bolted flange connection

7/27/2019 Bolted Flange Connection

1/6

MIT International Journal of Mechanical Engineering Vol. 1, No. 1, Jan 2011, pp 35-40

ISSN No. 2230 7699 MIT Publications 35

3-D Finite Element Analysis of Bolted Flange Joint of

Pressure Vessel

Nomesh Kumar, P.V.G. Brahamanandam and B.V. Papa RaoAdvanced Systems Laboratory,Defence Research and Development Organisation,

Ministry of Defence, Hyderabad, Andhra Pradesh, India

Abstract- The objective of this work is to find out the stressesin the bolts of the bolted flange joint of the pressure vessel so

that bolts/studs should not be failed during proof pressure test.

Bolted flange joints perform a very important structural role

in the closure of flanges in a pressure vessel. It has two

important functions: (a). to maintain the structural integrity of

the joint itself, and (b). to prevent the leakage through the

gasket preloaded by bolts. One flange is having a groove for

the gasket, and other flange is flat connected by a number of

bolts/studs. The preload on the bolts is extremely important for

the successful performance of the joint. The preload must be

sufficiently large to seat the gasket and at the same time not

excessive enough to crush it. The flange stiffness in conjunction

with the bolt preload provides the necessary surface and the

compressive force to prevent the leakage of the gases contained

in the pressure vessel. The gas pressure tends to reduce the

bolt preload, which reduces gasket compression and tends to

separate the flange faces. Due to flange opening , bending has

been noticed in the bolt. Hence the bolts/studs should be

designed to withstand against preload, internal pressure load

and bending moment. Due to existence of Preload, internal

pressure and bending moment at a time, the bolt behavior is

nonlinear which cannot not be evaluated by simplemathematical formulas. 3-Dimensional finite element analysis

approach is only the technique which shows some satisfactory

result.

Nomenclature

Fi = Initial PretensionTin = Torque Applied on each stud

D = Diameter of the bolt/stud

K= Nut factorFe= Total external load on all studs/ bolt

Fp= external load due to pressure on each bolt/stud

P= Proof Pressure in bar

n = number of studs/ boltsFt= Total load on each stud/boltFb = Prying force= Joint factor based on stiffness, dimensionless quantity

I. INTRODUCTION

Generally one dimensional analytical formulae are used to

design bolts and flanges of the bolted flange joint in which

the contact surfaces, which include the area between the

flanges and the gasket are assumed to be constant and

without friction between the flanges. The non-linear force-

deflection behavior of the bolted flange joint was presented

in Brickford [1]. In the paper ofZahavi[10] the numerical

analysis based on finite element method is presented and

non-linear stiffness characteristics of the bolt joint is

considered. In this paper, 3-dimensional finite element

analysis of the bolted flange joint of the pressure vessel has

been presented with the assumption of friction exist

between the flanges. BOOSTER motor is a major

propulsion system for surface to surface missile. It has

bolted flange joint between motor and nozzle. The bolted

flange joint consist 44 numbers of M14 x 1.5 studs of class

10.9 at the PCD of 590 mm. The detail of joint is shown in

figure-1and figure-2.. The motor is having M14 x 1.5

tapped holes with 21.0 mm threaded length and 24.0 mm

drilling depth. Convergent is having free holes of

14.5mm. The maximum expected operating pressure (MEOP) of

the motor is 70 Ksc and proof pressure (PP) is 77 Ksc (1.1

times of MEOP). The motor casing material is Maragin

steel and a fastener is of 15-5-PH steel. Booster motor act as

a pressure vessel and named as PC vessel. The main

objective of this work is to find out the strength of the

bolted flange joint so that bolt should not be failed during

proof pressure test.

II. STRESSES IN THE BOLT

Tightening the bolt on a flange sets up stress and

strain in both the bolt and flange members. The bolt is

placed in tension while the joint members are in

compression, at least in the vicinity of the bolt. To load the

joint by applying the torque in criss-cross pattern, so stress

is equally distributed in the flange. Short-form torque

preload equation is used to evaluate the initial preload

created in a bolt. Nut factor value ranges from 0.1 to 0.2. In

this case, 0.2 is assumed for worst condition.


2/6



Fig. 1. Joint Detail.

Fig. 2. 3-dimensional joint detail.

KxD

TF

ini =

Apart from this, internal pressure (proof pressure)

of 77 bar is acting. This pressure load enhance the total load

on the bolt. This is called a prying load, such a load can

drastically increase the amount of tensile and bending stress

produced in the bolt for a given external force as shown in

figure-3. To estimate the magnitude of bolt forces produced

due to prying, finite-element analysis has been carried out.

