design of compact flange joints

PVP2002-1087

1 Copyright © by ASME

ASME-PVP 2002, Bolted Flange Connections

August 4-8, 2002, Vancover, British Colombia, Canada

DESIGN OF COMPACT FLANGE JOINTS

Finn Kirkemo Seaflex a.s. P.O.Box 451

N-1373 Asker, Norway Phone: +47 66 76 16 58

Fax: +47 66 76 16 30 E-mail: [email protected]

ABSTRACT In the past 10 - 15 years, flange joints designed for metal-

to-metal face contact with self seating and pressure activated seal rings have been used extensively in high-pressure applications in industrial piping, pressure vessels, pipelines, risers and associated equipment. These flange joints are generally much smaller and lighter, with smaller bolts, than equally rated standard gasketed flange joints, and are often called compact flange joints. This paper provides all necessary information to design compact flange joints for pressure and external loads and made from any suitable material. The paper includes design methods for design of the seal ring, flange and bolts in addition to assembly guidelines. Weld neck flanges, where the hub is of uniform thickness are discussed in detail. Similar method as presented has been applied to design compact flange joints with great success for many years. INTRODUCTION

Design codes typically recommend the use of standard flanges, e.g. ASME B16.5, wherever possible. This recommendation is based on proven safety and that a standard flange usually will be less expensive than a special one. However, due to leakage problems within some applications and due their large sizes, the development of improved flange designs with higher leakage reliability and smaller sizes and hence lowers costs have emerged. Flanged joints designed for metal-to-metal face contact is one example of such a joint. Due to their size they are often called compact flange joints (CFJs).

For flat face flanges in metal-to-metal contact, separation occurs at the bore for low pressure and external loads. The amount of separation depends upon the stiffness of the flange and the amount of bolt preload. A high degree of preload also minimizes fatigue of the bolts during cyclic loading. On the other hand, such flanges require more bolting than comparable raised face flanges since the bolt load is increased as a result of the interaction of the flanges beyond the bolt circle.

By tapering the face of the flanges, so that contact at the bore occurs first during assembly, it is possible to design for zero separation at the bore or to limit the separation to an acceptable value based on the sealing characteristics of the seal ring. When a self-seating and pressure activated seal ring is used between the tapered faces, the major sealing force is applied where it is needed, i.e., inside the bolt circle close to the bore.

Haagen (1967) describes the design of a modified raised face flange where one flange has a lip machined at the outer edge. By controlling the initial gap between the lip and the mating flange, tightening the bolts to a predetermined stress place the flanges in "controlled" metal-to-metal contact. As a result, at the design pressure: (a) flange separation is eliminated; (b) bending stress in the hub is minimized; and (c) the bolt stress is independent of internal pressure.

Webjörn (1967) introduced a gasket free CFJ with a slight flange face taper using high strength bolts (ISO class 10.9) preloaded to 80 % of the bolt yield strength, see also Webjörn and Schneider (1980), Hyde et al (1988). Since then, other CFJ proprietary designs have been introduced in the marked.

mailto:[email protected]

PVP2002-1087

Most of these joints are utilizing a non-load carrying self-seating and pressure activated seal ring located either at the flange bore or in a seal groove. The increasing interest in the industry to apply CFJs has resulted in a new flange standard, Lassesen at al (2002). The standard CFJ has flanges with a slight face taper and is using a seal ring and high strength bolts with equivalent strength to ASTM A193 B7 preloaded to 70 % of the bolt yield strength, see Fig.1.

Flange half

Seal ring

Bolts

Fig. 1 Compact flange joint The CFJs were typically applied in conditions with high-

pressure, significant external loads and/or cyclic (dynamic) loading. However, the CFJs are applied in a larger extend in standard process piping due to their weight, size, cost and safety against leakage.

A CFJ may be designed to offer the structural strength and fatigue strength of a welded joint. However, there is no published well-established practice on the designing flange joints with tapered flange face, which is in contact outside the bolt circle after tightening the bolts. The calculation rules of ASME and EN do not apply for this type of joint. For the benefit of engineers whom design and use CFJs, the intention of this paper is to provide all necessary information to design CFJs in metallic materials. Similar design method as presented here has been applied in many years to design of CFJs for application to high-pressure vessels and piping in addition to pipelines and risers.

CFJ DESCRIPTION The CFJ described in this paper consists of two weld neck

flanges, bolting and a seal ring, see Fig. 1. The distance from the flange bore to the inside edge of the seal groove is named flange heel. The outer contact area of the flange face is named flange toe, see Fig. 2.

Heel

Girth weld

ToeSeal ring

5

1

2

5

6

3

Fig. 2 Flange joint characteristics

The flange ring is closely machined with a slightly face

angle to assure that upon assembly, bore contact is established first for the flange faces. This is resulting in a gap at the outside diameter prior to preloading the bolts, see Fig. 2. The joint is closed using closely spaced bolts with high preload spaced around a bolt circle that is close to the outside diameter of the pipe or nozzle. The main design characteristics of a CFJ are: 1. High contact stresses and local yielding is obtained at the

flange bore (heel) at bolting up, i.e. the heel is "seated". This means the heel may act as a seal if a certain minimum heel compression load is provided in operation. The smooth bore with heel contact eliminates turbulence, erosion and crevice corrosion on flange faces.

2. The self-seating and pressure activated seal ring is located in a seal groove isolated from bolting up and piping loads and is not directly exposed to internal fluids. The seal ring has sufficient leak tightness for a face separation occurring at the structural capacity of the joint. The seal ring also does the final guiding of the joint at bolting-up.

3. The joint is designed and preloaded such that flange face separation is avoided for normal operation conditions, hence the joint behaves like a rigid body with no moving parts. The bolt load is almost constant up to normal operating loads. This reduces the risk for bolt fatigue failure and the risk for leakage due to wear, corrosion or fretting of the seal ring during operation.

4. Most of the bolt preload and external forces are transferred as contact forces between the flanges within the bolt circle, hence bolt loads due to flange face contact forces outside the bolt circle, i.e. bolt prying effects, are insignificant.


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5. Closing of the small gap at the flange toe may be used as an indicator of obtaining the target bolt load during bolting-up. Excessive bolt tightening or compressive external forces cannot damage the seal ring or the flange as contact forces between the flanges balance these forces.

6. The flanges have an elliptical transition between the ring and hub to obtain low stress concentration factors (SCFs). Values in the range of 1.5 are normally observed with respect to unit axial stress in the connected pipe. The flange geometry and makeup influence the stresses at the girth weld connecting the flange to the pipe. A typical SCF at the weld ID is 0.9 and 1.1 at the weld OD. These values have to be included in addition to the SCF introduced by any geometry misalignment in a fatigue assessment of the girth weld.

