitations: Flange Type”, Proceedings of the ASME PVP 2008, ASME, Chicago, Illinois, PVP2008-61708

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Overview of

Developments in EN 1591

Manfred Schaaf, AMTEC Services GmbH © Manfred Schaaf

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Manfred Schaaf

Overview of Developments in EN 1591

Messtechnischer Service GmbHHoher Steg 1374348 LauffenGermany

Bolted Flanged Joints: New Methods and Practices16-17 March 2010, Newcastle upon Tyne

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EN 1591 – Part 1 to 5 • Status quo• Latest developments• Future work items

Content

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CEN TC 74 – Flanges and their joints

Scope:

Standardization of flanges and their joints in pipelines and piping systems for all applications excluding hydraulic and pneumatic load transmission.

- General: Definition of "nominal pressure" and "nominal size";- Flanges: Definition of dimensions and tolerances, selection of

materials, technical conditions of delivery, P/T ratings;- Bolts, screws and nuts: Selection of required bolts, screws

and nuts, dimensions, technical conditions of delivery, materials;- Gaskets: Definition of dimensions and tolerances, materials,

technical conditions of delivery;- Calculation methods for flanges design;- Determination of P/T ratings.

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CEN TC 74 – Working Groups

G. TaylorCalculation MethodsCEN/TC 74/WG 10

J. HoyesGasketsCEN/TC 74/WG 8A. PerceboisCast iron flangesCEN/TC 74/WG 3H.-D. EngelhardtSteel flangesCEN/TC 74/WG 2

H. KockelmannFlanges and their jointsCEN TC 74

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EN 1591 Rules

EN 1591Flanges and their joints - Design rules

for gasketed circular flange connections

CEN/TS 1591-3

Calculation method"Metal-to-metal contact"

EN 1591-1

Calculation method

prCEN/TR 1591-5

Calculation method"Full face gaskets"

CEN/TS 1591-4

Qualification ofpersonnel competency

EN 1591-2

Gasket parameters

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EN 1591-1: Status quo

EN 1591-1 released as an European Standard in 2001

Amendment A1 of EN 1591-1 released as an European Standard in 2009

• leak tightness and strength criteria are satisfied

• behaviour of complete flanges-bolts-gasket systemis considered

Calculation method for gasketed circular flange connectionswith gaskets inside the bolt circle and without metal-to-metalcontact of the flange faces

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EN 1591-1: Treated parameters

gasket characteristics

medium pressure

strength value of flange and bolt materials

thermal loads

external axial forces and bending moments

possible scatter due to bolting-up procedurenominal bolt load

changes in gasket force due to deformation of all componentsinfluence of connected shell or pipe

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EN 1591-1: Specifics

flange rotation and effective compressed gasket areaelastic deformation balance

iterative determination of the required bolt force in assemblyto fulfill tightness demands

force balance(interaction between all components)

virtual flange resistance of the flangeslimit load theory(admissibility of plastic deformation)

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DIN 28090-1 EN13555 Gasket Characteristic Testing Equipment

σVU/L QMIN(L)Minimum gasket stress

in assembly for tightness class L

σBU/L QSMIN(L)Minimum gasket stress

in service for tightness class L

σVO QSMAX(RT) Maximum allowable gasket stressin assembly

σBO QSMAXMaximum allowable gasket stress

in service

ED EG modulus of elasticity

ΔhD PQR Creep-relaxation factor

TEMESfl.ai1

TEMESfl.relax

gasket characteristics(prEN 13555 - draft 2001)

E0, KI

gC

Amendment

EN 1591-1: Amendment A1

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DIN 28090-1 EN13555 Gasket Characteristic Testing Equipment

σVU/L QMIN(L)Minimum gasket stress

in assembly for tightness class L

σBU/L QSMIN(L)Minimum gasket stress

in service for tightness class L

σVO QSMAX(RT) Maximum allowable gasket stressin assembly

σBO QSMAXMaximum allowable gasket stress

in service

ED EG modulus of elasticity

ΔhD PQR Creep-relaxation factor

TEMESfl.ai1

TEMESfl.relax

gasket characteristics(EN 13555 – 2004) Amendment

EN 1591-1: Amendment A1

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EN 1591-1: Future work items

CEN/TC74 Resolution 275/2009

Allocation of Joint Working GroupCEN/TC 54/TC 69/TC 74/TC 267/TC 269/JWG "Harmonized standard solution for flange connections"

CEN/TC74 Resolution 282/2009

Corrigendum to EN 1591-1:2001+A1:2009-03

CEN/TC74 Resolution 2/2008

Preliminary Work Item"Sample calculation and guidance on interpretation of calculations presented in EN 1591-1"

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EN 1591-1: JWG

EN 1591-1 or EN 13445-3

Chapter 10EN 12516-2 CEN/TC 69Industrial valves

"…in accordance to European Standards."

Chapter 9.3EN 12953-3CEN/TC 269Shell and water-tube boilers

EN 1591-1 + tables with gasket parameters

Annex P

Taylor ForgeAnnex DEN 13480-3 CEN/TC 267Industrial piping and pipelines

new equations derivedfrom EN 1591-1:2001

Annex GA

Taylor ForgeChapter 11 EN 13445-3CEN/TC 54Unfired pressure vessles

new Amendment released-EN 1591-1+A1CEN/TC 74Flanges and their joints

RemarksChapter / AnnexENTC

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EN 1591-2: Status quo

EN 1591-2 released as an European Standard in 2008

supersedes ENV 1591-2:2001

• results of research project(PERL – Pressure Equipment, Reduction of Leak rate)

• gasket characteristics are listed for types of gasket materials

• the values are no minimum required values, but typical values("generic data")

The standads details gasket parameters for use in EN 1591-1 during prelimenary calculations

characteristics are only informative(gasket characteristics must be supplied by manufacturer;alternative source: www.gasketdata.org)

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EN 1591-2: Types of gasket materials

• Modified PTFE sheet• Proprietary PTFE / Graphite gasket with metal eyelet • Metal jacketed with graphite filler• Graphite Covered Metal Jacketed with graphite filler & outer ring• Serrated metal core [kammprofile] with graphite facing• Proprietary type of graphite faced kammprofile with secondary

metal to metal seal • Corrugated metal core with graphite facing• Graphite sheet with tanged stainless steel core• Graphite sheet with multiple thin metal insertions• Non-asbestos, fibre based sheet• PTFE filled spiral wound gasket with both outer and inner rings• Low stress graphite filled spiral wound gasket with outer and

inner rings• Graphite filled spiral wound gasket with outer ring• Graphite filled spiral wound gasket with outer and inner rings

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EN 1591-2: Example 1IM

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EN 1591-2: Example 2

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CEN/TS 1591-3: Status quo

CEN/TS 1591-3 released as an Technical Specificationin 2007

• leak tightness and strength criteria are satisfied

• behaviour of complete flanges-bolts-gasket systemis considered

Calculation method for metal-to-metal contact type flangedjoints based on EN 1591-1

rejected as EN, released as TS(no experience with the modified calculation algorithm)

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CEN/TS 1591-3: Specifics

Calculation in 4 steps:

• determination of the bolt tightening to reach the MMC

• determination of the bolt tightening to maintain the MMC inall the calculation situations

• check of the admissibility of the leak-rate

• check of the admissibility of the load ratio

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CEN/TS 1591-3: Future work item

CEN/TC74 Resolution 285/20009

Review of CEN/TS 1591-5:2007:Extension of the life of CEN/TS 1591-5:2007 for another3 years

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CEN/TS 1591-4: Status quo

CEN/TS 1591-4 released as an Technical Specificationin 2007

• design codes increasingly require controlled bolt tightening

• ensure personnel are competent to assemble and tightenbolted joints for a leak-free status throughout its´ service life

• training, experience and assessment of knowledge arerequired to achieve competency

Process for training and compentency assessment of personnel in the assembly of bolted flanged joints fitted to equipment subject to PED

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CEN/TS 1591-4: Specifics

• procedural framework must be included within operator´squality management system

• route for achieving comeptency in the skills

- classroom training and workshop practice- written test- period of monitored work site experience- assessment by a qualified assessor

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CEN/TS 1591-4: General knowledge

• the principles of bolt elongation (strain), bolt load and stress;• importance of applied and residual bolt loads; • bolt load loss and the implications;• effect of coefficient of friction on bolt load when using torque;• bolt tightening methods and their relative accuracies;• joint assembly methods and tightening procedures;• the requirements to meet a specific class of tightness;• flange, bolt and gasket types and their limitations;• functionality of gasket and seal;• factors affecting the degradation of bolted assemblies,

e.g. corrosion;• common causes of joint failure and leakage;• specific health or safety requirements associated with joint

components;• maintenance requirements of bolt tightening systems;• importance of certification and records.

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CEN/TS 1591-4: Specific knowledge• general health and safety precautions;• procedure for preparing a joint for closure;• identification of correct joint components;• seal face preparation;• gasket handling, preparation and installation;• functionality of clamp or engineered joints;• importance of alignment and gap uniformity;• importance of using the specified lubricant;• manual and hydraulic torque joint tightening;• joint tightening using hydraulic bolt tensioners;• techniques for measuring bolt strain;• confirming joint can return to service;• identifying defects or faults;• variance or irregularity reporting;• safe joint disassembly;• safety requirements when selecting and operating bolt

tightening tooling; • calibration of bolt tightening tooling;• recording bolted joint activity and maintenance of records.

