artigo v - comparação entre tipos de juntas.pdf

Comparative study of the behaviour of conventional gasketed and compact

non-gasketed flanged pipe joints under bolt up and operating conditions

M. Abida,*, D.H. Nashb

aFaculty of Mechanical Engineering, Ghulam Ishaq Khan Institute of Engineering Sciences and Technology, Topi-23460, NWFP, PakistanbDepartment of Mechanical Engineering, University of Strathclyde, Glasgow, UK

Received 11 November 2002; revised 11 October 2003; accepted 10 November 2003

Abstract

Bolted flanged joints comprise an assembly of a number of important individual components, which are required to perform well together

in service. The ideal requirement for a bolted flange joint is a zero-leak condition. However, whilst recommended design procedures for

bolted flange joints are available in international codes and standards, leakage problems are still faced by industry. These are common in both

normal operating (internal pressure loading) and critical event conditions. The drive is, therefore, to find a flange joint assembly, which

provides zero-leak condition and requires little or no maintenance and handling. Considerable investigation in the area of optimised bolted

joints has been in progress for the past 10 years comparing traditional gasketed joints and compact non-gasketed joints, using both

analytical and experimental approaches. In this present study, two-dimensional non-linear finite element studies have been performed for

both gasketed and non-gasketed bolted flange pipe joints. Based on the stress results for the flange and the bolt and the flange

rotation/displacement, compact non-gasketed flange joints are shown to be a viable and preferable alternative to the conventional gasketed

flange joints. Recommendations are made for a best-fit flange model for static load conditions with zero-leak sealing in a flange joint.

q 2004 Elsevier Ltd. All rights reserved.

Keywords: Gasketed; Non-gasketed; Leakage; Flange joints; Finite element

1. Introduction

All modern international design codes and standards

provide calculation rules for the design of the flanges, which

are principally based on the TaylorForge method [1,2].

Flange joints designed in accordance with international

codes such as ASME [3], PD [4], and CEN [5] experience

leakage and this problem is continuously faced by industry.

In the present study, detailed analysis of flange joints is

undertaken in order to observe the actual stress behaviour in

flange, bolts and gasket. The reason for such a study is that a

bolted flange joint is a combination of different elements

which are inter-linked with each other to perform as a unit

for zero-leak condition. Mathematically, the zero-leak

condition is taken to be zero flange displacement of the

flange at the inside diameter of the mating flanges in the

flange joint. In conventional flange analysis, no consider-

ation has been given to this important issue, i.e. the flanges,

bolts and gasket are discussed and are treated entirely

separately. Therefore, with the availability of modern

analysis tools, it is possible to adopt a design-by-analysis

approach for the bolted flange joint in order to provide

accurate values for the stresses in key areas as required; for

example, at hub-flange fillets, in the bolts and in the gasket

element itself.

In addition for conventionally designed flanges, where

there is a measure of inconsistency of the dimensions

between flange sizes, it has been difficult to standardise on a

single procedure to overcome the stress and deflection

design versus leakage problem. Therefore, by means of

comparison to the conventional flange design, a new novel

flange design is also studied which is dimensionally very

compact, is light weight and is without a gasket. This is

designated the non-gasketed bolted joint. Stresses in the

flange ring, bolts and gasket, various displacements and bolt

bending results are studied for both the gasketed and non-

gasketed flange joints.

Previous work by Power [6] used a full three-dimen-

sional (3D) non-linear analysis for comparative study of

0308-0161/$ - see front matter q 2004 Elsevier Ltd. All rights reserved.

doi:10.1016/j.ijpvp.2003.11.013

International Journal of Pressure Vessels and Piping 80 (2003) 831841

www.elsevier.com/locate/ijpvp

* Corresponding author. Tel.: 92-938-71858~61; fax: 92-938-71889.E-mail address: [email protected] (M. Abid).

three different types of joints for the internal pressure

loading case only. Using 3D modelling the variation of

contact pressure and stress behaviour in the flange and the

bolts was studied. The contact, or interface pressure was

used as the main quantitative measure of the sealing ability

of the three joint styles. More recently, Cao [7] also

performed 3D FEA considering gasket non-linearity under

both bolt up and operating conditions and studied the joint

strength and sealing capability. Similarly, Fukuoka [8] has

studied the bolt stress scatter, effect of tightening sequence,

bearing surfaces and number of tightening passes. However,

flanged pipe joints have already been modelled using 2D

finite element analysis by Hyde et al. [9], Shoji [10] and by

one of the present authors [11].

