Joint strength of gasketed bolted pipe flange joint under combined ...

Original Article

Joint strength of gasketed bolted pipeflange joint under combined internalpressure plus axial load with different(industrial and ASME) bolt-up strategy

Niaz B Khan1, Muhammad Abid2, Mohammed Jameel1 andHafiz Abdul Wajid3

Abstract

Gasketed bolted flange joints are used in process industry for connecting pressure vessels and pipes. Design procedures

available in the literature mostly discuss structural strength, while sealing failure is still a big concern in industries.

Similarly, limited work is found in the literature regarding performance of gasketed bolted joints under combined loading.

A detailed 3D nonlinear finite element analysis is performed to study the strength and sealing of a gasketed bolted flanged

pipe joint under different bolt-up strategy (Industrial and ASME) and under combined internal pressure and axial loading.

Keywords

Bolt-up, gasketed, ASME, sealing, structural strength

Date received: 7 December 2014; accepted: 27 September 2015

Introduction

Strength or sealing failure of gasketed bolted flangejoints, whether small or large, is always a hazard andmay result in losses. Since there is rapid advancementin technology in the last 10 years for high pressure andexternal loading condition, there is a need for such ajoint which has adequate joint strength and sealing toperform safely under given operating conditions.

Limited work is available in the literature for thegasketed bolted pipe flanged joint under combinedinternal pressure and axial loading. Previously,three-dimensional nonlinear analysis of differenttypes of joints is performed by Power.1 Power studiedthe variation of contact pressure and stress variationin the flange and bolt. Contact pressure or interfacepressure was considered as main quantitative measurefor the sealing ability of the joints. Cao et al.2 alsoperformed three-dimensional analysis of gasket non-linearity to study the joint strength under bolt-up andoperating conditions. Fukuoka and Takaki3 mainlyfocus on the bolt scatter, tightening sequence andbearing surfaces. Nelson et al.4 studied the sealingcapability of gasketed flange joints subjected to inter-nal pressure at high temperature. Abid et al.5–8 alsoperformed 3D finite element modal analysis of gas-keted bolted flanged pipe joints under bolt-up.Gasketed flanged joint have been discussed in theliteratures9–11 under single loading. More recently,

Abid et al.12 performed the 3D nonlinear analysis ofnon-gasketed flange joint under combined internalpressure, axial loading and bending load. Jointstrength is evaluated at both the design pressure(DP) and proof test pressure (PT).

From previous studies,3,13–16 it is concluded thatthe key parameter responsible for the performanceof gasketed flanged joint is the bolt-up strategy.Improper bolt-up may result in failure of joint bothin term of structural strength and leakage. In thispaper, which is extension of previous study performedin the literatures,16–18 joint behavior is observed underbolt-up strategy recommended by the ASME guide-line19 as well as industrial bolt-up strategy which issuggested in the literature20 with addition to com-bined internal pressure plus axial loading. The studyis advancement in terms of bolt-up strategy(Industrial and ASME BUP) and in terms of

Proc IMechE Part E:

J Process Mechanical Engineering

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DOI: 10.1177/0954408915614460

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1Department of Civil Engineering, University of Malaya, Kuala Lumpur,

Malaysia2Interdisciplinary Research Center, COMSATS Institute of information

Technology, Wah Cantt, Pakistan3Department of Mathematics, COMSATS Institute of information

Technology, Lahore, Pakistan

Corresponding author:

Niaz B Khan, Department of Civil Engineering, University of Malaya,

Kuala Lumpur, Malaysia.

Email: [email protected]

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combined loading (Internal pressure plus axialloading) for the gasketed bolted flanged joint. It isalso observed that after improper bolt-up, the add-ition of external load on joint may result inworst condition, particularly combination of internalpressure and axial load have more pronouncedeffect.18

In this paper, the joint strength is examined underdifferent bolt-up strategies (ASME and industrial)plus combined internal pressure and axial load. Forthat purpose, 3-D finite element analysis (FEA) ofgasketed bolted flange joint is performed. Spiralwound gasket is used in this study. The model usedin current study is already verified in Abid andKhan.18 Table 1 shows the loading condition underwhich the flange joint is analyzed in this paper.Figure 1 depicts the dimensions of the model usedin current analysis. The dimensions are the same aspreviously used in Abid and Khan,18 except the pipewhich length has been increased to 300mm to studythe axial loading effects.

