saer-6061

Engineering Report SAER-6061 March, 2006

“Composite Repairs for Piping Systems, Tanks and Pressure Vessels”

(Amerada Hess, BP, Saudi Aramco, Shell)

Issue date: March, 2006 SAER-6061 “Composite Repairs for Piping Systems, Tanks and Pressure Vessels”

ii

Table of Contents

INTRODUCTION 1

SCOPE 2

NOMENCLATURE 4

DESIGN 6

Straight pipe repair design 7 Design based on pipe allowables 8 Design based on composite allowable strains 8 Design based on leaking pipes 8

Piping system components repair design 9 Bends 9 Tees 10 Reducers 11 Flanges 12

Cylindrical tank and pressure vessel repair design 14 End dome, main body connection 15 Supports/saddles/rigid attachments 15 Tees/nozzles 17 Supports/saddles/rigid attachments 19 Tees/nozzles 21

Overlap lengths – repair thickness increase factors 22 Stress decay within the repair laminate 22 Stress decay within the adhesive bond line 23 Stress decay from a circular hole within a flat plate 24

Pressure area method for composite repairs to tees and nozzles 27 Tee 29 Vessel nozzle 30

Summary of design method and design rules 31

FEA VERIFICATION OF DESIGN RULES 35

Cylindrical vessel 35

Spherical vessel 38

Overlay length 42

Summary of conclusions of FEA analysis 42


iii

EXPERIMENTAL VERIFICATION OF DESIGN RULES 43

Test arrangement 44

Defects 45

Repair design 46

Pressure testing 46 Repair Supplier 1 – test results and analysis 47

Short term test results and analysis 47 Medium term test results and analysis 48

Repair Supplier 2 – test results and analysis 52 Short term test results and analysis 52 Medium term test results and analysis 54

Repair Supplier 3 – test results and analysis 56 Short term test results and analysis 56 Medium term test results and analysis 57

Generic conclusions 61

GENERAL CONCLUSIONS AND RECOMMENDATIONS 63

BIBLIOGRAPHY 64

ATTACHMENTS: 1) AEAT - 57711, “Design of Composite Repairs for Pipework” 2) AEAT - 57756, “Installation procedures for composite repairs” 3) AEAT - 75394, “NDT Methods for composite repairs” 4) AEAT - 02529, “Documentation for the use of composite repairs” 5) AEAT - 57394, “Composite overwrap repairs - medium term testing and analysis”


1

Introduction Design codes and standards for pressurized equipment provide rules for the design, fabrication, inspection and testing of new piping systems. These codes do not address the fact that equipment degrades in service or may require to be up-rated due to a change in duty, nor do they consider options for remedial action should such events occur. This specification provides guidance for the design of one repair option: the external reinforcement and the repair of damage such as holes in pipe or pipe components using composite materials. The procedures described in this specification can be used to design composite reinforcements to allow damaged pressurized equipment to continue to operate safely. The development of this document was carried out in collaboration with a number of organizations representing material suppliers, users and regulatory agencies. Those involved included Shell, BP, Saudi Aramco, Amerada Hess, Petrobras, Petronas, Statoil, BG-Hydrocarbon Resources Ltd, Devonport Marine Dockyard Ltd., Walker Technical Resources, Clock Spring and Industrial Maintenance Group. This document is one of a number that covers the design (AEAT – 57711), installation (AEAT- 57756) and inspection (AEAT- 75394) of composite repairs. An overview of the documents is provided in a summary report (AEAT- 02529). This document covers the extension of the scope of the design document from pipework to more complicated geometries including, bends, tees, reducers, nozzle attachments for tanks and pressure vessels. The major design or technical challenges for the repair to part of a piping system or a tank or vessel are;

• The calculation of the local stress field around the defect • Limited length of repair due to geometrical constraints • Large diameter of vessels implying over-wrapping the whole diameter may not be practical


2

SCOPE Procedures in this document cover the repair of carbon steel pipework and related components, pipelines originally designed in accordance with a variety of pipe standards including ISO 15649/13623, ASME B31.1/B31.3/B31.4/B31.8 and BS 8010 and tank, vessel standards ASME Pressure vessel and boiler code and BS 4994/5719. The following circumstances or defects are addressed:

• external corrosion, which may or may not cause leaking, and structural integrity needs to be restored. In this case it is probable that with suitable surface preparation the application of a composite overwrap will arrest further deterioration;

• external damage such as dents, gouges, fretting (at supports) where structural integrity needs to be restored;

• internal corrosion, which may or may not be leaking, and there is a need to restore structural integrity. In this case it is probable that corrosion will continue and the assessment must take this into account;

• structural strengthening to account for an increase in pressure rating or other loads in local areas.