In the present study, in order to compare the two flange

styles and their relative performance, a flange size of 4 in.,

900# class was selected. 2D axi-symmetric non-linear finite

element analysis for both the joints was carried out for

internal pressure loading only. During the analysis, the

intention was not to study the variation in contact pressure

and contact area but rather to study the stress behaviour in

the flange and the bolts, especially the effect of bolt bending,

bolt pre-load and flange displacement/rotation. It is noted

that variation in contact pressure and contact area is

theoretically of importance but any deformation on the

flange surface, specifically for the metal-to-metal contact

non-gasketed flange joint, cannot stop leakage even with a

very high contact pressure. This has been noted during

published experimental tests [11]. However, it is recognised

that the contact pressure study is of course important from

the design point of view, in order to take account of bolt

behaviour and the contact force of the flanges, but will not

be reported herein.

2. Allowable stresses and flange joint configuration

2.1. Allowable stresses

The yield stress of the flange and pipe material selected

was 372 N/mm2, giving a nominal design stress of

248 N/mm2. High strength bolts of 30 and 10 mm were

used in the gasketed and non-gasketed joints. Material

properties for flange, pipe, bolts and gasket selected are

given in Table 1.

2.2. Flange geometry

The configurations under examination and parameters

used for modelling are shown in Fig. 1a and b for the

gasketed and non-gasketed joints.

Gasketed flange. For the flange and the gasket geometry,

the dimensions were obtained from BS1560 [12]. A washer

is not included in the present work, since for gasketed joints;

it is not a mandatory requirement of the PD or ASME code.

Gasket was modelled as a solid ring for the gasketed joint. In

actual practice during the bolt up condition, it is necessary to

compress the raised sealing part down to equal to

Nomenclature

A outside diameter of the flange

B inside diameter of the flange

C bolt circle diameter

g0 wall thickness of basic shell/pipe

h taper-hub length

t non-gasketed flange thickness

jh joint height

pid pipe inside diameter

pod pipe outside diameter

wth pipe thickness

sh shoulder height

sod shoulder outside diameter

hubht hub height

hod hub outside diameter

gth gasket thickness

gid gasket inside diameter

god gasket outside diameter

gcringid gasket centring ring inside diameter

gcringod gasket centring ring outside diameter

gsrght gasket seal ring height

bcd bolt circle diameter

fod flange outside diameter

bd bolt diameter

bhd bolt head diameter

bhh bolt head height

f element contact stiffness factor

Table 1

Material properties of gasketed and non-gasketed flange joints

Part Gasketed flange joint Non-gasketed flange joint

E(N/mm2)

y Design stress(N/mm2)

As per code E(N/mm2)

y Design Stress(N/mm2)

As per code

Flange and pipe 173,058 0.3 248.2 2=3sy ASTM A350 LF2 or A105 203,395 0.3 248.2 2=3sy ASTM A350 LF2 or A105Bolt 168,922 0.3 723.9 sy ASTM SA193 B7 204,000 0.3 640.0 sy ISO 898, Grade 8.8Gasket 164,095 0.3 206.8 2=3sy ASTM A182 F316

M. Abid, D.H. Nash / International Journal of Pressure Vessels and Piping 80 (2003) 831841832

the centring ring thickness in order to energise the proper

sealing capability of the gasket as recommended by the

gasket suppliers. During the previous experimental studies

[11], an inconsistency in the gasket outside diameter

dimension were noted and the outside diameter of the

gasket was subsequently machined to allow fit up into the

flange joint. Flange parameters used are given in Table 2.

Non-gasketed flange. A washer with the bolt is

recommended for the non-gasketed joint but it was observed

in preliminary studies that it does not have a significant

effect on stress and deformation results for the axi-

symmetric model. It was, therefore, decided not to include

the washer in the analytical model. For the non-gasketed

joint, to achieve better joint strength and sealing ability, a

higher initial pre-load and a positive taper angle of 0.038

on the flange surface is applied. Flange parameters used are

given in Table 3.

3. Finite element modelling

Previous work by Spence et al. [13] and Nash et al. [14],

showed the viability of a 2D axisymmetric model, for what

is essentially a 3D component. In neglecting the holes in the

flange and the presence of individual bolts round the flange,

the system can be considered as a continuous bolt ring

located at the bolt centre and running round the circumfer-

ence of the bolt circle. Schematic quarter models of both

flange joints are shown in Fig. 2a and b. The ANSYS finite

element code v5.7 was employed throughout this work.