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CEN/TS 1591-4: Future work item

CEN/TC74 Resolution 280/20009

New work item proposal:Conversion of CEN/TS 1591-4:2007-08 into an EN

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prCEN/TS 1591-5: Status quo

prCEN/TS 1591-3 under preparation in WG10PWI 00074056

• particular approach for full face gasketed joints

• leak tightness and strength criteria are satisfied

• behaviour of complete flanges-bolts-gasket systemis considered

Calculation method for full face gasketed joints based on EN 1591-1

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prCEN/TS 1591-5: Future work item

CEN/TC74 Resolution 280/20009

Activation of a new work item:preliminary work on project EN 1591-5 has reached a certain stage that a WI can be activated now.

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Contact Data

For more detailed information, please contact us:

Messtechnischer Service GmbHHoher Steg 1374348 LauffenGermanywww.amtec.deTel. +49 7133 9502-0E-Mail: [email protected]

84

Failure Mechanisms of

Bolted Joints - Bolting Aspects

Bill Eccles, Bolt Science Limited © Bolt Science Limited

85

86

Figure 1 Transverse joint movement

Failure Mechanisms of Bolted Joints – Bolting Aspects By Bill Eccles CEng BSc MIMechE, Bolt Science Limited SYNOPSIS The reliability of a flanged joint depends, in part, on the threaded fasteners that hold it together. Although threaded fasteners are generally considered a mature technology, significant problems exist with their use. The presentation briefly covers several failure modes of threaded fasteners including the problems arising from insufficient preload, self-loosening, tensile overload, fatigue and thread stripping. The presentation discusses some major accidents that have occurred as a direct consequence of particular failure modes. 1. INTRODUCTION It is known in principle how to design bolted joints in which bolting failures do not occur but in practice bolted related failures are not uncommon. Uncertainties about the applied forces, the magnitude of the preload achieved by the tightening process, inappropriate materials being specified and most notably, human error, in practice results in joint problems. On occasion such failures can have disastrous consequences. 2. INSUFFICIENT PRELOAD 2.1 Lack of Preload Flanged joints rely upon the preload, generated by the tightening of the bolts, to pre-stress the gasket so that a leak free seal is achieved and to resist the hydrostatic pressure tending to separate the flanges. The gasket relies upon the preload provided by the bolts to perform its sealing function effectively. Many leaks, which are frequently attributed to a gasket failure, are often as a result of insufficient clamp force provided by the bolts. This can be due to incorrect tightening or subsequent loosening following tightening. 2.2 Preload loss from gasket creep, bolt stress relaxation and self-loosening Bolts can lose preload without rotating. The loss of preload can be temporary; such as can occur as a result of differential thermal expansion, or permanent, for example from creep. There are several causes of non-rotational loosening, all of which involve either the bolt additionally elongating or the joint additionally compressing following installation. On flanged joints the issues commonly encountered are creep of the gasket material and stress relaxation of the bolts. Modern gasket material attempt to minimise creep. Stress relaxation can be mitigated by the appropriate choice of bolt material.

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Figure 2 A 'hard' joint

Figure 3 A 'soft' joint

Self-loosening is when the fastener rotates under the action of external loading. Flanged joints are largely exposed to axial loading. Although research indicates that some degree of slight loosening can result from axial loading, self-loosening of fasteners is usually as a result of transverse joint movement, illustrated in figure 1. Such transverse movement is undesirable for a flanged joint for several reasons. In the presentation a failure involving the self-loosening of nuts of a flanged joint on a pressure vessel containing an agitator assembly is discussed. 3. TENSILE OVERLOAD On conventional flanged joints the load increase experienced by the bolts can be significant. On a solid joint typical, the joint is relatively 'hard'. That is, the stiffness of the bolt is usually significantly lower than the joint stiffness. Figure 2 shows a joint diagram illustrating this condition. The proportion of the force that is applied to the joint which the bolt sustains depends upon the relative stiffness of the bolt to the clamped material. With a 'hard' joint, the bolt stiffness is low when compared with the stiffness of the joint. In such circumstances the increase in the bolt loading when an external force is applied to the joint is relatively small.

Conventional flanged joints have a relatively low stiffness due to the deflection of the flanges and compression of the gasket. This results in what can be termed a 'soft' joint which is illustrated in figure 3. In such a joint when an external force is applied, such as from hydrostatic pressure, the bolt can sustain a significant proportion of it.

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Figure 4 S/N Diagram

One consequence of this is that the bolt cannot be tightened near to yield since there is the risk that the bolt would be overloaded when the external load is applied. Typical target tensile prestress values for bolts used in flanged joints is 50% of the minimum yield strength. With a solid ('hard') joint, the target tensile prestress is more typically around 75% of the minimum yield strength. One consequence of this is that if the wrong bolt material is used on flanged joints it may only be revealed either during a pressure test or in service. On a solid joint, pre-stressed to a higher value, defective bolt material is more likely to fail at the time of assembly and hence more easily detectable. Mentioned in the presentation are details of an accident due to the bolts being overloaded during a pressure test on a flange. 4. FATIGUE FAILURE Fatigue is often quoted as the commonest reason for bolts to fail in service. It is well known that a part subjected to a varying load will fail at a significantly lower loading than one that has been statically loaded. Fatigue is a progressive cracking of a part under the action of alternating forces. Fatigue failure can take from thousands to millions of load cycles to occur, dependent upon the stress level in the part. It is well known that as the alternating stress increases, the number of cycles to failure decreases. This is represented by an S/N diagram as shown in figure 4. The S stands for stress and the N for the number of cycles. Most materials exhibit a knee in the S/N diagram. Beyond this knee failure will not occur no matter how great the number of cycles. The strength corresponding to this point is known as the endurance limit.

Possibly the most devastating engineering failure of 2009 occurred as a result of bolt fatigue at the Sayano–Shushenskaya hydroelectric power station in central Russia on the 17 August. The securing bolts on one of the turbine rotors failed resulting in water pressure lifting the 1650 tonne rotor into the turbine hall. This caused flooding of the turbine and engine rooms and a transformer explosion leading to the deaths of 75 people. A report released on the 21 December 2009 by a Russian parliamentary commission found that the failure was due to fatigue cracking in the 80 mm diameter bolts. Of the 80 bolts securing the turbine cover, at least 6 bolts had missing nuts and 41 had fatigue cracks. 5. THREAD STRIPPING Nut thickness standards have been drawn up on the basis that the bolt will always sustain tensile fracture before the nut will strip. If the bolt breaks on tightening, it is obvious that a replacement is required. Thread stripping tends to be gradual in nature. If the thread stripping mode can occur, assemblies may enter into service which are partially failed, this may have disastrous consequences. Hence, the potential of thread stripping of both the internal and external threads must be avoided if a reliable design is to be achieved. When specifying nuts and bolts it must always be ensured that the appropriate grade of nut is matched to the bolt grade.

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Figure 5 Thread stripping and bolt tensile fracture

In order to satisfy the above requirement when applied to tapped holes, the length of thread engagement required depends upon the relative strength of the threads. Rule of thumb is that when both male and female threads are of similar strength then a length of engagement equal to the diameter of the thread is usually required. For tapped holes in weaker materials longer lengths of engagements are needed - depending exactly of the relative strengths. One of the issues with thread stripping is that it is not obvious that it has occurred. Figure 5 illustrates what happen to the preload when thread stripping occurs. The nut stops in place but retains only a minimal preload.

To illustrate the possible consequences of thread stripping, mention in the presentation will be made of an accident that occurred on the USS Iwo Jima in the early 1990's. On October 30, 1990, the USS Iwo Jima experienced a catastrophic boiler accident whilst leaving Manama harbour in Bahrain. A valve failed resulting in large amounts of steam from both the ship's boilers being dumped into the boiler room. The valve controlled steam at a pressure of 40 bar and 450 C. All ten people that were in the room at the time of the accident were killed. The cause of the accident was attributed to the fitment of incorrect nuts.

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Seal failure from a gaskets

perspective Dene Halkyard, Flexitallic

© Flexitallic

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92

Seal failure from a gaskets perspective Dene Halkyard, Senior Applications Engineer, Flexitallic Ltd Seals fail not just gaskets is a wise and widely used adage in the industrial sealing industry. Seal failure is a phenomenon often attributable to a number of factors, of which the gasket is but one. From the gaskets perspective, creating and maintaining an adequate compressive force throughout the expected lifetime of the seal is paramount if containment losses are to be kept to an acceptable level. Most common failure modes can be characterised by insufficient, excessive or changes in gasket compressive force. Visual inspection of failed gaskets can reveal useful information about the failure mode and assist in preventing future leakage. Gasket failure attributable to insufficient and excessive compressive forces tends to occur during installation; whereas failure due to transient forces tends to occur under operational conditions. Correct gasket selection and adherence to established ‘best practice’ installation techniques play a major role in minimising emissions from bolted flanged connections.