Operational services that are considered are:

• utility fluids, diesel, seawater, air; • chemicals; • produced fluids, including gas and gas condensate.

The applicable pressure/temperature envelope is dependent on the type of damage being repaired. For all repairs continuous service temperatures should be limited to the range -50oC to 100oC. Where the pipe being repaired is leaking the upper limit for continuous service pressure should be 50 bar g (Piping Class 300). This limit should also apply for repairs where it is assessed that any continuing degradation of the substrate, e.g. through internal corrosion, will result in a through wall defect at some point during the remaining design life. Where leaking of the pipe being repaired is not a design factor, there is, in principle, no upper limit to continuous service pressure, although all repairs will be subject to a risk assessment where pressure will be an important consideration. The composite materials considered within the document are those with glass (GRP) or carbon (CFRP) reinforcement in a polyester, vinyl ester or epoxy matrix. Use of this document outside these service ranges is possible subject to the comments given below. Examples where the details of the application are outside the above scope, but where the intent of the design guidelines may be used coupled with a more complete analysis are:

• other pipe materials, e.g. alloyed steel; • other degradation mechanisms, e.g. wall loss due to erosion;


3

• other service conditions, e.g. process fluids or higher operating envelope in terms of pressure and temperature;

• other composite material systems. Due to the complicated geometry of the repair conditions, all repair considered will be deemed to be Class 3. The contents of this report are concerned with the development of the design rules for piping system components, tanks and vessels and their numerical and experimental verifications. In particular the following sections include;

• Section 3 – Nomenclature • Section 4 – Development of design rules • Section 5 – FE analysis of reference geometries • Section 6 – Experimental tests of repairs, both short and long term.


4

NOMENCLATURE 2Ca = axial length of attachment (m) 2Ch = hoop length of attachment (m) d = diameter (or diameter of the equivalent circle) of the leaking region (m). D = external pipe diameter (m) Da = diameter of attachment (m) Di = external diameter of component (m) Db = diameter (branch) (m) Do = diameter (outlet or dome minor axis) (m) Eh = tensile modulus for the composite laminate in the hoop direction (N/m2) Ea = tensile modulus for the composite laminate in the axial direction (N/m2) Es = tensile modulus for steel (N/m2) fleak = service factor for leaking repairs fth,overlay = repair thickness increase factor for available overlay length fth,stress = repair thickness increase factor for stress intensity factor corresponding to the

component F = sum axial tensile loads due to pressure, bending and axial thrust (N). (Note that the axial tensile load generated by an applied bending moment is (4M/D)) G = shear modulus for the composite laminate L = length of cylindrical vessel (m) Le = off-centre axial length of attachment (m) Lavailable = available axial extent of undamaged pipe section (m) mpsm = pressure stress multiplier M = bending moment (Nm) P = internal design pressure (N/m2) Ps = MAWP (maximum allowable working pressure) for the steel pipe (N/m2) Q = shear force or thrust (N) Rb = bend radius of component (m) s = allowable tensile stress for the steel (N/m2) tb = wall thickness (branch) (m) tf = wall thickness (flange) (m) ti = wall thickness (component, main section) (m) trepair,X = design thickness for a repair laminate for component X (m) ts = minimum remaining substrate wall thickness of the pipe (m) w = displacement (m) εc = allowable circumferential strain εa = allowable axial strain γ = toughness parameter (energy release rate) for the composite steel interface σ = stress (N/m2) ν = Poisson's ratio for the composite laminate in the circumferential direction


5

Subscripts a = axial

b = bending h = hoop

i = in-plane

o = out-of-plane

p = pressure sh = shear


6

DESIGN The design document (AEAT – 57711) provides details on the input data, necessary to specify the repair and required to perform the design calculation. The outputs of the design calculation are;

• Repair thickness • Repair overlay length (axial length of repair)

The scope of the design document (AEAT-57711) covers the repair situation for pipework. This document considers the extension of this scope to include;