Fig. 1. Parameters used in the finite element model of flange joints (a) gasketed, (b) non-gasketed.

M. Abid, D.H. Nash / International Journal of Pressure Vessels and Piping 80 (2003) 831841 833

Fully parametric finite element models were used through-

out so that the time involved in building the scaled geometry

models of other different sizes is minimised.

3.1. Element selection

Since the flange, bolt and gasket stresses were the

required outputs from this study, it was necessary to use two

classes of element: solid element PLANE82 to model the

solid entities and contact element CONTACT48 to measure

contact pressure or stress variation between the mating

surfaces. Contact stiffness (KN) for CONTACT48 element

quantifies the level of penetration that the target surface

allows. Over penetration results in inaccuracy if it is too

small a value and non-convergence of solution results if it is

too large a value. From the earlier study by Spence et al.

[14], it is noted that only a lower value of f ; the element

contact stiffness factor, affects the maximum axial stress at

the intersection of the taper hub and the vessel. When

varying the value of f between 0.1 and 100, at the value of

f $ 1 stress was almost levelled out, and therefore, a valuewas selected of f 10 for further analysis. The effect of thevariation of contact stiffness on stress is shown in Fig. 3.

3.2. Mesh refinement

In order to ensure accurate results, a reasonably refined

mesh was applied in the regions of interest such as at the

hub-pipe and hub-flange intersections. The mesh for the

gasketed flange was also refined at the flange shoulder to

measure the contact stresses at the gasket. Twenty-one and

seven elements across the pipe and hub thickness, 10 and 13

elements across the flange thickness and 10 and 12 elements

across the bolt width were defined for the gasketed and non-

gasketed flange joints. Illustrations of the meshed models

are shown in Fig. 4a and b.

3.3. Boundary conditions

The boundary conditions applied for both the joints are

shown in Fig. 4. Internal pressure is applied at the inside

diameter, and the loading due to the head is directly applied

as nodal forces across the wall of the pipe which are

obtained using the internal axial force Ppr2i divided by 2prt(i.e. Pr2i =2rt). The flanges are free to move in either theaxial or radial direction and rotate, providing the exact

behaviour of the stresses in the flange and the bolt.

Symmetry planes are constrained in both the axial and

radial directions. For the gasketed joint, due to the gap

between the flanges, internal pressure is also applied on

the gasket inside diameter as well as on the flange shoulder

in the gap. The gasket is constrained in the radial direction

to prevent it moving radially since, in practice, it is located

interior to the bolts. The bolts for both flanges are

constrained at the mid diameter in the radial direction as

Table 3

Non-gasketed flange joints geometric parameters

Flange hub

length, h (mm)

Pipe thickness,

g0 (mm)

Joint height,

jh (mm)

Bolt circle

dia., C (mm)

Flange outside

dia., A (mm)

Flange inside

dia., B (mm)

Flange thick.,

t (mm)

Flange surface

angle, taper (deg)

Bolt dia.,

bd (mm)

34 13.5 102.6 146 171 87.3 30 0.03 10

Fig. 2. Schematic quarter models for (a) gasketed, (b) non-gasketed flange

joints, showing bolts as continuous bolt ring (as used in axi-symmetric

representation).

Table 2

Gasketed flange joints geometric parameters

Flange parameters

Flange hub length,

hubht (mm)

Flange joint

height, jh (mm)

Bolt circle dia.,

bcd (mm)

Flange outside

dia., fod (mm)

Flange inside

dia., pid (mm)

Flange thick.,

fh (mm)

Flange hub

dia., hod (mm)

Shoulder height,

sh (mm)

Shoulder outside

dia., sod (mm)

69.9 71.67 235 292 87.3 44.4 159 6.4 157.2

Gasket, pipe and

bolt parameters

Gasket inside

dia., gid (mm)

Gasket outside

dia., god (mm)

Gasket thick.,

gth (mm)

Gasket seal

ring inside dia.,

gcringid (mm)

Gasket seal

ring outside dia.,

gcringod (mm

Pipe inside

dia., pid (mm)

Pipe outside dia.,

pod (mm)

Pipe thick.,

wth (mm)

Bolt dia.,

bd (mm)

120.65 149.35 4.5 106.43 206.50 87.3 114.3 13.5 30


they cannot move in this direction within the bolt hole. An

axial displacement is applied on the bottom nodes of the bolt

shank which was progressively optimised by sequential

iteration, to obtain the required pre-stress.