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94

European Emissions

Legislation Dr Brian Ellis, European Sealing Association

© Dr Brian S. Ellis

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96

EuropeanEuropeanEmission LegislationEmission Legislation

Dr Brian S Ellis

Acronyms!Acronyms!•• ESA ESA ……....•• IPPC IPPC ……..

•• BAT BAT ……....•• BREF BREF ……•• IPPC IEFIPPC IEF

•• PEDPED

EEuropean uropean SSealing ealing AAssociationssociationIIntegrated ntegrated PPollution ollution PPrevention revention

and and CControl Directiveontrol DirectiveBBest est AAvailable vailable TTechniquesechniquesBBAT AT RReference noteseference notesIPPC IPPC IInformation nformation EExchange xchange

FForumorumPPressure ressure EEquipment quipment DDirectiveirective

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ContentsContents•• Development of European environmental legislationDevelopment of European environmental legislation

-- types of EU legislationtypes of EU legislation•• Key elements of European legislationKey elements of European legislation

-- CommunityCommunity--widewide-- nationalnational

•• IPPCIPPC-- Directive basicsDirective basics-- BATBAT-- BREF notesBREF notes

•• Current legislation developments Current legislation developments •• ESA contributionESA contribution

-- IPPC IEFIPPC IEF-- Sealing Technology BAT guidance noteSealing Technology BAT guidance note-- revision of PED?revision of PED?

•• ConclusionsConclusions

•• European Sealing AssociationEuropean Sealing Association•• Fugitive emissionsFugitive emissions

Development of European Development of European environmental legislationenvironmental legislation

Over 1000 pieces of environmental legislation have been adopted since 1967

0

20

40

60

80

100

120

140

67 68 71 73 75 77 79 81 83 85 87 89 91 93 95 97 99 1 3 5 7

Items adopted

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Development of European Development of European environmental legislation environmental legislation -- 22

European European Council of Council of EuropeanEuropeanCommissionCommission MinistersMinisters ParliamentParliament

MemberMemberStatesStates

Legislation proposed

Refinements proposed

Opinion sought

Refinements proposed

Types of EU legislationTypes of EU legislation

•• RegulationRegulation•• DirectiveDirective•• DecisionDecision•• RecommendationRecommendation•• OpinionOpinion

increasing control from the EUincreasing control from the EU

DirectivesDirectives the preferred tool for environmental policies; the preferred tool for environmental policies; -- overall objectives + strategies defined by EUoverall objectives + strategies defined by EU-- allows Member States flexibility to transpose into national legallows Member States flexibility to transpose into national legislationislation

-- binding and applicable directlybinding and applicable directly

-- binding, but flexible through transpositionbinding, but flexible through transposition

-- binding on those to whom it is addressedbinding on those to whom it is addressed

-- nonnon--bindingbinding

-- nonnon--bindingbinding

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Key legislationKey legislation•• Emissions from industrial plantsEmissions from industrial plants•• Solvent (VOC) emissionsSolvent (VOC) emissions•• National emission ceilingsNational emission ceilings•• Large combustion plantsLarge combustion plants•• Waste incinerationWaste incineration

•• TATA--LuftLuft (D)(D)•• Integrated Pollution Control (UK)Integrated Pollution Control (UK)•• VDI VDI –– various (D) various (D) –– ““guidelinesguidelines””

•• Integrated pollution prevention and controlIntegrated pollution prevention and control

EU and National legislation EU and National legislation -- 11

EU legislationEU legislation(Regulations, Directives etc)(Regulations, Directives etc)

UK D F E I etc ..

Directives are Directives are ““transposedtransposed”” into national legislationinto national legislation

European Commission / Parliament / Council of Ministers

EU Member States

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EU and National legislation EU and National legislation -- 22

IPPC DirectiveIPPC Directive

UK

IPC

D

TA-Luft

VDI guidelines

F E I etc ..

IPPC IPPC -- 11•• IIntegrated ntegrated PPollution ollution PPrevention and revention and CControl (ontrol (IPPCIPPC))

Directive 96/61Directive 96/61 adopted in 1996adopted in 1996•• compliance for compliance for newnew plants required by end October 1999plants required by end October 1999•• compliance for compliance for existingexisting plants by end October 2007plants by end October 2007•• frameworkframework measure measure -- provides for common EU emission provides for common EU emission

limits to be adopted subsequentlylimits to be adopted subsequently•• integrated approach for a potential pollutant across all media integrated approach for a potential pollutant across all media

which might be affected which might be affected

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IPPC IPPC -- 22•• applies to 6 categories of industry: applies to 6 categories of industry:

-- chemicalschemicals-- energyenergy-- production and processing of metalsproduction and processing of metals-- mineralsminerals-- waste managementwaste management-- ‘‘otherother’’

•• specific obligations on operators specific obligations on operators

-- take all appropriate preventative measures against pollutiontake all appropriate preventative measures against pollution-- ensure no significant pollution is caused ensure no significant pollution is caused -- avoid waste production avoid waste production -- recover waste produced or dispose of safely recover waste produced or dispose of safely -- use energy efficiently use energy efficiently -- take necessary measures to prevent accidentstake necessary measures to prevent accidents-- protect and clean up site upon cessation of industrial activityprotect and clean up site upon cessation of industrial activity

IPPC IPPC -- 33•• identifies certain priority polluting substances, including:identifies certain priority polluting substances, including:

-- arsenic and its compoundsarsenic and its compounds-- asbestosasbestos-- carbon monoxidecarbon monoxide-- chlorine, fluorine and their compoundschlorine, fluorine and their compounds-- cyanidescyanides-- metals and their compoundsmetals and their compounds-- nitrogen oxides and other nitrogen compoundsnitrogen oxides and other nitrogen compounds-- organoorgano--halogen compoundshalogen compounds-- organoorgano--phosphorus compoundsphosphorus compounds-- organoorgano--tin compoundstin compounds-- substances and preparations which are carcinogenic, mutagenic substances and preparations which are carcinogenic, mutagenic or which may affect reproductionor which may affect reproduction-- sulphur dioxide and other sulphur compoundssulphur dioxide and other sulphur compounds-- volatile organic compounds (volatile organic compounds (VOCVOC’’ss))

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IPPC IPPC -- 44•• each facility is subject to authorisation through permittingeach facility is subject to authorisation through permitting•• emission limit and permits based upon emission limit and permits based upon Best Available Best Available

TechniquesTechniques ((BATBAT))•• BAT must consider:BAT must consider:

-- economic and technical viabilityeconomic and technical viability-- use of lowuse of low--waste technologywaste technology-- use of less hazardous substancesuse of less hazardous substances-- improvements in recovery and recyclingimprovements in recovery and recycling-- consumption of raw materials and waterconsumption of raw materials and water-- energy efficiencyenergy efficiency-- technical characteristics of the installationtechnical characteristics of the installation-- geographical location geographical location -- local environmental conditions local environmental conditions

IPPC IPPC -- 55•• BAT interpretation will result in differences across EUBAT interpretation will result in differences across EU•• hence, requirement for exchange of information on national hence, requirement for exchange of information on national

assessments of BAT and emission limits assessments of BAT and emission limits •• provides the basis for the publication of provides the basis for the publication of BAT ReferenceBAT Reference

((BREFBREF) notes) notes•• European IPPC Bureau established to publish BREF notesEuropean IPPC Bureau established to publish BREF notes•• IPPC IPPC IInformation nformation EExchange xchange FForum (orum (IEFIEF) established to ) established to

develop and review BREF notesdevelop and review BREF notes

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BAT reference (BREF) notes BAT reference (BREF) notes -- 11•• for all industry sectors covered within IPPCfor all industry sectors covered within IPPC•• usually industryusually industry--specific (specific (““verticalvertical”” BREF BREF

notes)notes)•• some cover more than one industry sector some cover more than one industry sector

((““horizontalhorizontal”” BREF notes)BREF notes)

BAT reference (BREF) notes BAT reference (BREF) notes -- 22Large V

olume O

rganic Chem

ical Industry

Mineral oil and gas refineries

Large Volum

e Inorganic Chem

ical Industry

Pulp and Paper Industry

“Vertical” BREF notes

Emission monitoring

Energy efficiency

“Horizontal” BREF notes

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ESA contributionESA contribution

Large Volum

e Organic C

hemical Industry

Mineral oil and gas refineries

Large Volum

e Inorganic Chem

ical Industry

Pulp and Paper Industry

Emission monitoring

Energy efficiency

Sealing Technology

•• ESA participates in IPPC IEFESA participates in IPPC IEF•• In deference to European Commission, ESA document entitled, In deference to European Commission, ESA document entitled, ““ESA ESA

Sealing Technology BAT guidance noteSealing Technology BAT guidance note””

ESA Sealing Technology BAT ESA Sealing Technology BAT guidance noteguidance note

•• ESA participating in IPPC IEFESA participating in IPPC IEF•• sealing technology involved in most industries covered sealing technology involved in most industries covered

by IPPCby IPPC•• ESA developed own, horizontal ESA developed own, horizontal ““BREFBREF”” voluntarilyvoluntarily

•• Sections covering BAT for sealing:Sections covering BAT for sealing:•• bolted flange connectionsbolted flange connections•• rotodynamicrotodynamic equipmentequipment•• reciprocating shaftsreciprocating shafts•• valvesvalves