DESIGN RULES

Safety and failure modes The overall goal by design, material selection,

manufacture, testing, assembly, safety systems and maintenance is to keep the failure probability for a flanged joint below an acceptable level in service. Safety is achieved by incorporating appropriate design factors or safety factors using calculations, e.g. formulas or finite element analysis, and experimental testing against relevant failure modes. The design factor(s) accounts for the integrated uncertainty and possible bias in load effects and resistance. A safety factor is defined as a failure load divided by the allowable or design load. The following failure modes are normally considered in flange joint design: excessive yielding (gross plastic deformation), leakage, fatigue failure and unstable fracture.

Excessive yielding means exceeding the plastic load carrying capacity of the joint. Leakage means exceeding a target leak rate. Fatigue design involves minimizing flange stress concentrations or stress raisers, keeping the operating bolt stress ranges low and avoidance of flange face separation to have no relative motions between seating surface and seal ring/gasket to avoid seal degradation. This may be obtained by an elliptical transition between flange and hub in addition to using high bolt preload. Materials selection and qualification are normally done to ensure that the materials are sufficiently ductile and have sufficient fracture toughness.

Design rules The codes provide design rules for raised face and flat

face to face connections, e.g. ASME VIII and EN 1591. However, the interaction between flange, gasket and bolts are treated different in the various codes and considerable discrepancies are found between these codes.

ASME VIII rules are applicable for design of two connection types subjected to pressure only; i.e. the ring type joint with the gasket as a load carrying element and full-face-contact type flanges with self-energized gaskets. In the former,

bolt load depends solely on gasket pressure and internal hydrostatic pressure. The elastic based calculation method for these joints is that developed by Waters, et al in 1937, and the gasket factors introduced by Rossheim and Markl in 1943. It is often named the Taylor Forge method. In the latter, the bolt load must also balance the contact force between mating flanges outside the bolt circle, and this involves the flange flexibility. The Taylor Forge method is subjected to several limitations, e.g. see list in PD 6438:1969.

The prEN 13445-3:1999 provides rules based on the Taylor Forge method for pressure design, however, it opens for use of a more modern alternative design method given in EN 1591. EN 1591 considers pressure, external axial forces and bending moments, nonlinear elastic behavior of the gasket and axial thermal effects. The EN 1591 applies limit load criteria for all parts of the flanged connection taking into account the scatter of the bolt preload. The leak tightness and strength criteria consider the life of the joint including bolting up, test and operation. The EN 1591 method is considered to be an improvement of the Taylor-Forge model.

Code safety factors The justification of the code design stresses in pressure

vessel and piping codes is experience, rather than rational analysis of the material response to the loading. ASME VIII was first published in 1915. The design (membrane) stress was originally taken as one fifth of the tensile strength. The so called “safety factors” have come down from 5 in the original ASME code to 3.0 in ASME VIII Div. 2, to 2.4 in draft EN codes, where other properties are also considered. However, for the brittle steels of that time, tensile strength was an adequate limiting property.

Presented code safety factors here are at the room temperature in order to simplify the comparisons. In present version of ASME B31.3 and ASME VIII Div. 2 the flange design stress for ferritic steels is limited to the minimum of Rp0.2/1.5 and Rm/3. Rp0.2 is the specified minimum yield strength at room temperature, and Rm is the minimum ultimate tensile strength at room temperature. For austenitic grades, the design stress is Rp0.2/1.5. Bolt design stresses for ASTM A193 B7 bolting is the lower of Rm/5 and Rp0.2/4 in general, however, in Appendices 4, 5, and 6 of Section VIII Div. 2 is the bolt design stress equal to Rp0.2/3. The allowable stresses above are for pressure loading only.

When discussing ASME design stresses and standard ASME B16.5 flanges, Rodabaugh (1972) makes an interesting remark: "B16.5 flanged joints do not necessary meets the criteria in the ASME Boiler Code. Experience and a more detailed analysis indicate that it is not necessary to meet the ASME Code rules in order to have a satisfactory flanged joint and, on the other hand, meeting the ASME Code rules does not necessary assure a good flanged joint for use in a pipeline".

For ferritic flange grades, the design stress is the smaller of Rp0.2/1.5 or Rm/2.4 in prEN 13445-3:1999. For austenitic


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grades, the design stress is Rp1.0/1.5. The bolt design stresses for the Taylor Forge method in prEN 13445-3:1999 is the lesser of Rp0.2/3 and Rm/4. Design stresses for bolts in EN 1591 are determined as for flanges.

In designing API 10 000 psi and 15 000 psi flanged joints in API 6A for wellhead equipment, the allowable stresses at design pressure were set to Rp0.2/2.0 of the flange and the bolting materials to arrive at reasonable dimensions, Eichenberg (1964). The target prestress in the bolts for API 6A flanges is Rp0.2/2.0, hence flange face separation is expected to occur for external loads in excess to the design pressure and for pressure testing. Properly made-up joints do not leak during pressure testing, as the crushed metal gasket is partly pressure activated. This makes the API joint unsuitable in cases where cyclic external loads occurs in excess of the design pressure.

Limit analysis Limit analysis addresses directly the design objective of

preventing gross plastic deformation with an agreed-upon safety factor. Limit analysis with safety factors on yield strength only presumes use of sufficient tough, ductile, sound and strain hardening materials to ensure that flange joints can attain the required plastic deformed state before premature failure. When the yield strength is applied, the resulting limit load provides a physical connection between the calculated load and the "real" capacity found by testing or elastic plastic finite element analysis, hence indicating the "true" safety factor.

In limit analysis, the loading includes only primary loads such as pressure and weights. Stresses and strains generated by bolt preload (fixed displacement) or temperature fields do not affect limit loads. Such constraints produce external forces (reactions) that are self-limiting.

For ferritic flange grades the code limit load is based on a "yield" strength equal to 1.5 x min (Rm/3 ; Rp0.2/1.5) and 1.5 x min (Rm/2.4 ; Rp0.2/1.5) in ASME VIII Div.2 and EN 1591, respectively. According to this approach, the calculated limit load will be less than the yield-point limit load of the flange unless Rp0.2/Rm is less than 0.5 or 0.63 for the ASME and EN code respectively. Consider a flange made of ASTM A105 with Rm=485 MPa and Rp0.2=250 MPa and ASTM A694 F52 with Rm=455 MPa and Rp0.2=360 MPa. In this case is the safety factor against its yield-point limit state 1.55 and 2.37 for A105 and A694 F52 for the ASME code limit load. High strength steels, duplex stainless steels (which are treated as ferritic) and steel bolts suffer from this as the ratio of yield to tensile strength for these steels is close to 0.9.