Finite element analysis

This paper is an extension to Abid and Khan,18 andtherefore all the geometrical dimensions of model areexactly the same. Boundary conditions, mesh gener-ation methodology, and element selection are also thesame as previously used. Since the model is symmetricin terms of boundary conditions and geometry, halfgasket (w.r.t thickness) is modelled. The bolt is alsomade by revolving the plane area pattern along itsshank axis in 360�. Only half of the bolt is modelledbecause of symmetry in terms of loading and geom-etry. The pipe of length 300mm is modelled withflange. Figure 2 shows the finite element model offlange, bolt and gasket.

Element selection

In addition to the contact stresses in spiral woundgasket, the required outputs from the analysis arestresses in flange, bolt, and pipe. In ANSYS,22

SOLID45 is a solid structural element used for struc-tural stress analysis of flange, pipe, and bolt in thejoint. It is defined by eight nodes with three transla-tional degree-of-freedom at each node. SOLID45

element has plasticity, creep, swelling, stress stiffening,large deflection, and large strain capabilities.

Contact elements are used to define contactbetween different surfaces of the model. In problemsinvolving contact between two boundaries, one of theboundaries is conventionally established as the‘‘target’’ surface, and the other as the ‘‘contact’’surface. CONTA174 (contact element) is used in com-bination with TARGE170 (target element) betweencontact of flange face and gasket, top of flange andbottom face of bolt, and bolt shank and inner face offlange hole. CONTA174 is defined by four nodes andis applicable to 3D structural contact analysis. It hasthe capability to simulate contact and sliding between3-D target surfaces.

For simulation of gaskets, 3-D linear interfaceelement defined by eight nodes, INTER195, isselected. INTER195 is used with 3-D structural elem-ent to simulate gasketed joint. The total number ofelements used in the finite element model of flange,pipe, bolts, and gaskets is 32,639. Commercial finiteelement analysis software ANSYS22 is used in currentstudy.

Material properties and boundaryconditions

In current study, bilinear kinematic hardening modelis used for elasto-plastic material properties, which ispreviously practiced in Abid.13This model has twolinear gradient section. In first linear gradient section,an elastic material is used which functions until theyield stress and Young’s modulus of elasticity is thegradient for this section. The second section is validbeyond the yield stress and the gradient (plasticmodulus) of this section is 10% of the Young’s modu-lus of elasticity.23 Table 2 shows the material proper-ties of flange, pipe, bolt, and spiral wound gasket,which are taken from the literatures.24,25 The spiralwound gasket consists of v-shaped metal strip andsoft non-metallic filler (asbestos fibre) under pressure.It has an inner and outer ring made of stainless steel.In this analysis, simplified modeling method of gasketnonlinearity for bolted flange joints is adoptedbecause of simplicity and cost.26

Figure 3 shows the boundary condition on theflange joint. The flange is permitted to move in the

Table 1. Operating conditions.

Sr. # Loading Type of loading Load

1 Bolt tightening Industrial strategy20 Torque¼ 505 Nm

ASME strategy28 Torque¼ 700 Nm

2 Internal Pressure Design pressure (DP) 15.3 MPa

Proof test pressure (PT) 23 MPa

3 Combined loading Design pressure plus axial load (DPþAL) DP¼ 15.3 MPa AL¼ 665 kN

Proof test pressure plus axial load (PTþAL) PT¼ 23 MPa AL¼ 665 kN

2 Proc IMechE Part E: J Process Mechanical Engineering 0(0)



Figure 1. Joint dimensions: (a) pipe flange and bolts, and (b) spiral wound gasket (all dimensions in mm).