Prying always bends the bolt, increasing stress on one side

more than the other as shown in figure-3. The prying force

is given as

)+1(=a

bFF pb

Hence total tensile load on each bolt is given by Ft.

Ft=Fi + x Fb

Fig. 3. Free body diagram for prying force.

The tension in the bolt which has been caused by the

external load and magnified by the prying action, and the

bending stresses created in the bolt as the joint members are

pried apart.

III. FINITE ELEMENT MODEL

Modeling of bolt in 3- dimensional finite element

application is still complicated. The analysis has been

carried out using ANSYS 12.1. As per cyclic symmetry,

One sector of 8.18 (360/44) has been analysed. SOLID95

(the element is having 20 nodes with three degrees of

freedom per node i.e. translations in the nodal x, y, and z

directions. This element has plasticity, creep, stress

stiffening, large deflection, and large strain capabilities.) has

been used for meshing. The flange interface and Nut-Flange

interface has been modeled with contact elements

CONTA174 & TARGE170 elements. The bolt is modeled

with solid elements. Cyclic symmetry boundary conditions

are applied at the edges of the model. Input parameters are

shown in table-1.

A cyclic symmetry analysis is required torepresent one part of a pattern that, if repeated N times in

cylindrical coordinate space, yields the complete model.

The angle (in degrees) spanned by the sector should be

such that n = 360, where n is an integer. In this case n is

44. The sector is constrained by symmetric boundary

condition on the corresponding surfaces. The cyclic sector

of 8.180 is shown if fig. 3. Here half of the sector is


3/6



considered for the analysis as shown in fig. 4 and Finite

element model in fig. 5.

The Preloads (initial tension) in bolts have

significant effect on deflections and stresses. The pretension

load is used to model a pre-assembly load in a joint fastener.The pretension section has been created in the shank portion

and pre-tightening force is simulated with PREST179Elements. Pretension section, across which the pretension

load is applied, must be defined inside the fastener. The

pretension load direction is along the bolt axis i.e. the bodythat contains Bolt Pretension has been meshed to partition

along the axial direction. The meshed pretension section is

flat with coincident nodes on the two sides (A and B) of thepretension section as shown in figure-6. The side A and B

on the pretension section are connected by one pretension

elements for each coincident node pair. The type ofelements and material properties use are mentioned in

Table-I and Table-II respectively.

TABLE 1

ELEMENT HISTORY

Sl.

No.

Component's name Element's Type

1. Flange and shell SOLID95

2. Bolt and Nut SOLID95

3. Contact Element CONTA174 and

TARGE170

4. Pretension PREST179

TABLE II

MATERIAL PROPERTIES

Sl.

No

Component's

name

Parameters Units Value

Modulus of

elasticity

Kg/mm2 190001. Flange

Poisson's Ratio ----- 0.3

Modulus of

elasticity

Kg/mm2 210002. Bolt and Nut

Poisson's Ratio ----- 0.3

TABLE III

INPUT PARAMETERS

Sl.

No.

Parameters Units Value

1. Bolt pre-tightening torque Kg-m 14

2. Bolt pre-tightening force with torque

coefficient of 0.2

Kg 5000

3. Bolt stress area mm2 125

4. Thread shear area at pitch line (Length

of engagement = 11.6 mm)

mm2 230

5. Bolt pre-tightening stress Kg/mm2 40.0

6. Bolt material class 10.9

7. Yield Strength of Bolt material Kg/mm2 90

8. Ultimate tensile strength of motormaterial (Maragin steel-250)

Kg/mm2 175

9. YS of motor material Kg/mm2 160

10 Coefficient of friction between the

flanges

---- 0.05

Fig. 4. Half sectorial detail of the flange joint.

Fig. 5. Finite element model of half sectorial bolted flange joint.


4/6



Fig. 6. Pretension sector in the shank portion of the stud.

IV. RESULTS

The 3-dimensional cyclic analysis has given the value of

axial forces and stress induced due to these axial forces as

shown in table-4 and table-5. It also has given the value of

bending moment on each bolt/stud. This bending moment

divided by the sectional modulus of the stud to obtain

bending stresses in the bolt/stud. The contour plot of

stresses is sector shown in figure 7. The total stresses in

each bolt/ stud is the sum of axial stress and bending stress.

The deflection of stud due to internal pressure is shown in

figure-8. The total stresses and bending stresses in the

bolt/stud is shown in figure-9 and 10 respectively. The axial

stress in the bolt/stud due to pretension is shown in figure-

11. The contact gap and stresses in flange are shown in

table-6 and in graph no 1,2,3 and 4 respectively. The cyclic

expansion of the sector is shown in fig. 12.