Based on the author's experience, yield point limit loads fit very well with elastic ideal plastic finite element analysis and gives lower bounds compared to experimental testing, hence the code safety factor seems to be varying. However, ASME VIII, Div. 2 Appendix 4 gives a safety factor of 1.5 on the capacity obtained by experimental testing. It may then be argued that yield-point limit load analysis may be performed

with a safety factor of 1.5 on the limit load also fulfills the requirements in ASME VIII, Div.2 Appendix 4. Using yield point limit loads or plastic design for design of components requires that the materials exhibit sufficient fracture toughness and ductility to ensure that it can attain the required plastically deformed state without premature failure. It should further be noted that a safety factor of 1.5 on yield strength is also applied in ASME B31.3 high-pressure piping, DIN 2505 for flanges and API Spec.6A for wellhead equipment including flanges in addition to several steel structural, pipeline, e.g. ISO 13623:2000, and riser codes.

As the load is restricted to a level of 2/3 of the limit loads, the degree of yielding or permanent deformation in a flange joint is restricted to small values, see Fig. 9, which will not cause leakage or malfunction. In the case of cyclic loading, the subsequent strain portions are linear, ensuring shake down, as long as the stress range is less than 2 times the yield strength. For load changes between zero and maximum load, swelling loads, differences of deformations are linear, if the safety factor of 1.5 against limit loads is used. The load characteristic of flanges is not swelling because the bolts preload the flange joint. Common ductile materials show hardening effects in the stress strain relation that increases the range of linearity compared to elastic perfectly plastic behavior.

CFJ DESIGN METHOD

General In ASME/EN ring type joints, the gasket separates the

flanges and is a load-carrying element. Therefore it must be strong enough to take the full bolt load when the bolts are tightened and no pressure exists in the flange. The bolt load in flange consists of the load caused by pressure and external loads trying to separate the flanges plus the load necessary to keep the gasket tight, which load is assumed to be a multiple of the unit pressure, exerted on the projected sealing area of the gasket. A vicious circle is established thereby: The greater the bolt load, the greater the gasket width and seating area to support it, in turn necessitating an increase in bolt load. Enormous gaskets and bolts can be designed this way.

If flanges are made up face-to-face, this arrangement will support the bolt load when no pressure is on the flanges; and if the seal ring is self seating only a small initial load is necessary to establish sealing. Therefore the bolts have to carry only the pressure and external load plus any small axial component of the seal ring contact pressure. Thus the seal ring cross section becomes independent of the bolt load. The present design method applies to circular bolted flange connections with self-seating and pressure activating seal ring with metal-to-metal face contact.

It is important to note that the operating bolt load is relative insensitive to the changes in preload up to a certain point where separation occurs and that thereafter the two loads are essentially the same, see Fig. 3. This is a desirable


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characteristic of CFJs; it means that if the assembly load (preload) in the bolts, FB0, is close to the normal operating loads the subsequent application of pressure and external loading will have no significant effect on the actual operating loads in the bolts.

Facecontactforces

Bolt force

Applied separation force

Applied separation force

Bolt

f or c

e

Zero preload

0BF

BF

Preload

Fig. 3 Bolt force – applied separation force

There are three separate elements of CFJ which must act

to provide a leak tight joint. They shall be considered in the following order: seal system design, bolting design and flange design. A well-designed CFJ must have sufficient contact pressure on sealing faces to keep the joint tight without overstressing the flange and bolting material. The contact pressure is applied to the seal ring sealing faces by means of elastic spring back forces and internal pressure and the heel seal contact forces are applied by means of the bolting.

Excessive yielding The CFJ strength sizing is performed by limit load

methods using design stress based on yield strength. As the CFJ has flange face to face contact only pressure and separation forces need to be considered in the CFJ design. In both flanges and bolts, a design stress equal to Rp0.2/1.5 of the respective material may be used for normal operating conditions considering design pressure and external loads. This ensures a balanced strength design between bolting and flanges. It is important to include pressure and resulting separation forces in the CJF design as this is governing the dimensions and bolting. For bolting up condition the bolt design stress is Rp0.2/1.05, see prEn 13445-3:2002 and EN 1591:2001. Bolt stresses are based on the root diameter. For extreme design loads and accidental (survival) loads a safety factor of 1.25 and 1.0 may be applied. A CFJ may also be designed to have equivalent limit load capacity as the connecting pipe. Analytical based load-bearing capacities for pipes subjected to pressure, tension and bending can be found in Kirkemo (2001).

Leak tightness A "tight" joint, implying one with zero leakage, is an

outdated concept, as a joint will always have a leak rate. This has been recognized in the EN 1591-2 and also in the proposed ASME Appendix BFJ entitled Bolted Flanged Joint Design. A "leak tight joint" may be a connection with a nitrogen gas leak rate less than 1x10-5 - 1x10-6 cm3/sec/mm sealing diameter, measured at atmospheric pressure at normal operating conditions. The seal rings leak tightness is to be checked at both low and high pressure due to the pressure-activating characteristic of the seal ring. Low pressure sealing performance of the seal ring may be improved by using O-rings on the outer flanks.

A safe and reliable seal against liquids and gases under pressure cannot be achieved with compressive forces that produce elastic deformations at the interface areas only, regardless the degree of surface finish, Butcher (1973). With plastic flow of the material, surface asperity differences are leveled out and the leakage flow passage is blocked. Sufficient leak tightness of the CFJs are achieved by the following experience based requirements: 1. Seating of the seals at bolting up by plastic deformation

of the seal interface areas. 2. Average contact pressure of 2 times the internal pressure

over a contact width of 1 mm during operation. 3. A surface roughness not exceeding Ra=0.8 µm as defined

in ISO 4287 applied for the heel and the seal ring and seal ring seating area. Lower surface roughness may be required for sealing helium and hydrogen. The seal ring or flange heel may be plated with soft

metals such as silver or gold or coated with a thin film of viscous oil, MoS2 or Teflon to provide a relatively soft surface, which flows into minor imperfections of the flange seating/seal ring at bolting up and improve leak tightness. The selection of plating or coating should based on the allowable leak rate, the viscosity (density) of the fluid, flange roughness and the application temperature.

SEALING SYSTEM DESIGN Seating of the sealing system is achieved by requiring that

a contact pressure corresponding to yielding are obtained over a fictitious contact width equal to 1 mm of the heel and the seal ring, see Fig. 4. The heel is seating during bolting-up due to the flange face taper and high bolt preload. The seal ring is seated when metal-to-metal contact occurs at the bore.

The seal ring must also perform a number of other different jobs in addition to create a seal between the mating faces, to function properly. It must do the final guiding of the flange halves during bolting up and be easily to install and remove.