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Current legislation developmentsCurrent legislation developments

New New Industrial Emissions Directive (IED)Industrial Emissions Directive (IED)

IPPC Solvent (VOC)

emissions

Emissions from

industrial plants

Waste incineration

Large combustion

plants

etc …

Intention for original Directives to be superseded by new IED; Intention for original Directives to be superseded by new IED; -- original Directive will be withdrawn if ALL areas coveredoriginal Directive will be withdrawn if ALL areas covered-- parts of original Directive will remain if NOT covered by IEDparts of original Directive will remain if NOT covered by IED

ESA contribution ESA contribution -- 22

Specifically of relevance to Specifically of relevance to bolted flange bolted flange connectionsconnections::

•• ESA developing programme to revise PEDESA developing programme to revise PED•• aim to have bolted flange connections considered an aim to have bolted flange connections considered an

““essential featureessential feature””•• relevant CEN standards would be relevant CEN standards would be ““harmonisedharmonised””•• would encourage fitters / installers to be suitably would encourage fitters / installers to be suitably

qualified (similar to requirement for welders)qualified (similar to requirement for welders)

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ConclusionsConclusions•• Development of European emission legislationDevelopment of European emission legislation

•• Types of European legislationTypes of European legislation

•• Relationship between CommunityRelationship between Community--wide and national legislationwide and national legislation

•• Key elements of European emission legislationKey elements of European emission legislation

•• IPPC DirectiveIPPC Directive

•• Current developments in European emission legislationCurrent developments in European emission legislation

•• ESA Sealing Technology BAT guidance noteESA Sealing Technology BAT guidance note-- available for download from available for download from www.europeansealing.comwww.europeansealing.com

•• ESA developing programme to revise PEDESA developing programme to revise PED

www.europeansealing.com

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European Sealing Association European Sealing Association -- 11•• panpan--European trade organisationEuropean trade organisation•• established 1992established 1992•• nonnon--profitprofit--making trade associationmaking trade association•• 40+ Member Companies40+ Member Companies•• representing a strong majority of the fluid representing a strong majority of the fluid

sealing industry in Europesealing industry in Europe•• organised as series of productorganised as series of product--focussed focussed

DivisionsDivisions•• Working Groups for common activitiesWorking Groups for common activities

Elastomeric & Polymeric

Seals Division

European Sealing Association European Sealing Association -- 22

PackingsDivision

Flange Gaskets Division

ESA Members

Executive Committee

Mechanical Seals

Division

Expansion Joints

Division

Industrial Materials Working GroupIndustrial Materials Working Group

Safety, Environment and Efficiency Working GroupSafety, Environment and Efficiency Working Group

108

Sustainable developmentSustainable developmentIndustry must reduce its overall emissionsIndustry must reduce its overall emissions

A large proportion of A large proportion of emissions are those emissions are those anticipatedanticipated from industrial from industrial processesprocesses

Fugitive emissionsFugitive emissions

Some emissions occur Some emissions occur through unanticipated leaks through unanticipated leaks in process systems in process systems ……..

….usually referred to as “fugitive emissions”

Sealing technology is playing a major role in helping industry to reduce fugitive emissions

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Definition of Definition of ““fugitive emissionfugitive emission””Any chemical, or mixture of chemicals Any chemical, or mixture of chemicals ……

… in any physical form …

… which represents an unanticipated or spurious leak …

… from anywhere on an industrial site

Fugitive emissionsFugitive emissions-- the cost the cost ““IcebergIceberg””

•• Lost materialLost materialVisible costs

Invisible costs Labour to repair leaksMaterial to repair leaksWasted energyProcess inefficiencyEnvironmental clean upEnvironmental finesClaims for personal injuryLost sales due to poor image

Companies which invest to reduce their fugitive emissions can achieve a fast pay-back

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Tension Control, the key to

Bolted Flange Reliability Rod Corbett, James Walker Rotabolt

© Rod Corbett

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112

Tension Control, the key to Bolted Flange Reliability

Rod Corbett, Managing Director, JamesWalker Rotabolt There are three basic factors that ensure bolted flange joint reliability:- Joint Design Bolt/Component quality Achieving design bolt tension/joint compression/gasket seating stress on installation Measure and control all three and flange reliability is assured. The investment in managing design and quality assurance over the last twenty years has been substantial. The investment however in measuring and controlling installed bolt tension has been negligible despite the technical fact that the sole objective of any bolt used in tension is to deliver a known level of clamp force on the joint. The vast majority of flanged bolted joints are tightened in an uncontrolled manner i.e. the residual, installed bolt tension is unknown. This is remarkable when you consider that 90% of all bolted joint failures can be attributed to incorrect bolt tension. This is against a back drop of industry demanding greater levels of safety and reliability from its plant, equipment and structures. Millions are spent on controlling and measuring process parameters such as temperature, pressure, flow rates, speeds etc but it is reluctant to measure the parameter that holds all the pressure containment together – bolt tension. Maybe this is due to a lack of understanding as to the limitations of traditionally controlling tightening through tightening power or effort from torque or hydraulic tensioning. This is probably the case because most of the time the concluding reason as to why the joint has failed lies elsewhere from the installation – maybe with the gasket, the flange surface or the severe process thermal swing. Whatever the reason, design bolt tension objectives can be measured and controlled reliably and cost effectively. Operators who embrace this technology driven bolting route are inevitably rewarded with assured reliability – leak free performance on hydro test, start up and in service. This also results in lowest maintenance cost. The science is such that the Oil and Gas industry, upstream and downstream can realistically expect to eliminate all future bolted flange leaks by taking the technology driven route. The paper describes commercially available tension control systems along with their relative merits. Factors that effect the variations in these systems such as operating environment, temperature, operator skill, system datum face integrity and the crucial physical calibration of bolt extension versus bolt tension are discussed in detail. One state of the art, market leading system is described along with an explanation of the calibration methodology employed. Results of the systems independent test and accreditation programme outlines the systems overall integrity for industrial usage.

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114

Management of Integrity of

Bolted Joints for Pressurised Systems

Robert Noble, Hydratight © Robert Noble

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116

MANAGEMENT OF INTEGRITY OF BOLTED JOINTS FOR PRESSURISED SYSTEMS.

Robert Noble Technical Services Leader Hydratight

Comparison with the Welded Joint?Welded Joint

Coded Welder

Material Control

Documented Procedure

NDT Verification

Hydro-tested

Competent Personnel

Documented Procedure

Hydro-tested

Material Control

Integritytested

Bolted Joint

In Service InspectionRecords Records

Permanent joint Subject to Breakout

117

My Arms are

calibrated!

Just Nuts and Bolts!

This will seal it

TYPICAL RESULT

Flanged Joints – Are easy?Gasket not on compression stop

Gasket on compression stop

Flanges rotating due to over tightening

Green TagLeak Test Passed!

Would you be confident in the Would you be confident in the performance of this joint?performance of this joint?

Lubrication

?

Applying Integrity Management - new build

%Reduction in leaks 75% 75%

(Note: Total leak number includes all vendor leaks)

Phase 1No System

Total Joints - 8,691Total Leaks - 518Leak Rate - 5.9%

Phase 2JDMS Used

Total Joints - 5,413Total Leaks - 84

Leak Rate - 1.5%

Phase 3JDMS Used

Total Joints - 15,640Total Leaks - 234

Leak Rate - 1.49%

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Applying Integrity Management – Operational Major Operator Multiple Asset

% Leaks Year to Year

2.85%

0.52%

1.55%

0.70%

4.80%

4.30%

2.65%

0%

1%

2%

3%

4%

5%

6%

% Leaks 2002 % Leaks 2003 % Leaks 2004 % Leaks 2005 % Leaks 2006 % Leaks 2007 % Leaks 2008YTD

Management of Bolted Joints Evolution

HSE SAFETY NOTICE 2/2000GUIDELINES FOR THE

MANAGEMENT OF INTEGRITY OF BOLTED PIPE

JOINTS

2000 2002

GUIDELINES FOR THE MANAGEMENT OF

INTEGRITY OF BOLTED JOINTS FOR PRESSURISED

SYSTEMS

2007

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Management of Bolted Joints: Evolution

Ownership

“There should be an identified owner of the management system, responsible not only for its implementation and ongoing maintenance, but also for communicating its aims and objectives throughout the organisation. The owner should state the expectations for the system and monitor its effectiveness.”

Appoint a Champion

Support them with expertise

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Technology and Practice

• “Good practice with regard to selection and control of assembly, tightening and assurance of bolted joints should be applied. Understanding of the theory and practice of bolted joints and development of appropriate procedures should be encouraged throughout the organisation.”

Establish Standards

Ensure they are applied

Criticality Assessment

“The range of services, pressures and conditions which bolted joints experience varies considerably. Each joint should undergo a criticality assessment which will determine the levels of inspection, assembly control, tightening technique, testing, assurance and in-service inspection relevant to the joint.”