TABLE IV

TABULATED LOAD RESULT

Sl. No. Parameters Units values

1. Pretension load Kg 5000.0

2. Axial load due to externalload with prying action

Kg 7218.7

3. Bending moment Kg-mm 19008.9

TABLE V

RESULTED STRESS VALUES

Sl.No

Press-ure

Load

(bar)

Axialstress due

topretension

(Kg/mm2)

Axialstress due

to internalpressure

(Kg/mm2)

Bending

stress(Kg/

mm2)

Totalstress

(Kg/mm2)

Shearstress

on nut(Kg/m

m2)

1. 77 44.90 57.75 70.56 173.21 30.0

TABLE VIRESULT OF CONTACT ELEMENTS

Sl.

No.

Parameters Units values

1. Contact gap between flanges just

below o-ring (in intermediate zone

between two bolt)

mm 0.36

2. Contact gap between flanges below

o-ring (just below the bolt)

mm 0.41

3. Maximum Principal stresses in

flange (in intermediate zone between

two bolt)

Kg/mm2 31.8

4. Maximum Von-misses stresses in

flange (in intermediate zone betweentwo bolt)

Kg/mm2 45.76

5. Maximum Principal stresses in

flange (just below the bolt)

Kg/mm2 52.0

6. Maximum Von-misses stresses in

flange (just below the bolt)

Kg/mm2 48.0

Fig. 7. Distribution of Longitudinal stress in sector.


5/6



Fig. 8. Distribution of deflection due to bending of the stud.

Fig. 9. Distribution of total stresses in the stud.

Fig. 10. Distribution of bending stresses in the stud

Fig. 11. Distribution of axial stresses due to preloading.

Graph 1: Distribution of contact gap (between two bolts).

Graph 2: Distribution of contact gap (just below the bolt)).


6/6



Graph 3: Distribution of contact gap (between two bolts).

Graph 4: Distribution of contact gap (just below the bolt).

Fig. 12. Contour plot of stresses in cyclic expansion.

V. CONCLUSION

20 numbers of PC vessel have been proof tested

successfully. The experimental result of proof pressure

shows that the stresses predicted by 3-dimensional FEM

analysis are more than realistic stresses. The PC vessel has

been busted at 141.0 bars. The joint withstand upto 141.0

bars without any leakage/signal of failure. Hence the

stresses shown by the FEM analysis are for the reference

only. The actual stresses are less than 3D Finite Element

analysis result. The Finite Element analysis is very

conservative and shows no factor of safety on fasteners on

Proof Pressure. But experiment results have been shown

sufficient factor of safety is available on the bolt. Alsostresses in the stud/bolt depend upon the friction exist

between the flanges. As the coefficient of friction increases,

stresses in the bolt/stud decreases. To predict actual stresses

in the bolt/stud, measured the actual coefficient of friction

between flanges and same will be used in the analysis.

REFERENCES

[1] John H. bricford, Introduction to the design and behavior of bolted

joint," fouth edition", ; CRC press, taylor & francis group", pp.

259298.[2] BaroLomieg Zylinsky and Ryszard Buczkowski, Analysis of Bolt

Joint using the Finite Element Method," The Archieve ofMechanical Engineering",Volume LVII, 2010.

[3] NanBu, Naohiro, Ueno and Osamu Fukuda, Finite element Analysisof Contact stress in a full metallic pipe for hydrogen pipeline, "AISI

, Japan", pp. 184189.[4] Kathryn J. Belisle, Experimental and Finite Element Analysis of a

Simplified Aircraft wheel bolted Joint model," The Ohino state

University,2009.

[5] Yasumasa Shoji and Santoshi Nagata, analysis of Gasket Flangeswith Ordinary Elements using APDL Control," Toyo Engineering

Corporation, chiba, Japan,2002.

[6] JCharles S. Hseih, Steven R. Massey and Dennis H. Martens, Design

of flanged joint subjected to pressure and external loads, PVP,

ASME, New York, 1999.[7] ASME code , Pressure vessel Design ,"chapter nine", pp. 110112.

[8] Robert D. Cook Finite element modeling For Stress Analysis", "John

Wiley & Sons,Inc,2001" , pp. 41168.[9] Zahavi E., A finite Element Analysis of flange connections,Journal of Pressure Vessel Technology, ASME, 115-1993, pp.

327-330.[10] Joseph E. Shigley, Machine Design," second edition", ; Mcgraw-hill

", pp. 250280.[11] Jerome Montgomery, Method for modelling bolt in bolted joint,

Siemens Westinghouse Power Corporation, Orlando, FL".

bolted flange connection

Documents