The flange bore, B, may be established as follows Po eDB ×−= 2 (1)

where Do is the pipe/hub outside diameter in mm, eP is the pipe wall thickness. The inside diameter of the free seal ring


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DRi is chosen equal to the pipe/neck outside diameter, see Fig. 5,

oRi DD = (2) This ensures that the distance from the inside edge of the

groove to the bore is almost equal to the connecting pipe wall thickness, hence, the inner flank of the groove resists any pressure and external loads applied to the connecting pipe. The flange surfaces are slightly conical so that they only meet at the inner edge after seating the seal ring. This ensures contact stresses in excess of yield strength at the inner edge after bolting-up, i.e. heel seating.

Stand off

Free position

Stand off

Free position

Stand off

Free position

Stand off

Free position

Contact forces

Seating loads

Self seating

Pressure activated

Assembled position

Operating

Contact forces

Fig. 4 Sealing system

The height of the ring, HR, is arbitrary taken as,

oR DH ×= 2 (3) This produces a ring height of 15.5 mm for a 2 in. pipe, and 33 mm for a 10 in. pipe.

The flank angle of the groove ϕ is set equal to 15°. This angle is also applied by the Grayloc type seal ring (1964). The ring is double cone, with cone angles of 15-2=13° and 15+2=17°. A theoretical line contact for sealing is neither desirable nor feasible. The double cone seal ring has therefore a contact radius of 5 mm, Butcher (1979). The height of the upper flank is 1/6 of the total height, HR, of the seal ring as shown in Figs. 5 and 6. This gives an axial distance between the two sealing lines HR,s as follows

RsR HH ×= 32, (4) The initial seating stress of the seal ring is created by the

wedging action of an inclined seal surface, Fig. 5. The wedging action of the seal groove compresses the ring in the

hoop direction. The groove seal surface bears against a contact radius on the seal ring. The radial force on the ring, FR,r, generating a contact pressure corresponding to yielding over a 1 mm height, when neglecting the effect from the flank angle as cos(15°)=0.97, is given by

RpRirR RDF ,, 1×××=π (5) where Rp,R is the seal ring yield strength. Naming bRs as the ring thickness at the sealing diameter Ds, the hoop stress in the ring subjected to a radial force FR,r becomes, see Fig. 6,

RpRs

RiRRi

rR

h Rb

DHD

F

,

,

2

2

=×

×××

×

= πσ (6)

Hence

R

RiRs H

Db = (7)

( 2tan3

−×+= ϕRRsR

Hbb ) (8)

where bR is the total radial width of the seal ring and (ϕ-2) = 13° is the lower flank angle of the seal ring, see Fig. 6.

RiD

RHRsR HH

32

, =

Rb

Rsb

SO

o15=ϕ

N

Q

B

Free position

Made up positionSD

B

Original position

GoD

Fig. 5 Seal ring and groove dimensions It should be noted that the seal width is independent of

the yield strength. This method of sizing has been applied with great success for seal ring metals with yield strength in the range of 350 MPa to 720 MPa. Compressive stresses in the range of yield stress in the ring direction might result in buckling of the ring if the slenderness is low even if the ring is guided in the seal groove with outside contact pressure. Based


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on calculations and testing it can be indicated that buckling will not govern the width of the seal ring for the applied design method and yield strength range. Experience has also shown that these rings have sufficient shear strength to do the final guiding of the flanges during bolting-up.

rRF ,

rRF ,

Rsb

o17

o13

Rb

RH32

ip

sD

Fig. 6 Seating of seal ring

The width of the groove, N, is made such there is a radial

interference, I=Rorig-Rfinal, between the unseated and seated (made up) seal ring diameter sufficient to generate yielding in the ring direction during bolting-up to insure initial seating. The amount of initial radial compression I necessary to generate yield stress in hoop direction of a ring with an actual yield (or flow) strength 50% higher than minimum specified, is given as

( ) 2.05.1

25.0 , +

×××+×=

R

RpRRi E

RbDI (9)

where ER is the seal ring modulus of elasticity in N/mm2. A margin of 0.2 mm is included in Eq.(9) to cope with manufacture tolerances of ring and groove. The gap between the flanges at the seal groove when the seal ring is in initial contact is named stand off. The stand off, SO, necessary to generate the radial interference, I, is given as, see Fig. 5,

( )ϕtanISO = (10)

The depth of groove, Q, is made sufficient deep to avoid interference with the seal ring considering compression of the heel regions and not too deep in order to avoid to large rotations of the ring in the groove during make up,

2.051.0 +×= RHQ (11) The width of seal ring groove including mating clearance to groove inner diameter, N, outside diameter of the seal ring groove, DGo, and fluid sealing diameter Ds, and seal ring cross section area AR becomes:

( ) ( )ϕtan3

5.0;5.1max, ×+++= RsR

HIbN (12)

( )

×+×+= ϕtan

32 R

RsRiGo

HbDD (13)

( )ϕtan32 ××−= RGoS HDD (14)

( )2

RsRRR

bbHA +×= (15)

The groove width N is made such that the ring will no interfere with the groove when flanges are lined up and bolts are inserted in the bolt holes. The corners of the seal ring and groove are rounded with radiuses. To assist assembly, the seal rings can be retained in the flanges by making an outer recess in the ring, see Fig. 6, and using a retainer fixed to the flange face.

The seal rings have shown by elastic plastic element analyses and testing to have sufficient gas leak tightness at normal operating conditions and sufficient water tightness up to the structural capacity of the CFJ. Note that the flange rotation at the limit capacity of the CFJ increase the sealing action as the seal groove moves inwards due to flange rotation.

During assembly, the compressed seal ring exerts an axial force on the flange seat. This make-up (seating) force becomes

( )θϕπ +×××= tan,,0 RRpaR ARF (16) where θ is the friction angle in °. θ = atan(µR), where µR is the friction coefficient between the seal ring and seating face. The axial component of the seal ring retaining force FR,a during testing and operation conditions is

( )

××+××−×=

2tan ,

,,iSsR

RRpaRpDH

ARF θϕπ (17)

where pi is the internal fluid pressure in N/mm2. The first part in Eq. (17) is the retaining force for zero pressure, i.e. elastic spring back force and the last part is the pressure induced retaining force, see Fig. 6.

BOLTING AND FLANGE OUTLINE DESIGN With the size and shape of the seal ring and groove

established, the next step is a calculation of bolt size and a determination of the flange outline, except the thickness. The bolting should be selected to maintain the required compression on the flange faces with internal pressure and external loads acting, i.e. the flange face contact when subject to normal operating design loads. Fig. 7 illustrates the notation used for dimensions, forces and lever arms. The forces are considered to be uniformly distributed on the circumference.