3.4. Initial bolt up and operating (internal pressure) loads

A nominal pre-load of about 35% (245 N/mm2) of

the yield stress of the bolt (723 N/mm2) was chosen for

the gasketed joint as per the achieved maximum strain

in the bolt at the applied torque of 505 N m during the

experimental study [11]. The ASME code does not

specify a magnitude of pre-load for the bolts, rather only

a minimum seating stress that relates to the gasket style

and composition. In addition, in general practice, the

pre-load for the gasketed joint is based partially on the

practical knowledge that most fitters of flanged joints

manually tighten the bolts as hard as possible. For the

non-gasketed joint a pre-load approximately 80% of the

yield stress of the bolt [11] was applied for better joint

strength due to the fact that the higher the pre-load the

better the joint is and is easily achieved for metal-to-

metal contact flanges due to no gasket stress limitation.

As the yield strength of the bolt material is 640 N/mm2,

so the applied bolt pre-stress is 512 N/mm2. This stress

is based on the bolt shank area of 78.54 mm2, whereas

the bolt strength is taken as per thread root area

(58 mm2), resulting in pre-stress to be applied was

378 N/mm2.

After the bolt up, an internal pressure of 15.3 N/mm2

was applied as per the pressure rating recommended for

the flange size of 4 in., class 900# as defined by the

codes. This applied pressure is the maximum permissible

working pressure, whereas, in order to compare the stress

pattern or structural integrity of both the flange joints,

the pressure required was raised to the proof test pressure

of 23 N/mm2 and thereafter to a maximum value of

40 N/mm2 (i.e. 2.6 times the normal operating pressure

of 15.3 N/mm2).

4. Results and discussion

4.1. Comparison of calculated and FEA results

To verify the models, the main pipe wall axial and hoop

stresses were compared with the calculated stress results and

were found in good agreement.

Fig. 3. Effect of contact stiffness on maximum axial stress (N/mm2).

Fig. 4. Element plots for flange joints (a) gasketed, (b) non-gasketed with

enlarged hub portion.


4.2. Comparison of results for both the joints

4.2.1. Non-gasketed flange joint

With reference to Fig. 5ac and Table 4, the following

observations may be made. At the maximum allowable

working pressure of 15.3 N/mm2, the maximum stress

intensity in the flange is less than the allowable stress of

the flange material. The stress is higher at the hub-flange

fillet, than at the inside diameter at the hub. A high stress

is also observed at the area under the bolt head, due to the

high-applied pre-load in the bolt. The maximum principal

stress is less than the allowable of the flange material even

at a pressure of 2.6 times the maximum permissible

working pressure. The contact stress variation between the

two flanges is very small from the inside diameter (a

value of 19 N/mm2) to the outside diameter (a value of

35 N/mm2) with the applied internal pressure. The rotation

from inside to outside diameter of the non-gasketed flange

is almost negligible at the maximum permissible, the

proof test and the maximum applied pressures, and is

recorded in Table 4.

4.2.2. Gasketed flange joint

With reference to Figs. 6a,b and 7a,b and Table 4, the

following observations were made. For the gasketed flange

joint, even during pre-loading, stress intensity and principal

stress is significantly more than the allowable of the flange

material at the hub-flange fillet. The application of internal

pressure added to this effect resulting in increase in stress.

This indicates yielding at hub-flange fillet during both bolt

up (pre-loading) and operating (pressure) loading (Fig. 6a

and b). At other locations, i.e. hub-pipe and hub-centre, the

resulting stress was less than the allowable.

It is also important to note that the local yielding is more

than 20% of the hub-flange fillet depth, Fig. 7a. The

maximum stress at the inside diameter of the hub is also

higher than the allowable of the flange material. In addition,

a high stress is present at the edge of the shoulder, which

acts as the pivot point due to the flange rotation. The higher

stresses in this region are possibly due to the geometric

stress concentration and idealised finite element modelling

(Fig. 7a). From Fig. 7b and c stress intensity plots, overall

stress variation in the flange and bolt is obvious. The rotation

Fig. 5. Non-gasketed flange joint, (a) stress intensity plot for full joint at internal pressure of 15.3 N/mm2, (b) stress intensity plot for flange only at internal

pressure of 15.3 N/mm2, (c) maximum principal stress plot at internal pressure of 40 N/mm2.