Figure 2. (a) Flange, (b) gasket and (c) bolt.

Khan et al. 3



z-direction which provides flange rotation and actualbehavior of stresses in the gasket, bolt, and flange. Onthe gasket bottom face, symmetry condition isapplied. The bolt is constraint in the radial direction.To achieve pre-stress in the bolt, a displacement con-straint in the y-direction is applied on the lower faceof the bolt shank. For internal pressure loading, aload is applied on inner face of the pipe and flange.

Furthermore, endcap loading is applied across thewall of the pipe end.

Figure 4 shows the loading sequence of BUP, IPand axial loading in the current analysis. It is depictedfrom the Figure 5 that first of all, the bolt is tightenedup as per ASME or Industrial strategy. After bolt-up,internal pressure is applied on the inner surface ofpipe. After BUP and IP, axial load is applied at the

Figure 4. Loading sequence.

Figure 3. Boundary conditions for combined internal pressure plus axial loading.

Table 2. Material properties of flange, bolt and shank.

Parts As per code E (MPa) u Allowable stress (MPa)

Flange and pipe ASTM A350 LF2 173,058 0.3 248

Bolt ASTM SA193 B7 168,922 0.3 724

Gasket (Spiral wound) ASTM A182 164,095 0.3 207




end of pipe surface to study the combined effect ofloading.

Bolt preload and tightening sequence

Bolt preloads and tightening sequence are themajor parameters responsible for joint strength andperformance.

In this work, the methodology used for bolttightening sequence is the same as previously used inKhan et al.16 All the bolts are tightened according toASME strategy19 and industrial recommendation20

using torque control method15,27 as shown inTable 3. In order to achieve required pre-stress, a cer-tain displacement (UY) is applied at the lower surfaceof the bolt.

Figure 6. Nomenclature of selected nodes on bolt shank.

Figure 5. Numbering of bolts (a) sequence-1 and (b) sequence-2.14

Table 3. Target stress at end of each pass.

Industrial BUP strategy ASME BUP strategy

Applied

torque (Nm)

Bolt

preload (kN)

Target

stress (MPa)

Applied

torque (Nm)

Bolt

preload (kN)

Target

stress (MPa)

210 37 57 140 24 38

310 54 86 420 73 115

400 70 112 700 122 191

505 89 145 –

Khan et al. 5



In ASME BUP strategy,19 bolts are tightened withincrement of torque 140, 420 and 700 Nm, according tosequence-1 (1, 5, 3, 7, 2, 6, 4, 8) (Figure 5(a)), and inlast pass all the bolts are tightened again to 700 Nmaccording to sequence-2 (clockwise) (Figure 5(b)). Inindustrial BUP strategy, bolts are tightened with incre-ment of torque 210, 310, 400, and 595 Nm according tosequence-1 and in last pass all bolts are tightened againto the 595 Nm according to sequence-2.

Results and discussion

Bolt stress variation

Results for stress variation along the bolt nodes (asper nomenclature shown in Figure 6) at bolt plusinternal pressure plus axial loading are show inFigure 9. It is observed that more bolt bending behav-ior is found during ASME BUP strategy in compari-son with industrial bolt tightening strategy.

At ASME BUP, bolt stress variation of 200–250MPa between inner and outer nodes is observedin almost all the bolt except in bolt 6 where 150–170MPa stress variation is observed (Figure 9).

Lower stress variation (20MPa) in bolt-6 shows lessgasket contact stress in the vicinity of bolt 6. Overall,more scatter stresses in ASME bolt-up strategy isobserved in comparison with industrial bolt-upstrategy.