• Piping systems components o Bends o Tees o Reducers o Flanges

• Cylindrical vessels

o End dome, main body connection o Supports/saddles/rigid attachments

Thrust loading Axial and hoop moment loading

o Tees/nozzles Pressure loading Axial and hoop moment loading Thrust loading

• Spherical vessels

o Supports/saddles/rigid attachments Thrust loading Moment loading

o Tees/nozzles Pressure loading Moment loading Thrust loading

The design approach for repairs to piping systems, tanks and vessels is comparable to that for pipework. For each component a comparative approach is adopted based on the equivalent straight pipe component. The design process is to calculate the repair thickness for an equivalent straight pipe section then calculate further additional multiplicative factors (called repair thickness increase factors) allowing for both the stress intensification due to the geometry of the component and the possible reduction in overlap area available for repair. The first step in the design approach, therefore is to calculate the thickness of the repair for the equivalent pipe section (see Section 0), i.e. same diameter and wall thickness, trepair,straightpipe.


7

The next step is to calculate repair thickness increase factors for;

• Available overlay length (less than the design overlap length), fth,overlay • Stress intensity factor corresponding to the piping system component, fth,stress

The repair thickness for the piping system component is given by the product of these repair thickness increase factors times the repair thickness for the equivalent straight pipe section, i.e.;

overlay,thstress,theraight piprepair, stmponentrepair, co fftt = (1)

For most components the repair thickness is calculated to withstand both hoop and axial applied loads. The design repair thickness is taken as the larger of these two values. For the leaking repair situation, for all geometries it is assumed that the flat plate solution as derived in AEAT – 57711 is valid, i.e.,

⎪⎪

⎭

⎪⎪

⎬

⎫

⎪⎪

⎩

⎪⎪

⎨

⎧

+⎪⎭

⎪⎬⎫

⎪⎩

⎪⎨⎧

+−

=24

3

2

6431

51231 d

Gtdd

tE)(

fP

pestraightpi,repairpestraightpi,repair

leak

πυ

γ (2)

where the repair thickness is then multiplied by fth,overlay and fth,stress to obtain the final repair thickness, i.e.


The input data for the design calculation is similar to that for pipework and therefore the design input data sheet quoted in AEAT – 57711 is appropriate.

Straight pipe repair design The basic design equations for straight pipe are derived in AEAT-57711. To calculate the repair thickness, one of two design options is chosen, depending on whether or not the pipe substrate is assumed to contribute to carrying the applied load. Either option ensures that the strength of the repair is sufficient to withstand the applied loads. For leaking repairs, a third option, must also be used, to check that the adhesion of the repair laminate to the substrate is sufficient. The equations for each design option are given in the following Sections.


8

Design based on pipe allowables This design option is chosen for the situation when the load carrying contribution of the substrate pipe is considered. The thickness for the repair (laminate), trepair, is given by the larger of the hoop and axial load carrying requirements, respectively:

( ) ⎥⎦

⎤⎢⎣

⎡⎟⎠⎞

⎜⎝⎛ −⋅⎟⎟

⎠

⎞⎜⎜⎝

⎛⋅−⋅⎟⎟

⎠

⎞⎜⎜⎝

⎛⋅= s2s

22,2 P

DF

EE

sDPP

EE

sDMaxt

a

s

c

srepair π

(4)

Design based on composite allowable strains This design option is chosen for the situation when the load carrying contribution of the substrate pipe is ignored. The thickness for the repair (laminate), trepair, is given by the larger of the hoop and axial load carrying requirements, respectively:

⎥⎦

⎤⎢⎣

⎡⎟⎟⎠

⎞⎜⎜⎝

⎛−⎟⎟

⎠

⎞⎜⎜⎝

⎛−=

caacccrepair E

PDED

F,ED

FE

PDMaxt υπε

υπε 2

1112

1 (5)

Design based on leaking pipes For leaking pipes, in addition to either design option 0 or 0, an analysis of the interfacial delamination resistance (or interfacial fracture toughness) of the repair system is required. The repair thickness, trepair, is related to the design pressure, P, by;

⎪⎪

⎭

⎪⎪

⎬

⎫

⎪⎪

⎩

⎪⎪

⎨

⎧

+⎪⎭

⎪⎬⎫

⎪⎩

⎪⎨⎧

+−

=24

3

2

6431

5123)1( d

Gtdd

tE

fP

repairrepair πυ

γ (6)

The design of repair thickness for a leaking defect or hole is generally governed by the delamination failure of the repair, rather than the strength requirement of carrying the applied load.