Table 4

Non-gasketed flange joint, maximum principal stress and flange rotation results

Internal Pressure (N/mm2) Maximum principal stress (N/mm2) Flange rotation (degree)

Gasketed flange joint (mm) Non-gasketed flange joint (mm) Gasketed flange joint (mm) Non-gasketed flange joint (mm)

15.3 309.59 94.64 0.1589 0.0300

23.0 316.62 100.31 0.3085 0.0307

40.0 321.66 137.95 0.3131 0.0424


of the gasketed flange joint is proportional to the applied

internal pressure. For the gasketed flange joint, flange

rotation means the flange portion between

the shoulder outside diameter and the flange outside

diameter. Shoulder rotation means the rotation of the

shoulder portion only.

4.3. Bolt stress behaviour of joints

4.3.1. Non-gasketed flange joint

With reference to Fig. 8a,b and Table 5, the following

may be concluded. There is high stress at the top of the bolt

shaft close to the bolt head and the shank corner. However,

Fig. 6. Gasketed flange joint, (a) principal stress plot after pre-loading or bolt up, (b) stress intensity (exaggerated) plot for flange at a pressure of 15.3 N/mm2.

Fig. 7. Gasketed flange joint, (a) stress intensity plot showing the location and the depth at which the yielding starts at a pressure of 15.3 N/mm2, (b) stress

intensity plot for full joint model, (c) stress intensity plot for flange only at internal pressure of 15.3 N/mm2.


no yielding resulted and the maximum stress in the bolt is

less than the allowable stress of the bolt material. From

the stress variation in the bolt, the bolt bending is very

small, due to the very small flange rotation with applied

internal pressure.

4.3.2. Gasketed flange joint

With reference to Fig. 9a,b and Table 5, similar to the non-

gasketed joint bolt, a high stress occurred at the corner of the

bolt head and the bolt shank with no yielding. However, bolt

bending is obvious for the gasketed joint due to a difference

of stress of about 200 N/mm2 at the inside and outside nodes.

This difference is a maximum of 40 N/mm2 for the non-

gasketed joint, which is very small compared to the values

obtained for the gasketed joint. This stress variation in the

bolt shows the flange rotation during pre-loading (bolt up)

and operating (pressure) loads.

4.4. Contact stress behaviour for gasketed and non-gasketed

joints

Due to the flange rotation, the contact pressure on the

gasket varies from the gasket inside to the outside diameter.

The gasket was modelled as a solid ring. The deformed

shape of the gasket is shown in Fig. 10 and the gasket

contact stress variation is shown in Fig. 11. From the high

stress at the seal ring outside diameter, the flange rotation is

again obvious and shows that during the bolt-up, the seal

ring can be damaged for any applied higher pre-load. As

discussed earlier, it is important for the gasketed joint to

achieve the proper seating stress recommended by the gasket

Fig. 8. Non-gasketed flange joint bolt only, (a) stress intensity plot for the full joint due to the operating load, (b) stress variation in the bolt due to the operating

load (zoomed bolt portion only).

Table 5

Bolt stress variation for non-gasketed and gasketed flange joints with the variation of pressure

Internal pressure

(N/mm2)

Stress in bolts

(N/mm2)

Non-gasketed flange joint Gasketed flange joint

Inside diameter

(mm)

Mid diameter

(mm)

Outside diameter

(mm)

Inside diameter

(mm)

Mid diameter

(mm)

Outside diameter

(mm)

15.3 392.20 374.13 356.06 343.65 242.87 141.86

23.0 391.88 373.10 354.30 346.86 243.17 141.86

40.0 402.26 376.38 350.46 347.50 243.82 139.93


supplier for the no leak condition. It is obvious from the

gasket stress variation that a very low stress value is

achieved on the contact area between the flange and the

gasket. In addition, in axisymmetric modelling, the gasket is

modelled as per the ideal required condition.

In Fig. 12, the stress variation at the contact surface with

the symmetry plane for the non-gasketed joint is shown

which increased gradually from the inside to the outside

diameter during the application of internal pressure.