Theoretically, the hydrostatic pressure extends only to the inside diameter of the flange. However, mechanical damage of the flange heel and not sufficient bolt tension tend to permit the confined fluid to creep over the heel face. In order to be on the safe side, the design of the CFJ is based on the worst possible sealing condition, namely, a hydrostatic pressure extending to the sealing diameter of the seal ring.


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aRF ,

BF

Dh

LFeRh

Th

KO

Pe

TF

DF

eqF

θ

B

oD

ey

exHl

g

Fig. 7 Flange dimensions and loads

The bolt load must balance the sum of the total

hydrostatic end force, the axial component of the seal ring retaining force, and the equivalent axial separation force. Therefore the minimum required root area of the bolts becomes

B

eqaRQB f

FFFA

++= ,

min, (18)

where

iSQ pDF ××= 2

4π (19)

EEeq MK

FF ×+= 4 (20)

and where fB is the bolt design (allowable) stress in N/mm2, FQ is the equivalent axial separation load due to pressure (pressure trust) in N, FE the is external (additional or effective) axial tension in N, ME is the external bending moment in Nmm and K is the bolt circle diameter in mm. The axial separation force from the external moment is calculated as in EN 1591.

The number of bolts should be divisible by 4 and bolt sizes may not be selected less than ½ in. Smaller bolt sizes are prohibited in some codes due to the risk of overstressing during make-up. With these considerations, the size and number of bolts are selected, so that the actual bolt cross

section area using the root diameter of all bolts, AB,act, equals or exceeds the minimum required bolt area AB,min give by Eq. (18). Single bolt root areas are given by Eq. (55).

The bolt spacing and bolt circle diameter must be sufficient to provide the necessary makeup tool clearances. Make-up tools may include standard socket, hydraulic torque wrench or tension tool as appropriate, see Fig. 8. The bolt data given in Table 1 is based on access for use of hydraulic torque tools available in the marked. The selected tools should have a torque capacity of minimum 30 – 50 % in excess that necessary to make-up the lubricated bolt. The reserves are considered necessary for disassembly after a period in service.

Fig. 8 Hydraulic torque and tension tool

The wrench clearance Rmin in Table 1 is added to the

minor half axis ye to determine the minimum bolt circle diameter, K:

( min2 RyDK eo +×+= ) (21) The wrench clearance is the radial distance from bolt circle diameter to start ellipse. The minor half axis ye of the neck ellipse is given by

=

5.2;3min P

e

ey (22)

The major half axis xe is 3.5 times the minor half axis ee yx ×= 5.3 (23)

The selection of the elliptical transition ensures low fillet stresses between the flange and hub.

Next, the distance between bolts must be calculated and checked against the minimum bolt spacing dimension in Table 1, to guard against torque tool interference,

=

KB

nBminarcsin

π (24)

where Bmin is the minimum pitch (bolt center-center distance). The flange outside diameter O then becomes

min2 EKO ×+= (25) where Emin is the radial distance from bolt circle diameter to flange outside diameter, assuming nut corner is flush with flange outside diameter, see Table 1. The hub length lH in mm is estimated as


PVP2002-1087

( )( )25 ; 2/10max peH exl ++= (26) where a minimum length of lH is assumed to be 25 mm to allow for weld access during welding/NDT. The length in excess of xe is sufficient to account for a straight part between end ellipse and weld bevel.

FLANGE RING THICKNESS AND FACE ANGLE At this time, all flange main dimensions except the flange

thickness eF are know. The internal flange (warping) moment M due to load acting on the flange is the product of the resulting load and its moment arm, see Fig. 7. The applied moments have to be resisted by the moment capacity of the flange, hence, the flange thickness can be determined.

The internal flange moment for operation conditions is resulting from the sum of pressure end load, external loads and the seal ring retaining load for the relevant conditions as follows:

( ) RaRTTDeqDF hFhFhFFM ×+×+×+= , (27) where

iD pBF ××= 2

4π (28)

( ) isT pBDF ×−×= 22

4π (29)

( ) 2pD eBKh −−= (30) ( ) 42 sT DBKh −−= (31) ( ) 2sR DKh −= (32)

and FD is hydrostatic end force applied via the pipe to flange in N, FR is seal ring retaining load in N, FT is hydrostatic end force due to pressure on flange face in N. The moment arms hD, hR and hT are the radial distances from bolt circle to circle on which FD, FR and FT acts in mm. The loads acting on the flange are assumed uniformly distributed around the circumference of the circles of diameters.

Proper allowance has to be made if connections are subjected to external loads. In cases where the external loads are not know, the equivalent axial tension acting on the CFJ may be chosen as

ioeq pDF ××= 2

4π (33)

The internal flange moment capacity, i.e. the limit load, of the flange including support from the neck is given by:

×××

+××××××

+×××

×=

PppM

PpPPFS

FFF

F

fedc

fedeec

feb

W2

2

2.2

2

4π (34)

where

PP

PiQ ef

dp××

×=2

δ (35)

PPP

RF edf

F×××

=π

δ (36)

+

×−×

+−= 2

22

43

123

4R

QR

QMc δ

δδ

δ (37)

( )RQMS cc δδ ×+×−×= 4.06.08.0 (38)

( ) LBObF −−=2

(39)

and fF and fP are the flange and pipe/hub design stresses, respectively, in N/mm2, bF is the radial width of flange ring excluding the bolt hole diameter in mm, δQ and δR are pressure and external loading parameters, and cS and cM are correction factors. Eq.(34) is based on Draft.2, 1992 of the EN 1591. The limit load of the flange ring in EN 1591 is corrected to be in line with the theoretical flange ring limit load. EN 1591 subtracts only a partial bolt hole diameter, while limit load analysis require that the total bolt hole diameter L shall be applied to establish the flange radial width.

Bolt size

AB1 Bmin Rmin Emin L

in.