9

Piping system components repair design The following piping system components are considered;

• Bends • Tees • Reducers • Flanges

The following sections describe the derivation for fth,stress for each component.

Bends The repair thickness increase factor for bends, fth,stress is derived in this section and is based on ISO 14692 – Part 3. The average hoop, σav,h and axial stress, σav,a, within the component is given by;

22

22

41

4

⎟⎟⎠

⎞⎜⎜⎝

⎛+⎟

⎟⎠

⎞⎜⎜⎝

⎛+=

⎟⎟⎠

⎞⎜⎜⎝

⎛+⎟

⎟⎠

⎞⎜⎜⎝

⎛+=

p,a

sh

p,a

b,ap,aa,av

p,h

sh

p,h

b,hpsmp,hh,av m

σσ

σσ

σσ

σσ

σσ

σσ

(7)

where 223

81oi

i

h

p,h

b,h MMDS

P+=

πσσ

, 4

3813

,

sh

iph

sh MDP πσ

σ=

223

161oi

i

a

p,a

b,a MMDS

P+=

πσσ

, 4

31613

,

sh

ipa

sh MDP πσ

σ=

The definitions of stress intensity factors are taken from ISO 14692;

32

2611324

2111

161/

i

bi/

i

i/

i

b

h

h DRt

tD

tR

EP.

.S ⎟⎟⎠

⎞⎜⎜⎝

⎛

⎟⎟⎠

⎞⎜⎜⎝

⎛⎟⎟⎠

⎞⎜⎜⎝

⎛+

= ,

32

22314

25321

1760/

i

bi

i

i/

i

b

h

a DRt

tD

tR

EP.

.S ⎟⎟⎠

⎞⎜⎜⎝

⎛

⎟⎟⎠

⎞⎜⎜⎝

⎛⎟⎟⎠

⎞⎜⎜⎝

⎛+

=


10

As with most piping system components, the only applied load that will be known with any degree of certainty is the internal pressure, all other applied loads will be difficult to quantify. Therefore, the applied moments are assumed to be related solely and conservatively to the internal pressure by;

016

3

=== shi

oi MandPDMM π

For bends the pressure stress multiplier, mpsm, is given by ISO 14692, mpsm = 1 Therefore, repair thickness increase factors in the hoop and axial directions are given by;

ap,a

a,ava,th

hp,h

h,avh,th

S.f

S.f

41411

70701

+==

+==

σσ

σσ

(8)

The stress intensity factors, Sh and Sa, both have a maximum value of 2.5, implying that fth,h = 2.76 and fth,a = 4.5. These maxima are defined in ISO 14692. For pressure loading only acting on the bend, Equation (8) simplifies to;

11

=

=

a,th

h,th

ff

(9)

Tees The repair thickness increase factor for tees, fth,stress is derived in this section and is based on ISO 14692 – Part 3.

The following definition is used throughout this section ⎟⎟⎠

⎞⎜⎜⎝

⎛⎟⎟⎠

⎞⎜⎜⎝

⎛=

i

i

b

bt t

DDt

2

2λ

The average hoop, σav,h and axial stress, σav,a, within the component is given by;

22

22

41

4

⎟⎟⎠

⎞⎜⎜⎝

⎛+⎟

⎟⎠

⎞⎜⎜⎝

⎛+=

⎟⎟⎠

⎞⎜⎜⎝

⎛+⎟

⎟⎠

⎞⎜⎜⎝

⎛+=

p,a

sh

p,a

b,ap,aal,av

p,h

sh

p,h

b,hpsmp,hh,av m

σσ

σσ

σσ

σσ

σσ

σσ

(10)

where; 223

,

, 81oih

iph

bh MMSDP

+=πσ

σ ,

4381

3,

sh

iph

sh MDP πσ

σ=

223

,

, 161oia

ipa

ba MMSDP

+=πσ

σ ,

43161

3,

sh

ipa

sh MDP πσ

σ=

The definitions of stress intensity factors and pressure stress multiplier are taken from ISO 14692;