Leak PotentialLeak PotentialService FluidService Fluid

Loss PotentialLoss PotentialLocal factorsLocal factors

Criticality RatingCriticality RatingLowLow MedMed HighHigh

CompetenceCompetence MethodMethodWitnessWitness VerifyVerify

Integrity TestIntegrity Test InspectInspect

Assess

Determine

Control

121

Training and Competence

“Everyone with an influence on joint integrity in the organisation should be aware of the management system, its objectives, expectations and effects on project planning and day-to-day working. Good awareness needs to be maintained. Any staff working on bolted joints should be appropriately trained and competent.”

Records, Data Management and Tagging

“The certainty of achieving joint integrity increases if historical data exists on the activities carried out in the past, ideally from original construction of the joint, linked to the design specification of the joint. Providing and recording traceable data encourages best practice at the time of the activity, and will provide useful planning data for the next time the joint is disturbed.”

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In-service Inspection

In-service inspection of bolted joints is an integral activity to ensure the continued integrity of the joints and as such should be built in to all relevant inspection programmes. This section looks at the possible damage that can occur, the inspection methods available for detection of defects and mitigation measures that can be put in place to minimise such degradation.

Management of Leaks• “The objective of a

correctly designed and installed bolted joint is to provide a long-term tight seal and prevent ingress or egress of fluids through the joint. However, leaks can occur and managing the investigation and repair of the leak is essential to avoid recurrence. It can also provide useful data for prevention on other projects.”

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Analysis, Learning and Improvement

• Analysis of leakage and inspection data coupled with formal reviews of the management system should occur at agreed intervals by the owner and users. The results obtained from commissioning, incident analysis and in-service inspections should be used to generate ideas for continuous improvement.

• Easily monitored but meaningful performance standards should be put in place at launch to quantify the contribution being made by the management system and evaluate user satisfaction. Feedback on good practice in integrity issues and causes and solutions to incidents should be provided both internally and to industry to contribute to continuous improvement.

Collect Data

Analyse

Improve

0 100 200 300 400 500 600 700 8000

0.25

0.5

0.75

1Bolt Stress Relaxation from BS4882:1973 - Fig 9

100%

0

σ.res.B7

σ.res.B16

σ.res.B8

σ.res.B8M

8000 temp.B7 temp.B16, temp.B8, temp.B8M,

Summary

••A Management system is criticalA Management system is critical••Cover all of the elementsCover all of the elements••Appoint a championAppoint a champion••Apply Standards and ProceduresApply Standards and Procedures••Assess criticalityAssess criticality••Trained and competent people are keyTrained and competent people are key••Maintain a record and tagging systemMaintain a record and tagging system••Inspect joints and manage leaksInspect joints and manage leaks••Analyse and Improve. Analyse and Improve.

Copies of EI Guidelines available at Copies of EI Guidelines available at www.energypublishing.orgwww.energypublishing.org

124

ASME PCC-1 Updates

Warren Brown, The Equity Engineering Group © Warren Brown

125

126

ASME PCC-1 Updates

Warren Brown, Ph.D., P.Eng. Principal Engineer, The Equity Engineering Group Shaker Heights, Ohio, USA Email: [email protected] Introduction

The ASME Post-Construction Committee released the first version of ASME PCC-1 “Guidelines

for Pressure Boundary Bolted Joint Assembly” in 2000. At the time, the document was unique in addressing the issues with the assembly of bolted joints from a standards perspective. Since the initial version, there have been advances in gasket technology, bolting assembly procedures and calculation methods that enabled the improvement of both the integrity and efficiency associated with bolted joint assembly. In order to capture these advances, the ASME PCC-1 sub-committee was tasked to update the document beginning in 2006. The updates planned were extensive and have resulted in an increase in the length of the document from 33 pages to more than 80 pages. As evident from the almost three-fold increase in content, the updates are significant and are primarily in the form of additional new information, rather than modifications to the original information from the first version. This paper is intended to briefly summarize the major modifications to the document and, in the interests of length, will leave out many of the minor improvements also made. Please also keep in mind when reading both this paper and PCC-1 that PCC-1 is a guideline only. It represents what is considered to be best practice for the majority of joints in industry. However, it is not possible to cover all possible joint configurations within such a document, therefore the status as a guideline (only) is appropriate in that it leaves the possibility of modification based on specific need or experience up to the end user.

Changes to the Main Body of the Document

The most significant changes made to the main body of the document are outlined following:

In Section 4.0 “Cleaning and Examination of Flange and Fastener Contact Surfaces”, three changes were included, based on industry experience with best-practice and also from experience with joint failure. The wording was modified to allow graphite material to remain in the flange surface finish grooves after cleaning of the joint for inspection when using graphite faced gaskets. This modification was made in the interests of efficiency and based on extensive field experience indicating that graphite that remains in the facing grooves is time-consuming to remove and, if left in place, simply melds with the graphite facing on the new gasket to form a cohesive sealing element without degradation of the joint integrity. During the document public review phase, concern from several gasket manufacturers was expressed that it would be difficult to judge the amount of graphite remaining on the face and that excessive graphite may cover flange facing imperfections and/or affect the gasket sealing characteristics. However, the key to understanding why this will not occur is in the wording of the guideline; the only graphite allowed to remain is in the grooves of the surface finish and therefore sufficient quantity must be removed so as to allow flange facing inspection and the small amount left will not affect gasket performance.

127

The second change made to this section was the inclusion of a requirement to remove any flange paint or coating from the nut seating surfaces when the paint or coating thickness exceeds 0.13mm (0.005 inches). This requirement was based on industry experience with joint leakage in an offshore platform environment where the paint on standard flanges was excessively thick, led to additional bolt load relaxation and contributed to joint leakage. The thickness limit guidance was chosen to be an indication that a relatively thin layer of paint does not seem to affect joint performance (as most standard flanges are supplied with some form of protective coating), but that more than a thin layer is likely to lead to leakage and should be removed prior to joint assembly. The third change is guidance that the machining of large diameter bolts for reconditioning the threads is the preferred method. However, this will involve material removal and, therefore, a finite life for the bolt. Periodic replacement of the bolts should be planned if multiple reconditioning procedures are required on the same bolt.

In section 6.0 “Installation of Gasket”, commentary has been added to recommend that gaskets are not re-used. This inclusion was made based on field experience with joint leakage or flange facing damage where gaskets, in particular RTJ gaskets, are reused. Most gaskets are designed to plastically deform in order to obtain a seal. This results in a reused gasket being harder than a new gasket, which means that higher assembly bolt loads are required to obtain a seal, the gasket will not seal as effectively, and damage to the flange facing may occur during assembly. An exception to this recommendation is mentioned and that is the re-use of the metal core in grooved metal gaskets with soft facing (kamprofile gaskets). For these gaskets, it has been shown that it is possible to recondition them with a new facing layer and successfully reuse them in the same joint.

Table 1M and Table 1 were included in the first version of the document to be used as the basis for establishing the required assembly torque value by multiplying the listed torque value with the desired assembly bolt stress divided by the table reference bolt stress (345 MPa, 50ksi). However, it was common practice within industry to quote these tables as PCC-1 recommending 50ksi as an appropriate assembly bolt stress level. In fact, this was never the intent and so steps were taken in the revised document to clarify this. The steps included changes to the table titles to include the words “Reference Values for Calculating…”, some updates to the wording on how to apply the tables and also the inclusion of a new appendix, which outlines methods for determining the required assembly bolt stress.

Section 7.0 “Lubrication of Working Surfaces” was updated to include a recommendation that bolts be checked for free-running nuts during the bolt lubrication stage of assembly. This requirement was introduced based on field and laboratory experience which indicated that relatively small imperfections on the bolt or nut thread can have a significant impact on the obtained bolt load when tightening the joint using torque or tension techniques.

In section 13.0 “Joint Pressure and Tightness Testing”, a caution has been added with regard to the use of temporary gaskets during pressure and tightness testing (gaskets for which the joint was not designed). This caution is based on industry experience where temporary gaskets have blown out during pressure and tightness testing and caused personnel injury and fatality.

Appendix A: Training, Qualification and Certification of Joint Assembly Personnel

The lack of standardized qualifications for bolted joint assemblers has been identified as an issue by many in industry and is a leading cause of joint leakage due to poor assembly practices. In an effort to improve the status-quo, a significant revision to the existing PCC-1 Appendix A was drafted. The new appendix outlines the requirements for a certification entity to create and administer a training and assessment program for bolted joint assemblers that provides certification of the assembler.

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The appendix contains requirements for the minimum course content that must be taught in the theoretical portion, requirements for a series of practical demonstrations, a practical assembly exam that must be administered, requirements for maintenance of the certification and the requirements for the certification entity to establish and maintain their ASME accreditation in order to supply the certified assessment program. The appendix has three levels of assembler qualification: Certified Bolting Specialist, Certified Senior Bolting Specialist and Certified Bolting Specialist Instructor. Initial review of the available draft of PrEN/TS 1591-4 was conducted at the start of preparation of PCC-1 Appendix A and alignment was sought in overall format and context for the general requirements. In this way, it is hoped that the two certification requirements will be compatible in such a manner that it will be possible to have one training and assessment system that achieves both qualifications. One of the main differences between the two documents is that the training curriculum and practical demonstrations are outlined in greater detail in PCC-1 Appendix A.