Threads per inch

mm2 mm mm mm mm 1/2 13 81.1 29.1 16.3 12.8 15.0 5/8 11 130.2 35.1 19.5 15.6 18.0 3/4 10 194.8 42.3 24.0 18.3 22.0 7/8 9 270.4 49.3 28.2 21.1 25.0 1 8 355.4 56.6 32.8 23.8 29.0

1 1/8 8 469.4 62.1 35.6 26.6 32.0 1 1/4 8 599.3 70.7 41.4 29.3 35.0 1 3/8 8 744.9 76.3 44.2 32.1 39.0 1 1/2 8 906.5 82.3 47.5 34.8 42.0 1 5/8 8 1083.8 90.2 52.6 37.6 45.0 1 3/4 8 1277.0 95.7 55.4 40.3 48.0 1 7/8 8 1486.0 101.5 58.4 43.1 51.0

2 8 1710.9 110.1 64.3 45.8 54.0 2 1/4 8 2208.1 122.3 71.0 51.3 61.0 2 1/2 8 2768.6 138.4 81.5 56.8 67.0 2 3/4 8 3392.5 149.7 87.4 62.3 73.0

3 8 4079.7 161.0 93.2 67.8 80.0 3 1/4 8 4830.3 172.1 98.8 73.3 86.0 3 1/2 8 5644.2 181.8 103.0 78.8 92.0 3 3/4 8 6521.4 194.3 110.0 84.3 99.0

4 8 7462.0 205.8 116.0 89.8 105.0 where AB1 is the cross section area of a single bolt using the root diameter in mm2, see Eq. (55) L is the bolt hole diameter

Table 1 Bolt and torque wrench data. The first and last part of Eq.(34) is ring and pipe wall

thickness internal flange moment resistance. The reduction factors cM and cS take into account the reduction of the


PVP2002-1087

bending-carrying capacity and shear force capacity of the pipe cross section assuming von Mises yield criterion. The factors are based on pipe wall yielding and not the actual cross section yielding capacity, see Kirkemo (2001). The middle part contains the support effect of a radial force from the pipe for the ring. If the value in the root giving cM is negative the hub/pipe is overloaded. Hoop stress caused by internal pressure is neglected in the flange ring, however, included in the strength contribution from the connecting hub/pipe.

The flange ring thickness can now be calculated by requiring that WF should be equal to MF by an iterative solver available in spreadsheets. The initial flange face angle θ in radians is calculated as

FKM min0=θ (40)

DBB hFnM ××= min1min0 (41) where M0min is the minimum applied bolting up internal flange moment in Nmm, KF is the elastic stiffness of the integrated flange ring and cylinder and FB1min is the minimum bolt force for one bolt in N. KF is given by

FF

FFeFF cd

ebEK××

×××=3

3π (42)

and the correction factors are as follows

( ) ( )

××+

×+×+×+×+×=

42

2

3

6641191.0

ϑγϑϑϑγ

ϑγFc (43)

( ) ( )PFeFP dbde ××=γ (44)

FPP eed ××= 4.0ϑ (45) ( )

eFe LBOb −−=2

(46)

KnLLL B

e ××=

π (47)

PP eBd += (48) ( )

2BOdF

+= (49)

The flange stiffness takes into the adjoining effective cylinder shell by multiplication with . The factor

is modified by the 0.91 factor compared to factor given in EN 1591. Furthermore, the constant in

( PP de , ) Fc

Fcϑ is 0.4 compared to

the 0.55 factor applied in EN 1591. The effective gap at the flange toe g is calculated as, Fig. 7,

( )2

tan9.0 BOg −××= θ (50)

As the toe gap is 90% of the theoretical elastic value, closing of this gap during bolting up is an indicator of some minimum applied bolt preload.

Due to the initial flange face angle, most of the bolt preload and external loads are transferred as contact forces between flanges within the bolt circle due to flange taper. This in combination with stiff flanges and flexible bolts and a

design with balanced strength between flange and bolts, excludes any flange interacting outside the bolt circle, hence any additional bolt stress generated due a prying effect can be neglected. The back face of the flange is made parallel to front face in the made-up position; hence, bending in the bolts is reduced to a minimum.

Considerable elastic and elastic plastic finite element analyses, Fig. 9, have been performed to justify the applied limit load based design and stiffness equations. Capacities should be determined using elastic-plastic finite element analysis to avoid the necessity of dividing the stresses into primary and secondary stress categories and linearisation of stresses as required in elastic analysis. The structural capacity is determined by increasing the loads nearly to the point of instability (maximum) or when the local strains exceed 5 %. The design capacity is found by dividing the structural capacity by 1.5. Only limited permanent deformation occurs at this load level, see Fig. 9.

Flange capacity

0.0

1000.0

2000.0

3000.0

4000.0

5000.0

6000.0

7000.0

0.00% 1.00% 2.00% 3.00% 4.00% 5.00% 6.00%Total strain in neck

Tota

l sep

arat

ion

forc

e

Design capacity = 2/3 yield point limit load

Calculated yield point limit load

Elastic plastic FEA

FEA notch strain limit

Flange ring rotation

0.0

1000.0

2000.0

3000.0

4000.0

5000.0

6000.0

7000.0

0.0000 0.2000 0.4000 0.6000 0.8000 1.0000 1.2000 1.4000 1.6000 1.8000Rotation, deg

Sepa

ratio

n fo

rce

Elastic plastic FEACalculated yield point limit load

Design capacity = 2/3 yield point limit load

Calculated elastic rotation

Fig. 9 Finite element analysis of flange ring


PVP2002-1087

ASSEMBLY CONSIDERATIONS Successful sealing a flanged connection depends on all

components of a well-designed flange system working well together. These include not only design of sealing system, bolting and flange as a system but also assembly guidelines.

Initial bolt loads in ASME B16.5 flanged joints have not always been accurate. Rodabaugh (1972) states: "In field installation of B16.5 flanged joints the initial bolt stress is seldom controlled; the pipe fitter simply tightens the bolts to what he considers to be an appropriate amount". Tightening groups of bolts in a gasketed ASME joint results in significant elastic interaction. Individual bolts can loose up to 95 % of their initial preload, Bibel (1995). Bibel (1995) further states that final bolt load can be as low as 45 % of design even after three pass bolt-up procedure. However ASME has recognized the importance of guidelines for flange joint assembly by issuing ASME PCC-1 (2000). Typical target bolt prestress in ASME bolted flanged joints has changed from 275 MPa (40 ksi) in 1972, Rodabaugh (1972), to 345 MPa (50 ksi) today, ASME PCC-1 (2000).

Only qualified assemblers with calibrated torque wrench, hydraulic or other tensions shall assemble bolted flange connections like ASME B16.5 joints and CFJs. Assembly must be to a written procedure, which is qualified by test to achieve the minimum residual bolt load. Typical steps in assembly of CFJs are as follows: 1. Clean and examine the CFJ components before assembly

is started. All sealing surface shall be free from mechanical damage and rust and have appropriate surface finish.

2. Align flanges and bolt holes such that the bolts easily can be installed.

3. Install the seal ring carefully between flanges, check that the seal ring slightly rocks in the groove (stand off) and bring the flanges together without damaging the seal ring.

4. Lubricate nut load-bearing surfaces and bolt ends with specified lubricant.

5. Install bolts and nuts hand-tight, then "snug up" to 15 Nm to 30 Nm. Number each bolt.

6. Tighten the bolting evenly to specified torque values in a cross-pattern tightening sequence. After full torque is applied, apply at least one final torque to all nuts in a clockwise direction until all torque is uniform and check that the flange gap is closed.