After BUP, with addition of external combinedloading, i.e. internal pressure plus axial loading, it isobserved that bolt stress variation increased linearly.Almost the same pattern is observed in all bolts, i.e.stresses are increased resulting in more stress variationin all the bolts. In the case of ASME case study, boltstress variation of 300–370MPa between inner andouter nodes is observed in bolts 2 and 8, where inbolts 4 and 6 stress variation is 250–300MPa(Figure 9). It is observed from Figures 7 and 8 thatthe maximum stress is at the corner of bolt head andshank. The exaggerated views show the bolt bendingbehavior. It is also observed that stresses in all thebolts are within the allowable stress limits (723MPa)at the combined DP and PT plus axial loading (100–665 kN). In the case of Industrial BUP plus IP plusAL, the same behavior is observed with linear increasein stresses. Figure 9 shows the bolt bending behaviorof four bolts for ASME and Industrial case study.

Figure 7. Exaggerated bolt stress variation at (a) ASME BUP þ DP, (b) ASME BUP þPT, (c) ASME BUPþDPþAL and (d) ASME

BUPþ PTþAL.




Hub flange fillet

In ASME case study, maximum stress (S1 and SINT)at hub flange fillet increased from 350MPa and289MPa to 374MPa and 304MPa at design pressureand to 378MPa and 314MPa at PT pressure plus. Itis found that maximum principal stress is more thanthe ultimate stress of flange material at DP and PTpressure, whereas the bending stress (SY) increasedfrom 245–250MPa to 272–276MPa at additionalaxial load of 665 kN to DP and PT, which ismore than the allowable stress of flange material(248MPa) but less than ultimate strength of theflange material (370MPa). Stresses observed duringindustrial bolt-up strategy with additional combinedloading are less than ASME bolt-up plus combinedloading.

From the above discussion, it is concluded thatstresses are higher than the ultimate stress limit atcombined internal pressure and axial loading. It isalso concluded from the above discussion thathigher bolt-up is better for improving sealing capabil-ity, but it results in higher stresses at hub flange filletwhich reduces its strength.

Gasket contact stress variation

The nomenclature of gasket model is shown in Figure10. G1 shows the gasket node in front of bolt1, G2show the gasket node in front of Bolt2, and so on. Theinner and outer sealing ring has been defined in orderto study the variation along the gasket width.

Figure 11 shows the contact stress variations forASME and Industrial cases along the outer andinner sealing ring under additional axial loading tothe design pressure (Figure 11(a) to (d)) and prooftest pressure (Figure 11(e) to (h)). Gasket contactstress pattern remains the same with additionalaxial loading. However, reduction in stresses isobserved in both case studies under axial loading inaddition to the internal pressure resulting in gasketrelaxation.

In ASME case study, along the outer sealing ring,it is observed that along G1–G3 locations, higherstresses are observed compared to G5–G7 location.Minimum contact stress of (97–100MPa) is observedalong G5–G7 which is more than the seating stress(68MPa) recommended by the gasket supplier.28

Hence, it is concluded that at the outer sealing ring,

Figure 8. Exaggerated bolt stress variation at (a) industrial BUPþDP, (b) industrial BUPþ PT, (c) industrial BUPþDPþAL, and (d)

industrial BUPþ PTþAL.

Khan et al. 7



there is no chance of leakage and the joint is safeunder additional axial loading.

Along the inner sealing ring, more scatter isobserved along the gasket locations G1–G4, whereasalong the G4–G8 location, the stress variation issmall. Minimum stress is observed at G6 locationwhere the gasket contact stress reduces from�89MPa at bolt-up to -63MPa and 62MPa at DP

and PT pressure plus axial loading of 665 kN, respect-ively. At additional axial load of 665 kNto DP and PT, the gasket contact stress alongG5–G7 locations is less than the recommendedseating stress along the inner location only. It is con-cluded from the above discussion that the joint is safeat additional axial load of 665 kN to DP and PTpressure.