The new version of Appendix A will not be issued with the main document when it is published in March 2010. This is due to the need approve and create the body within ASME that will administer the program once published. The appendix will be on hold until this has been done and will be released as an update via web page link to users of PCC-1 once everything is in place.

Appendix D: Guidelines for Allowable Gasket Contact Surface Flatness and Defect Depth

Previous industry guidelines for flange face flatness were based on manufacturing tolerances and often did not reflect what was practical to achieve in the field. The guidelines also did not address acceptable levels of minor local imperfection in the flange facing (pits, gouges and scratches). In addition, the acceptable imperfections in the flange facing are dependent on the type of gasket being employed. In terms of the flange flatness, which defines the amount of variation that will be seen in gasket compression, the new limits in PCC-1 were set based on the amount of compression that the gasket is subject to during assembly. Typical soft gaskets will compress in excess of 1mm (0.04 inches) and therefore are much more tolerant of flange face flatness variation than harder gasket types that compress much less than this amount. The amount of gasket compression stress lost due to flange flatness out-of-tolerance will be proportional to the variation divided by the gasket assembly deflection, so the tolerances specified in the appendix are varied depending on whether a hard or soft gasket is employed. The caution is also made that a soft gasket material (PTFE for example) may not exhibit soft behavior when applied as a thin gasket. The flatness tolerances are related to separate radial and circumferential acceptance limits and when these are combined the acceptable level of variation can be two to three times that of existing flange fabrication flatness guidelines.

A note is also made regarding the acceptability of complementary distortion of mating flanges, such as often occurs in shell and tube exchanger joints. For those, or similar joints, the orientation of the flanges is fixed by pass partitions or nozzle locations and it is possible to have thermally induced distortion on one flange that follows the other flange and does not therefore reduce the joint integrity. In those cases, it is acceptable to apply the flatness tolerances to the gap between the flanges, rather than for each flange independently. In addition, there is now a tolerance noted for the acceptable height difference for pass partitions on exchanger flanges to ensure both that it is not under or over compressing the gasket at that location. This requirement is based on experience where neglecting to specify this value leads to machining only of the periphery of the gasket, leaving the pass partition proud of the main seating surface, which often results in joint leakage.

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A second set of guidance is listed in the appendix for acceptable levels of local flange facing imperfections (pits, gouges, scratches,…). Once again, the acceptable levels are outlined relative to the gasket material. Harder facing materials (steel, for example) will not conform to the imperfection and will, therefore, be more sensitive to imperfections than gaskets that have a softer facing material. The limits include assessment of closely-spaced imperfections and have acceptable depth tolerances that are dependent on the type of gasket employed and the distance the imperfection extends radially across the flange seating surface. The intent is that these limits can be employed by an inspector to assess the flange facing condition as part of the standard equipment inspection process and only if the noted damage falls outside of the listed limitations will the joint be flagged for engineering inspection.

Appendix E: Flange Joint Alignment Guidelines

Previous flanged joint alignment guidelines were primarily obtained from fabrication specifications (ASME B31.3, for example) and did not address the fact that the initial alignment was not as critical as the inter-relationship between the initial alignment and the force required to bring the joint into perfect alignment (system stiffness). The alignment guidelines for PCC-1 were completely re-written to focus on geometry limits for alignment coupled with applied alignment force limits. The new limits address the maximum acceptable load to bring the joint into alignment in terms of the specified assembly bolt load. The acceptable load to bring the flanges parallel (angular misalignment) is listed as a maximum of 10% of the specified bolt load for any bolt. The maximum load to close an excessive axial gap between flanges is also a total of 10% of the specified bolt load, with a maximum individual load of 20% for any given bolt allowed for the combined limit. Simple figures illustrating the different types of misalignment have been added to clarify the listed tolerances. Additional considerations, such as the importance of joint alignment load on rotating equipment to avoid affecting shaft alignment and limits for when the assembler must seek engineering guidance if alignment forces are excessive are also included.

Appendix F: Alternative Flange Bolt Assembly Patterns

The original version of PCC-1 contained a bolt assembly pattern and procedure that involved tightening in a pattern pass at three different levels of assembly bolt load, completing a final circular pass and then an optional additional circular pass four hours afterwards. This method has been retained in the document for continuity and is referred to as the Legacy method. However, since the initial release of PCC-1, considerable effort in research has gone into proving that faster methods of assembly can be used that will achieve equal or better joint integrity. The theory behind these improvements is based on using an appropriate pattern for the gasket being employed and by increasing the bolt load at a much more rapid rate than the Legacy method. Increasing the bolt load more rapidly is applicable to all gasket types. It reduces the number of pattern passes required before proceeding to circular passes and generally results in a higher average gasket stress being achieved prior to commencing the circular passes. If the gasket stress is higher when the circular passes are commenced, the final compression on the gasket will be more uniform. The relationship between the gasket type and the assembly pattern is determined by how stiff the gasket is (how much compression occurs during assembly). For gaskets with relatively little compression (kamprofile gaskets for example) it has been proven that a pattern pass is not required and all that must be done is to tighten four opposing bolts in sequence to ensure that the joint has initial alignment prior to proceeding to tighten the remaining bolts in a circular fashion.

130

In addition, pattern passes using multiple tools have been included in the appendix in order to reflect this common industry practice. All of the new pattern passes do not include the optional final pass after a 4 hour wait and all include the additional instruction to continue tightening the bolts until they no longer turn for the final pass. There are three new patterns introduced for single tool application and two patterns for multi-tool. The single tool patterns include:

• Modified Legacy Pattern: Similar to the Legacy pattern, but with bolt load increased to the next level after every 4 bolts tightened, rather than after a full pattern pass. The pattern includes one or two pattern passes (second optional, depending on gasket type) and then a final circular pass until no nut turns.

• Quadrant Pattern: Similar in configuration to the Modified Legacy, except the bolts do not require numbering as, instead of using a cross-pattern for tightening the bolts, the joint is divided into quadrants and the next bolt in each quadrant is tightened in order. Bolt numbering is not required, as the next loose bolt in the next quadrant is always the bolt that must be tightened. Two patterns are presented, one for flanges with ≤ 16 bolts, where opposite quadrants are tightened successively and one for joints with > 16 bolts where the next quadrant in a circular order is tightened.

• Four-Bolt pattern: similar to the modified Legacy, except only four opposing bolts are tightened in sequence and then a circular pattern is commenced.

The multi-tool patterns are similar to the Modified Legacy pattern and the Four-Bolt pattern. In addition, the appendix contains guidelines for suitable measures for assessing the efficacy of other alternative tightening patterns/procedures that are not included in PCC-1.

Appendix M: Hardened Washer Usage Guideline and Purchase Specification

The existing specification often referenced for through-hardened washers is ASTM F436, which is actually a structural washer specification. That specification did not include higher alloy materials and the washer outer diameters were in excess of common flange spot-face diameters used at the nut contact surface. This resulted in the washer bridging the spot face, creating an undesirable bending of the washer during assembly. The new PCC-1 Appendix M was written with the intent to rectify these two issues and also to provide guidance on the service limits for the different materials listed for washer manufacture. The service limits are based on single use (where softening during operation will be acceptable, since they will not be reused) and multiple use (where softening is not desirable). The service temperature limits outlined in the appendix are based on metallurgical behavior for multiple usage and service experience for the single use limits. The four materials listed in the appendix are intended to match commonly applied bolt materials and significant effort was made to ensure that the washer thickness and material specification resulted in washers that could be easily manufactured. The intent is for this appendix to eventually be replaced by an ASTM specification, which is an effort that is already underway.

Appendix N: Reuse of Bolts

In many common joint sizes, it is practical to replace the bolting at every assembly in order to maximize the chances of joint integrity. However, there is often a cost barrier that prevents this from occurring. Appendix N has been written to ensure that more than cursory consideration of the bolt material cost is assessed when making the decision. The cost of the new bolting material is offset by the cost of reconditioning the old bolts and also the benefit to accuracy in achieved bolt preload with new bolts. Guidelines are given as to when to re-use and when to replace bolts. In addition, there is commentary on the appropriate methods for reconditioning bolts.

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Appendix O: Assembly Bolt Load Selection

This appendix outlines two methods of determining the appropriate assembly bolt load for a given joint. The first method is the use of a standard assembly bolt stress across all joints. It is recognized that the simplicity of that method may assist in its adoption and success on some sites. However, that method may also result in insufficient or excessive gasket stress or damage to the flange due to excessive bolt load in some cases. Therefore, the second method of determining assembly bolt load involves the calculation of the maximum limits for each component and the determination of the minimum required gasket stress to both seat the gasket during assembly and to seal the gasket during operation. Once gasket relaxation and hydrostatic end force have been allowed for in the calculation, there is a band within which the assembly bolt load may be selected that will ensure that no joint components will be damaged and that sufficient gasket stress is present during all phases of operation such that no leakage will occur. Using this comprehensive approach allows the end user to be more aware of the reasons as to why the selected bolt load is being applied and therefore to explore opportunities to improve the joint integrity based on the limiting factors for the joint, as determined by calculation. The appendix contains tabulated values of maximum allowable assembly bolt stress to avoid damage to the flange for standard B16.5 and B16.47 “Series A” flanges in sizes from DN 50mm (NPS 2) to DN 1200mm (NPS 48). A worked example for determining the assembly bolt load for a DN 75mm (NPS 3), cl. 300, flange and an example assembly bolt torque table is also provided.