5. Conclusions

In view of the stress distributions resulting in the flange

and bolts together with the observed flange rotation, bolt

bending, bolt fatigue and bolt relaxation it is concluded that

a dynamic behaviour in the gasketed flange joint is present

during both the bolt-up and operating conditions. This is

primarily due to the inclusion of a flexible gasket element

between the flanges leading to a continually varying

situation where, when bolt or pressure load changes, the

flange faces move and rotation results in response. This

provides permanent damage in the flange and repeated use

of this joint cannot ensure its performance for zero-leak,

even though the joint may be well assembled. Secondly, due

to limitations of the choice of gasket seating stress to avoid

crushing, a higher initial stress cannot be applied in the

bolts, and therefore, the available strength of the high yield

bolts cannot be utilised.

In comparison, stresses observed in the non-gasketed

flanged joint are within the specified allowable levels even

up to an internal pressure of 2.5 times design pressure.

Fig. 9. Gasketed flange joint bolt only, (a) stress intensity plot for the full joint due to the operating load, (b) stress variation in the bolt due to the operating load

(zoomed bolt portion only).

Fig. 10. Gasketed flange joint, deformed gasket plot.


Fig. 11. Gasketed flange joint, gasket stress variation plot along the gasket, after the operating load.

Fig. 12. Non-gasketed flange joint, stress variation at the symmetry plane (contact surface).


A non-gasketed metal-to-metal contact joint without a

separating gasket is concluded as having static mode of

load due to no significant movement of the flange faces with

a change in bolt or pressure load. Due to this no bolt bending

and so no bolt fatigue results, therefore, it is considered a

safe joint. Finally due to the compact dimensions, it is

comparatively 10 times lighter than the conventional

gasketed flange joint, resulting in easy handling and joint

assembly.

Hence, as a result of the detailed finite element study

presented herein, compact non-gasketed flange joints may

readily be considered as a viable alternative to the

conventional gasketed flange joints presently in use.

References

[1] Waters EO, Taylor JH. The strength of pipe flanges. Trans Mech

Engng 1927;49:53142.

[2] Waters EO, Wesstrom DB, Rossheim DB, Williams FSG. Formulas

for stresses in bolted flanged connections. Trans ASME 1937;59:

1617.

[3] American Society of Mechanical Engineers. ASME boiler and

pressure vessel code. Section VIII, Division 1. App. Y, 2001 Edition.

[4] PD5500:2000. Unfired fusion welded pressure vessels. British

Standards Institution, London.

[5] EN 13445. European unfired pressure vessels, CEN; 2002.

[6] Power DJ. A study of conventional and unconventional flanged pipe

joint styles using non linear finite element analysis techniques. MPhil

Thesis; 1997.

[7] Cao D, Xu H. 3-D finite element analysis of bolted flange joint

considering gasket non-linearity. Pressure Vessel Piping Conf Boston

1999;382:1216.

[8] Fukuoka T, Takaki T. Evaluations of bolt-up sequence of pipe flange

using three dimensional finite element analysis. Pressure Vessel

Piping Conf Boston 1999;382:2135.

[9] Hyde TH, Lewis LV, Fessler H. Bolting and loss of contact between

cylindrical flat-flanged joints without gaskets. J Strain Anal 1988;23:

18.

[10] Shoji Y, Nagata S. A new analysis method for flange-gasket systems.

ASME Pressure Vessel Piping Conf Boston 1999;382:11320.

[11] Abid M. Experimental and analytical studies of conventional

(gasketed) and unconventional (non gasketed) flanged pipe

joints (with special emphasis on the engineering of joint strength

and sealing). PhD Thesis. University of Strathclyde, Glasgow;

2000.

[12] British Standard BS1560: Section 3.1. Circular flanges for pipes,

valves and fittings. BSI London; 1989.

[13] Spence J, Macfarlane DM, Tooth AS. Metal-to-metal full-face

taper-hub flanges: finite element model evaluation and preliminary

plastic analysis results. Proc Inst Mech Engrs 1998;212(Part E):

5769.

[14] Nash DH, Spence J, Tooth AS, Abid M, Power DJ. A parametric study

of metal-to-metal full face taper-hub flanges. Int J Pressure Vessel

Piping 2000;77:7917.


Comparative study of the behaviour of conventional gasketed and compact non-gasketed flanged pipe joints under bolt up and operIntroductionAllowable stresses and flange joint configurationAllowable stressesFlange geometry

Finite element modellingElement selectionMesh refinementBoundary conditionsInitial bolt up and operating (internal pressure) loads

Results and discussionComparison of calculated and FEA resultsComparison of results for both the jointsBolt stress behaviour of jointsContact stress behaviour for gasketed and non-gasketed joints

ConclusionsReferences

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