All preload methods involve some degree of inaccuracy, which should be accounted for. The scatter in bolt preload is accounted by the scatter value ε for the bolt preload as follows:

ε−=

1min10

10B

nomB

FF (51)

( )ε

εε−

∆++=∆++=1

11 min1010max10 BnomBB FFF (52)

where FB10min is the minimum bolt force in N, FB10nom is the nominal (average) bolt assembly force in N, FB10max is the

maximum bolt assembly force in N, ε is the residual bolt preload scatter value and ∆B is the bolt transfer loss for tension tool, (= 0 for torque tool). The tension tool preload transfer loss may be calculated by

0

00

8.029.09.0

BF

B

B

BB de

dl

d×+×

×=×=∆ (53)

where lB is the effective (clamp) bolt length in mm and dB0 is the nominal bolt diameter in mm (=25.4 mm for 1").

Bolt preload scatter values (standard deviation) 5 - 8 % have been obtained for lubricated (MoS2) galvanized ASTM A193 B7 bolts using a friction value of 0.12. Using friction values on the high side ensures that the mean bolt preload are on the high side, hence the mean value minus the scatter is higher than the minimum required bolt preload. The bolt-preload scatter for ASME value of a CFJ B16.5 ring type joints are typical the double of what is obtained by CFJs. Gasketed ASME bolted flange joints have higher bolt interaction and larger bolt bending due to flange rotation than CFJs which have metal-to-metal face contact and almost zero bending due to parallel flanges after make-up.

An adequate estimate of the relationship between tightening torque and axial force in the bolt for ASTM A193 bolts and ASTM A194 heavy hex nuts is computed as follows

( nomBBBnomBt FdpM 100, 23.116.0 ×××+= )µ (54) where MBt,nom is the nominal (target) bolt torque in Nmm, p is the thread pitch in mm (=25.4/n), n is the number of thread per inch (=8 for 8UN threads), µB is the average friction coefficient on thread and under nut, FB10nom is the nominal (target) axial preload in the bolt in N (=fB0nom x AB1), fB0nom is the nominal (target) initial bolt stress in Nmm2 and AB1 is the bolt root area of a single bolt in mm2 given by

( 21 3.1

4pdA BB −= )π (55)

The actual minimum bolt preload should be in the level of 2/3 of the bolt tension yield capacity. This ensures that the sealing surfaces are in a stable condition (static) for normal design conditions, i.e. there are no relative movements of sealing surfaces. The bolt utilization ratio UR at bolting up is

112312

30

max,

2

1

max0

0

≤

×

×+

×=

BBt

B

B

B dM

AF

fUR

π (56)

where UR is applied load divided by allowable load, fB0 is the bolt design stress at bolting-up and the maximum torque value is given as:

( )ε+×= 1,max, nomBtBt MM (57) and MBt,max=0 for hydraulic tensioners. During bolting-up the torque is primary load while the wrench is loaded, however, after make-up, the torque is secondary. This means that torque can be neglected in the subsequent load conditions.

THERMAL CONSIDERATIONS Bolted flanged joint materials should be applied below

the lower bound of the creep range, e.g. 370 °C for ferritic


PVP2002-1087

steels, due to creep, causing relaxation in bolt and seal ring, and eventually the joint may leak. The load capacity for the CFJs at temperature is established by using the actual yield strength at temperature. For material strength at temperature, it should be noted that EN uses minimum yield and tensile strength values while ASME uses strength values based on average temperature dependent trend curve.

For thermal applications the bolt, seal ring and flange materials should not have coefficients of thermal expansion, which are differing too much. The bolt load will in general change with temperature. The axial bolt load at temperate FB1,T may be expressed by

( ) ( oFBTBBF

TFBTB TTEA

EE

FF −×−××+×= αα,10,

,min10,1 ) (58)

where EF,T and EF,0 is the flange elastic modulus at temperature T and assembly temperature T0, respectively, EB,T is the bolt elastic modulus at temperature T, αB and αF is the thermal expansion coefficients of the bolt and flange, respectively. The following may be observed from the expression: 1. The bolt force will reduce with increasing temperature

with equal thermal expansion in bolt and flange due to the drop in elastic modulus with increased temperature.

2. Higher thermal expansion in bolts than in flange will reduce the bolt load with increasing temperature.

3. Lower thermal expansion in the bolts than in flange will increase the bolt load with increasing temperature.

The axial bolt load at temperature including primary and secondary axial load effects should be kept below the yield strength at temperature to avoid permanent deformation of the bolt, hence avoid reduction of bolt preload when the joint is returned to room temperature.

Note that the seal ring and bolts are thermally shielded against direct influence from internal fluids and external thermal sources like fire.

EXAMPLE OF CFJ DESIGN An example of a CFJ sizing is given in this section. The

CFJ consists of 2 weld neck flanges with materials according to ASTM A694 F52. The stud bolts strength and threads are in accordance ASTM A193 B7 while the seal ring material is ASTM A694 F65. The flanges are connected to pipes with Do=273.1 mm (10") and wall thickness eP=26 mm. The flange connection is designed for a pressure of 258 bar, an equivalent tension equal to 1511 kN, Eq.(33), and a temperature of 20°C. The minimum target prestress is 2/3 yield strength. For more details see Table 2.

The comparable ASME B16.5 flanged joint is a 10" CL1500 ring type joint. The CFJ is considerably lighter and smaller than the ASME B16.5 flanged joint including torque tools, see Table 3. In Fig. 10, the CFJ is compared with the B16.5 flange joint. Main dimensions and weights are given in Table 4.

It should be noted that selection of other materials, pipe wall thickness and external loading would change the dimension of the CFJ.

Table 3 10" CL1500 CFJ and ASME comparison

Characteristic CFJ ASME B16.5 Outside diameter 418.2 mm 584 mm Thickness 71.1 mm 108 mm Total length 130.5 mm 254 mm Bolting 16 x 1 1/8" x

215 mm 12 x 1 7/8" x

345 mm Weight each flange1) 57 kg 205 kg Weight bolting 21 kg 73 kg Weight torque tool 2.5 kg 12.5 kg 1) Weight of one flange half with pipe length equal to total ASME flange length is 73 kg.

12 stud bolts1 7/8x345 mm

16 stud bolts1 1/8x225 mm

415 mm

584 mm

71.1mm

119.1 mm

273 mm

Fig. 10 Comparison of 10" CL1500 CFJ and

equivalent ASME B16.5 joint (dotted)

CONCLUSIONS Conventional flange designs with load carrying gaskets

have major shortcomings wrt. to leakage reliability and inability to cope with cycling loading and temperature. A design method for CFJs is presented and applied in an example for a flange design. The design principles of a CFJ presented in this paper are sound and offer many fundamental


PVP2002-1087

advantages over the conventional type of joint, apart from reduced weight and size. In the author's opinion, CFJs should gradually find their way into general industrial applications due to their leak reliability records. However, design codes should address these types of joints in future.