Appendix P: Guidance on Troubleshooting Flanged Joint Leakage Incidents

One of the most important activities that can be undertaken in any leak free bolted joint program is a diagnosis of the cause of any leaks that occur. This includes an assessment of what the original joint configuration was, assembly history, operating conditions and condition of the joint and gasket subsequent to joint disassembly. The new Appendix P in PCC-1 provides guidance and a series of checklists designed to guide the user through an investigation of joint leakage. It contains a sample “Flanged Joint Leak Report” and additional lists of considerations for common flange design issues and some potential resolutions for those issues. It also lists some best practice guidance for basic flanged joint design problems. The diagnostic troubleshooting checklists are written to key from when leakage occurred and to narrow in on conditions and clues as to why the leakage occurred.

132

Conclusions

The ASME PCC-1:2010 version represents a step change in the level of detail provided for guidance on bolted joint assembly and will represent a significant body of work for the international improvement of bolted flanged joint integrity.

The undertaking and commitment by the committee members (listed following) was significant; however it is believed that the benefit to industry from this revision will be commensurate.

Chair:

Mr. Clyde Neely (Becht Engineering Co., Inc.)

Members:

Mr. Joseph Barron (Northrup Grumman Newport News)

Dr. Warren Brown (Equity Engineering Group)

Mr. Edward Hayman (Superior Plant Services)

Mr. David Lay (Hytorc)

Mr. Gary Milne (Hydratight)

Mr. James Payne (JPAC, Inc.)

Mr. Clay Rodery (BP North American Products, Inc.)

Mr. Jerry Waterland (Virginia Sealing Products, Inc.)

133

134

Qualification of Personnel

Competency – DD CEN/TS 1591-4

John Hoyes, Flexitallic Ltd © J. R. Hoyes of Flexitallic

135

136

The Evolution of a Pan-European Norm on Competency Assurance of Flange Assembly Technicians

John HoyesFlexitallic

Sections of Presentation

Background Considerations

CEN Standardisation

Harmonisation with PED

137

Background Considerations

Joints Fail – Not Just Gaskets

Installation critically important

138

Objective

To Raise the Status, in the context of the PED , of a Joint

Assembly Technician to that of a Welder responsible for the welds

of the flanges being sealed

Loss of Time Served Maintenance Personnel

Progression Towards Contractors

Previous Knowledge & Experience Base Lost

139

HSE Concerns Over Safety Record In North Sea

Attendance at a Training Course Does Not Demonstrate

Subsequent Competency

140

Competency Assessment Systems Added to Training

Courses for North Sea Technicians

Outcome was a Significant Reduction in Incident Rate

CEN Standardisation

141

TC 74 “Flanges and their Joints”set up to Implement the

Requirements of the Pressure Equipment Directive

Chairman : Hans Kocklemann of MPA , Stuttgart

TC 74 WG 10 “Calculation Methods”

Convenor, Robert Noble, Hydratight

TC 74 WG 8 “Gaskets”Convenor, John Hoyes, Flexitallic

142

Competency Document Drafted by Hydratight Member of WG 8

based upon the North Sea Experience

TS 1591 Part 4 : 2007

“Qualification of Personnel Competency in the Assembly of

Bolted Joints Fitted to EquipmentSubject to the

Pressure Equipment Directive”

TS --- Technical Specification

143

A TS Has A Lower Status Than an EN [European Standard]

A TS is Intended to be a Pre-Standard that Leads Within Three

to Five Years to a full EN

Before Approval an EN is Subject to Public Enquiry, twice, and a

Weighted Formal Vote

Adoption of a published EN is not Mandatory

144

TS 1591 Part 4

Intended to be an Umbrella Document Augmenting Current Training Schemes by Adding

Competency Assessment

TS 1591 Part 4 : 2007For the sections on General and Specific

Knowledge Curriculum Requirements of a Training Course only the suggested subjects

to be covered are listed

There is a Requirement for Work Site Experience with Guidance by Certified Competent Person and Log Keeping

A Competency Assessment has to be carried out once candidate has had Sufficient Work

Site Experience

145

Guidance for Work Site Experience before Competency

Assessment

12 MonthsSporadic

6 MonthsInfrequent but with Intense Periods

3 MonthsFrequent & IntenseEarliest AssessmentWork Site Experience

Method of Competency Assessment

Theoretical Question Paper

Practical Assessment during typical Simulated on site assembly

Documented Work Place Evidence

146

Refresher Training Guidance

12 MonthsSporadic

2 YearsInfrequent but with Intense Periods

3 YearsFrequent and Intense

Period After Achieving Competency

Work Site Experience

Decision Taken by TC 74 to Upgrade TS 1591 Part 4 to be a

Full EN Standard

This follows both the natural intended path for a TS and the German chemical

industry view that a full EN is more likely to be adopted

147

Harmonisation with PED

Bolted Connections are not recognised as an “Essential

Feature” of the PED

This should be changed

Then EN 1591 Part 4 would have to be Harmonised with the

Requirements of the PED & thus create further encouragement for its

adoption

148

Perhaps Operators would be able to achieve insurance cost reductions by specifying only competent, as defined by 1591 Part 4, technicians were used

on site

THANK YOU

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150

A regulatory perspective on bolted joints at high hazard

sites Iain Paterson, HSE Offshore Division

© Health & Safety Executive

151

152

A Regulatory Perspective on Bolted Joints at High Hazard Sites Iain Paterson, Team Leader Mechanical Engineering, HSE Offshore Division The presentation addresses:

• the relevant safety legislation aimed at ensuring the integrity of pressure systems both offshore and onshore,

• a few photographs showing what we find in the ‘real world’, • a few statistics showing the equipment where hydrocarbon releases occur

offshore, and • Some of the benchmarks that HSE uses to help judge compliance.

We need safety legislation to prevent catastrophic events such as the Piper Alpha disaster in 1988 where 167 lives were lost. Lord Cullen’s enquiry into the disaster led to many wide ranging recommendations including changes in the offshore safety legislation. The principles embedded in the Cullen report have stood the test of time but we still need to be vigilant. In 2005, a major accident incident occurred at the Bombay High complex in which 350 of the 367 persons on board the platform survived. That’s a testament to Lord Cullen’s recommendations and the developments in major hazard accident prevention and mitigation since Piper Alpha. One of the principle recommendations arising from Lord Cullen’s enquiry is the Offshore Installations (Safety Case) Regulations. Regulation 12 specifies the central theme. All major accident hazards must be identified, and all major accident risks must be evaluated and controlled. Major accidents include fire and explosion, and major damage to the structure that affects its stability. The essence of this regulation is that duty holders must have a robust safety management system and an effective auditing regime to ensure, amongst other things, that the integrity management of the hydrocarbon containment envelope is maintained. Onshore, the Control of Major Accident Hazards Regulations (COMAH) applies to sites such as refineries and chemical works etc where significant inventories of hazardous material are used. Regulation 4 requires the operator to take all measures necessary to prevent major accidents and to limit their consequence on the local population and environment. The effect is the same as for offshore, the duty holder needs to put in place a robust safety management system and an effective audit function. The COMAH regulations define major accidents as major emissions, fires and explosions that could lead to serious danger to human health or the environment. Regulation 5 of the Management of Health and Safety at Work Regulations (MHSWR) applies at all work places and requires employers to implement an effective safety management system commensurate with the risks that they create.

153

The Pressure Equipment Regulations 1999 address the design and initial integrity of new plant both onshore and on fixed offshore installations. Examples are given in the presentation showing poor practice on new equipment including; • missing flange bolts, • tack welded vibration supports • Unsuitable material used on a pipe support pad. The Provision and Use of Work Equipment Regulations (PUWER) address in-service integrity and apply both onshore and offshore. Regulation 6 is relevant in that it address the inspection of piping and flanged joints etc. It requires that deterioration such as corrosion is detected in good time so as to allow remedial action before a dangerous situation occurs. In practice, this means an inspection scheme where someone competent has considered the anticipated deterioration modes. It means adopting suitable inspection techniques where you have confidence in detecting deterioration. In other words an inspection regime that considers the scope, the nature and the frequency of inspections. COMAH and the Offshore Installations (Safety Case) regulations both require a safety management system to ensure that this actually takes place together with periodic review and audit to confirm or otherwise, that the inspection regime remains valid. Examples are given in the presentation showing in-service deterioration including;

• a perforated gas pipe suffering from corrosion under insulation, • rapid erosion of a vessel by a sand wash nozzle, • instrument tubing susceptible to vibration induced fatigue failure, • a flanged joint fretting against another pipe • Galvanic corrosion due to dissimilar metals on bolts and flange.