In case of Industrial case study, FEA results showsthat at internal pressure (DP and PT pressure), gasketcontact stress along G4–G7 outer sealing ring is 48–53MPa and 44–50MPa, respectively, which is lessthan recommended seating stresses (68MPa).28

Additional axial load of 665 kN with DP and PTpressure further decreased the gasket contact stressto 26–32MPa and 25–30MPa, respectively, alongthe G4–G7 outer sealing ring providing more chanceof leakage. Although the contact stresses along G4–G7 is not zero, but very less than recommendedseating stress (68MPa), it provides some chances ofleakage. It is also observed that gasket contact stressis less than the seating stress along all locationsaround the gasket, at axial load of 525 kN in additionto the DP and at axial load of 400 kN in addition tothe PT pressure.

0

100

200

300

400

500

BU

PD

P10

052

566

5

BU

PD

P10

052

566

5

BU

PD

P10

052

566

5

BU

PD

P10

052

566

5

BU

PPT 10

052

566

5

BU

PPT 10

052

566

5

BU

PPT 10

052

566

5

BU

PPT 10

052

566

5

Stre

ss (

MP

a)

B2 B4 B6 B8 B2 B4 B6 B8

0

100

200

300

400

500

BU

PD

P10

052

566

5

BU

PD

P10

052

566

5

BU

PD

P10

052

566

5

BU

PD

P10

052

566

5

BU

PPT 10

052

566

5

BU

PPT 10

052

566

5

BU

PPT 10

052

566

5

BU

PPT 10

052

566

5

Stre

ss (

MP

a)

B-2/1 B-2/2 B-2/3 B-2/4 B-4/1 B-4/2 B-4/3 B-4/4

B-6/1 B-6/2 B-6/3 B-6/4 B-8/1 B-8/2 B-8/3 B-8/4

B2/1 B2/2 B2/3 B2/4 B4/1 B4/2 B4/3 B4/4

B6/1 B6/2 B6/3 B6/4 B8/1 B8/2 B8/3 B8/4

(a)

(b)

Figure 9. Bolt stress variation: (a) ASME case study and (b) industrial case study.

Figure 10. Gasket nomenclature.




Conclusions

1. Gasketed joint load capacity is concluded betterusing ASME bolt-up strategy compared to theindustrial bolt-up strategy especially in terms ofsealing capability. ‘‘No leakage’’ criteria isobserved in the joint even with addition of com-bined loading application, while in terms of struc-tural strength the joint capability is adverselyaffected. It is also concluded that higher bolt-uptorque value resulted in weak structural strength.

2. Lower bolt-up in industrial bolt-up strategy isconcluded better for structural strength of thejoint, but there is chance of leakage.Furthermore, the addition of combined internalpressure and axial load results in worst scenarioin terms of sealing capability, where the stressesobserved in majority of gasketed portion are muchlower than the recommended seating stresses.

Declaration of Conflicting Interests

The author(s) declared no potential conflicts of interest with

respect to the research, authorship, and/or publication ofthis article.

Funding

The author(s) disclosed receipt of the following financial

support for the research, authorship, and/or publicationof this article: This research was supported by UniversityMalaya Research Grant (UMRG – Project No. RP004E-

13AET), University Malaya Postgraduate Research Fund(PPP – Project No. PG102-2014B).

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-140

-120

-100

-80

-60G1 G2 G3 G4 G5 G6 G7 G8

Gas

ket

Stre

ss (

MP

a)

Outer Sealing

BUP

DP

100

525

665

-120

-100

-80

-60G1 G2 G3 G4 G5 G6 G7 G8

Gas

ket

Stre

ss (

MP

a)

Inner Sealing

BUP

DP

100

525

665

(a) (b)

Figure 11. Gasket contact stress variation in ASME BUP study [(a), (b), (c) and (d)] and Industrial BUP study[(e), (f), (g) and (h)].

Khan et al. 9



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Nomenclature

AL axial load (kN)BUP bolt-upDP design pressure (15.3MPa)IP internal pressure (MPa)PT proof test pressure (23MPa)SINT stress intensity (MPa)SWG spiral wound gasketSY stress in axial (Y) direction (MPa)S1 principal stressUY displacement in axial direction (mm)



www.klinger.co.uk


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