REFERENCES 1. API Spec. 6A, 1999, Specification for Wellhead and

Christmas Tree Equipment. 2. ASME, 2001, Boiler & Pressure Vessel Code, Section

VIII, Division 1 and 2, ASME International, New York, NY.

3. ASME B16.5, 1996, Pipe flanges and flanged fittings. 4. ASME B31.3, 1996, Process piping. 5. ASME PCC-1-2000, Guidelines for pressure boundary

bolted flange joint assembly. 6. Bibel, G., 1995, "Summary of PVRC research on bolted

flange assembly," PVP-Vol.307, ASME. 7. BS PD6438:1969, A review of present methods for design

of bolted flanges for pressure vessels. 8. Butcher, H.H., 1973, "Fundamental principles for static

sealing with metals in the high pressure field," ASLE Transactions, Volume 16, 4, pp.304-309.

9. DIN 2505 Part 1 Draft 1990, Calculation of flanged joints.

10. Eichenberg, R., 1964, "Design of high-pressure integral and welding neck flanges with pressure-energized ring joint gaskets," ASME Paper No.63-Pet-3, J. of Engineering Industry, May 1964, 86, (2), 199-2-4.

11. EN 1591-1:2001, Flanges and their joints – Design rules for gasketed circular flange connections – Calculation method.

12. prEN 13345:2002 (March), Unfired pressure vessels. 13. Haagen, T., 1967, "New flange connection for large

pressure vessels," First International Conference on

Pressure Vessel Technology, Part 1, Design and Analysis, September 29 – October 2, ASME, pp.155-164.

14. Hyde, T.H., Lewis, L.V. and Fessler, H., 1988, "Bolting and loss of contact between cylindrical flat-flanged joints without gaskets,", Journal of strain analysis Vol.23, No.1.

15. ISO 13623:2000, Petroleum and natural gas industries – Pipeline transportation systems.

16. ISO 4287:1977, Geometrical Product Specifications (GPS) - Surface texture: Profile method - Terms, definitions and surface texture parameters.

17. Kirkemo, F., 2001, "Burst and gross plastic deformation limit states equations for pipes: Part 1 – Theory," ISOPE 2001.

18. Lassesen, S., Nybråten, O. and Eriksen, T., 2002, "NORSOK L-005; Compact flanged connections (CFC) – the new standard," ASME PVP 2002.

19. "Pipe connection", Chemical Engineering, April 26, 1965, 72, (9), 183-4.

20. Rodabaugh, E.C., 1972, "Background of ANSI B16.5 pressure-temperature ratings," API, Preprint 54-72.

21. Rossheim, D.B., Markl, A.R.C., 1943, "Gasket loading constants," Mech. Eng., Vol.65, p.647-648.

22. Scwaigerer, S., 1954, "Die berechnung der Flanschverbindungen im Behälter- und Rohrleitungsbau," Z.VDI 96, pp. 7-12.

23. Waters, E.O., Wesstrom, D.B., Rossheim, D.B. and Williams, F.S.G., 1937, "Formulas for stresses in bolted flanged connections," Trans.ASME, April.

24. Webørn, J., 1967, "Flange design in Sweden," ASME Paper 67-PET-20.

25. Webørn, J. and Schneider, R.W., 1980, "Functional test of a vessel with compact flanges in metal-to-metal contact," WRC Bulletin No. 262.


PVP2002-1087

Table 2 Example sheet of CFJ sizing WELDING NECK AND INTEGRAL COMPACT FLANGE JOINT DESIGN

DESIGN BASIS Pipe/hub outside diameter Do 273.1 mm Yield strength, flange/hub Rp,F 360.0 N/mm2

Pipe/hub wall thickness eP 26.0 mm Yield strength, bolting Rp,B 720.0 N/mm2

Design pressure pi 25.8 N/mm2 Yield strength, seal ring Rp,R 450.0 N/mm2

External equivalent load Feq 1.51E+06 N Safety factor, operating SP 1.50 Elastic modulus, flange EF 200000 N/mm2 Safety factor, bolting up S0 1.05 Elastic modulus, seal ring ER 200000 N/mm2 Groove flank angle ϕ 15 °

Seal ring/seating friction coef. µR 0.10 Minimum target prestress fB0min 480.0 N/mm2

Bolt/nut friction coefficient µB 0.12 Bolt preload scatter ε 0.05 SEAL RING AND GROOVE CALCULATIONS Flange/pipe bore B 221.1 mm Radial interface I 0.70 mm Inside diameter of ring DRi 273.1 mm Stand off SO 2.60 mm Height of ring HR 33.1 mm Width of groove N 12.71 mm Minimum ring width at Ds bRs 8.26 mm Outside diameter of groove DGo 295.53 mm Width of ring bR 10.81 mm Depth of groove Q 17.06 mm Ring cross-section area AR 315.1 mm2 Fluid seal diameter Ds 289.6 mm BOLTS AND FLANGE OUTLINE CALCULATIONS Ring retaining load FRa 2.73E+05 N No. of bolts nB 16 Total hydrostatic end force FQ 1.70E+06 N Bolt size dB 1 1/8 in Minimum required bolt area ABmin 7258 mm Bolt hole diameter L 32.0 mm Actual bolt area ABact 7511 mm Diameter of bolt circle K 365 mm Minor half ellipes yE 10.4 mm Outside diameter of flange O 418.2 mm Major half ellipse xE 36.4 mm Hub length lH 59.4 mm FLANGE THICKNESS AND INITIAL FLANGE TAPER CALCULATIONS Pipe hydrostatic end force FD 9.906E+05 N Operation internal flange

moment MF 1.966E+08 Nmm

Flange hydrostatic end force FT 7.092E+05 N Flange ring thickness eF 71.1 mm Internal flange moment, FD MD 1.475E+08 Nmm Bolting up internal flange mom. M0 2.045E+08 Nmm Internal flange moment, FT MT 3.888E+07 Nmm Flange rot. due to min. preload θ0 0.27 o

Internal flange moment, FR MR 1.028E+07 Nmm Initial gap at flange toe gF 0.42 mm ASSEMBLY CALCULATIONS FOR SINGLE BOLT Maximum tension, torque FB1max 2.592E+05 N Tension tool transfer loss ∆B 16 % Bolt torque, target Mt 1225 Nm Tensioner tension, target Ft 2.970E+05 N Bolting up load ratio - torque URMt 0.97 Bolting up load ratio - tension URFt 0.92


design of compact flange joints

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