In the offshore sector, industry and HSE are working to reduce the number of hydrocarbon leaks. Duty holders formally report all of their leaks to HSE and these are stored in a database. HSE research report RR672 summarises the statistics from HSE’s offshore hydrocarbon release database. Over the eight year period 2001 to 2008, there were a total of 579 ‘major’ and ‘significant’ hydrocarbon releases, decreasing from 110 such releases in 2001 to 60 in 2008. RR672 indicates that major and significant leaks occur most often at: piping (21%), instruments (18%), and flanged joints (10%). However, it’s difficult to pin point exactly what proportion of hydrocarbon leaks occur at flanges. A study looking at gas leaks greater than 25 kg (a substantial release that would have serious implications if ignited) revealed that instruments, piping, flanges and valves are the priority areas where industry and the regulator need to focus our attention. HSE uses evidence such as this to inform our inspection priorities. Typically, HSE interventions to inspect the integrity of the hydrocarbon containment plant are based on our loss of containment manual that is publically available on our web site. It addresses several key risk areas including bolted joints, leaks from small bore fittings, and vibration induced fatigue failure of small bore piping connections. For bolted joints, we use the Energy Institute guidelines as a model of good practice. Bolted joints can be safety critical parts of the high hazard process plant and that their integrity must be effectively managed throughout their life time.

154

155

156

Health and Safety Executive

A regulatory perspective on bolted joints at high hazard sites

Iain Paterson CEng MIMechE

Team Leader, Mechanical Engineering

HSE, Offshore Division

Why do we need safety legislation?

To provide adequate integrity management of high hazard plant

Piper Alpha – 6th July 1988 – 167 lives lost

157

Why do we need safety legislation?

Relevant legislation includes …

Offshore Installations (Safety Case) Reg’s 2005Regulation 12: Management of health and safety and control of major accident hazards(1) ……..include in the safety case sufficient particulars to demonstrate that …..(c) all hazards with the potential to cause a major accident have been identified; and(d) all major accident risks have been evaluated and measures have been, or will be, taken to control those risksto ensure that the relevant statutory provisions will be complied with.This means put in place a robust safety management system and effective audit

158

Relevant legislation includes …

Control of Major Accident Hazards Regulations 1999Regulation 4 General dutyEvery operator shall take all measures necessary to prevent

major accidents and limit their consequences to persons and the environment.

This means put in place a robust safety management system and effective audit

Relevant legislation includes …

Management of Health & Safety at Work Reg’s 1999Regulation 5

Every employer shall make and give effect to such arrangements as are appropriate, having regard to the nature of his activities and the size of his undertaking, for the effective planning, organisation, control, monitoring and review of the preventive and protective measures.

This means put in place a robust safety management system and effective audit

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Initial integrity …

Pressure Equipment Regulations 1999 (applies to new equipment onshore and offshore)

Missing bolts

Initial integrity …?

160

In-service integrity …

Provision and Use of Work Equipment Reg’s 1998Regulation 6: Inspection (2) Every employer shall ensure that work equipment exposed to conditions causing deterioration which is liable to result in dangerous situations is inspected -(a) at suitable intervals; and(b) each time that exceptional circumstances which are liable to jeopardise the safety of the work equipment have occurred,to ensure that health and safety conditions are maintained and that any deterioration can be detected and remedied in good time.

In-service integrity …?

161

Bolted flanged joint integrity …?

Bolted flanged joint integrity …?

162

Offshore hydrocarbon releases

595 major and significant leaks in 8 year period

The 3 most common types of equipment where offshore hydrocarbon leaks occur:

Piping 126/595 21%

Instruments 107/595 18%

Flanges 59/595 10%

163

Major & Significant Gas HCRs > 25kg by Equipment Type

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2008-92007-82006-72005-62004-52003-42002-32001-2

Flanges

www.hse.gov.uk/research/rr672

HSE Loss of containment manual www.hse.gov.uk/offshore/lossofcontainmen.pdf

UKOOA Hydrocarbon release reduction toolkitwww.stepchangeinsafety.net/ResourceFiles/toolkit%20final%20version.pdf

Energy Institute Guidelines for the management of the integrity of bolted joints for pressurised systems

Energy Institute document Guidelines for corrosion management in oil and gas production and processing

HSE Offshore external corrosion guidewww.hse.gov.uk/offshore/corrosion.pdf

Benchmarks:

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Offshore Installations (Safety Case) Regulations 2005

www.hse.gov.uk/pubns/priced/l30.pdf

COMAH 1999

www.hse.gov.uk/pubns/priced/l111.pdf

Management of Health & Safety at Work Regulations 1999

www.hse.gov.uk/pubns/priced/l21.pdf

PUWER 1998

www.hse.gov.uk/pubns/priced/puwer.pdf

Pressure Equipment Regulations 1999

www.berr.gov.uk/files/file11284.pdf

Legislation downloads:

165

166

Leak Management

Ed Versluis, James Walker Rotabolt © James Walker

167

168

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Achieving:

Ed VersluisSales Manager

James Walker Benelux

Maximum bolt forceMaximum

bolt force

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169

Q8 - HDS unit:- Friday 13th February, 2009

Some are very costly…

Costs ofSteam Quenching...

170

How effective is leak sealing?

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85%of flange leaksare caused byincorrect bolt loads

WhyWhy……does everyone blame thegasket?...?...

172

Gaskets don’t fail !...

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3 Keysto Reliable Bolted Joints

Component

Quality

Join

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BoltTension

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173

Bolt Force

HydrostaticForce

OperatingGasket Stress

Possibleflange

bending or"rotation"

174

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CNAF Gasket 1.5 mm thick, 20 bar

0.01

0.1

1

10

0.000 10.000 20.000 30.000 40.000 50.000 60.000

Operating Stress MPa

Leak

Rat

e m

g/se

c/m

Leak Rate

Torque Tightening

Friction Estimate

175

Tightening Case History:H.P. Heat Exchanger

•Studs: 36 x M52 x 360mm long, on 1720mm PCD

•Grade: ASTM A193 B7

•Duty: 30bar, 350ºC

• Calculated bolt load: 412 kN / 42 Tonnes

• Calculated torque (� = 0.2): 4368 Nm

• Min. torque: 2259 Nm (- 48.3%)

• Max. torque: 5874 Nm (+ 34.5%)(Thread and nut surface – lubricated)

Tightening Case History:H.P. Heat Exchanger

176

Torque Scatter

0

1000

2000

3000

4000

5000

6000

7000

1 3 5 7 9 11 13 15 17 19 21 23 25 27 29 31 33 35Bolt No.

Torq

ue (N

m)

Bolt #Torque [Nm]Calculated torque

Theoretical vs. Actual Torque

-48%

+34%

Hydraulic Tensioning

177

Hydraulic Tensioning

Load-Transfer Relaxation

1. Loading 2. Localised deformation 3. Distribution 4. Load losses

Step 2Step 1 Step 3 Step 4

Limited access

3 Bolts tightenedby “ flogging ”

178

© James Walker 2006

Traditional bolting

Minimum Bolt force

0 kN

F [kN]

Required bolt force

Traditional bolting (without tension control)

Bolt force

+/- 40%F min F max

Tem

pera

ture

Pres

sure

Tem

p.

Flanges Bolts

DestructionZone

LeakageZone

Maximum bolt force

© James Walker 2006

Traditional bolting

Minimum Bolt force

0 kN

F [kN]

Required bolt force

Traditional bolting (Thermal cycling)

Bolt force

+/- 40%F min F max

Tem

pera

ture

Pres

sure

Tem

p.

Flanges Bolts

Proces upsetDestruction

ZoneLeakage

Zone

Maximum bolt force

179

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© James Walker 2006

JW RotaBolt Tension control & JW gasket science

Minimum Bolt force

0 kN

F [kN]

IdealBolt force

+/- 5%

DestructionZone

LeakageZone Bolt tension

Maximum bolt force

180

© James Walker 2006

Minimum Bolt force

0 kN

F [kN]Maximum bolt force

Bolt tension

+/- 5%

Tem

pera

ture

Pres

sure

DestructionZone

LeakageZone

Tem

p.

Flanges Bolts

JW RotaBolt Tension control & JW gasket science

© James Walker 2006

Minimum Bolt force

0 kN

F [kN]Maximum gasket,

flange- orbolt force

+/- 5%

Tem

pera

ture

Pres

sure

Safety Margin

Tem

p.

Flanges Bolts

No leaksguaranteed

Proces upsetDestruction

ZoneLeakage

Zone

JW RotaBolt Tension control & JW gasket science

Bolt tension

181

RefineryEXAMPLES2

3 Case Histories:

a) Valves in Catalytic Reforming

b) Cat. Reforming Heat Exchanger

182

Powerformer

Powerformer naphtha

From Storage

183

Unit had a track record of leaks at RCV’s since 1958

RCV flanges

184

Naptha 550°C / 37 bar

Thermal cycling

12” & 16” Class 600

Silver FacedKammprofiles

20 RotaBolts

Steam quench on 40+RCV flanges in 2002

All flanges 7 years leak free

GoodEngineering

Practice

GoodEngineering

Practice

185

© James Walker 2006

Heat exchanger leaks (how to avoid?)

Tubesheet

255255°° CC 425425°° CC

��T over tubesheet = 170T over tubesheet = 170°° CC

186

�1.7/8” bolts(B16 grade)

�approx 20”clamp length

�10.6:1 ratio

�30.3 tonnestarget load

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Actually took around

60 tonnes per bolt

to set the tension…

187

© James Walker 2006

© James Walker 2010

188