11

5.066.0

thS

λ= and 5.0

66.0

taS

λ= with 25.0

4.1

tpsmm

λ=

As with most piping system components, the only applied load that will be known with any degree of certainty is the internal pressure, all other applied loads will be difficult to quantify. Therefore, the applied moments are assumed to be related solely and conservatively to the internal pressure by;

016

3

=== shi

oi MandPDMM π

This relationship comes the assumption that the axial load i.e. axial stress induced by moments is the same as that induced by internal pressure. Therefore, repair thickness increase factors in the hoop and axial directions are given by;

50

50250

9301

47041

.tp,a

a,ava,th

.t

.tp,h

h,avh,th

.f

..f

λσσ

λλσσ

+==

+==

(11)

Reference standard IGE TD/12 lists stress concentration factors for various tees. Equation (11) is comparable to the bending and pressure stress concentration factors under fatigue loading. Therefore the above listed approach will be conservative for continuous loads. For pressure loading only acting on the tee, Equation (11) simplifies to;

1

41250

=

=

a,th

.t

h,th

f

.fλ (12)

The pressure stress multiplier has a maximum value of 3, implying that fth,h = 3. This maximum is defined in ISO 14692.

Reducers The analysis of the repair thickness increase factor for reducers is based on Timoshenko [1]. The stress intensification in reducers is caused by the difference in diameter along the axial length of the component under internal pressure. This difference in diameter results in a differential radial displacement with the maximum given by;

⎟⎟⎠

⎞⎜⎜⎝

⎛−=Δ 2

22

14 i

o

i

i

DD

EtPDw

This difference in displacement results in an induced shear force within the wall of the component and is given by;


12

)(

EtwQ i2

33

112 υβ

−Δ= implying ⎟⎟

⎠

⎞⎜⎜⎝

⎛−

−= 2

2

21

1124

11

i

o

ii DD

)(tDPQ

υβ

The average hoop, σav,h and axial stress, σav,a, within the component is given by [1];

2

2

64

262

i

a

i

ial,av

ii

a

i

ih,av

tM

tPD

DwE

tM

tPD

+=

Δ++=

σ

υσ

Inserting the applied moment, ββζ /)x(QM a −= , and displacement, Δw gives;

( )

)x(Qt

PD

)x()()x(Qt

PD

i

ia,av

i

ih,av

βζβ

σ

βθυβυζβ

σ

64

11262

2

+=

−++=

(13)

where ζ(βx) and θ(βx) are defined in [5]. Inserting the shear force, Q, in the above and locating the maximum values of the average hoop stress (which occurs at βx = 1.8574) and average axial stress (which occurs at βx = π/4), it can be shown that the hoop and axial repair thickness increase factors simplify to;

⎟⎟⎠

⎞⎜⎜⎝

⎛−+=+==

⎟⎟⎠

⎞⎜⎜⎝

⎛−+=+==

2

2

2

2

158601173871

10640118301

i

o

iip,a

a,ava,th

i

o

iip,h

h,avh,th

DD.

tDPQ.f

DD.

tDPQ.f

βσσ

βσσ

(14)

Flanges The analysis of the repair thickness increase factor for flanges is based on Timoshenko [1]. The stress intensification in flanges is caused by the difference in thickness along the axial length of the component under internal pressure. This difference in thickness results in a difference in radial displacement with the maximum difference given by;

⎟⎟⎠

⎞⎜⎜⎝

⎛−=Δ

f

i

i

i

tt

EtPDw 14

2


)(EtwQ i

2

33

112 υβ

−Δ=

implying ⎟⎟⎠

⎞⎜⎜⎝

⎛−

−=

f

t

ii tt

)(tDPQ 1

1124

112υβ


13

Inserting this shear force, Q, in Equation (13) and locating the maximum values of the average hoop stress (which occurs at βx = 1.8574) and average axial stress (which occurs at βx = π/4), it can be shown that the hoop and axial repair thickness increase factors simplify to;

⎟⎟⎠

⎞⎜⎜⎝

⎛−+=+==

⎟⎟⎠

⎞⎜⎜⎝

⎛−+=+==

f

i

iip,a

a,ava,th

f

i

iip,h

h,avh,th

tt.

tDPQ.f

tt.

tDPQ.f

158601173871

10640118301

βσσ

βσσ

(15)


14

Cylindrical tank and pressure vessel repair design The following connections and attachments with corresponding loads are considered for cylindrical tanks and pressure vessels;

o End dome, main body connection o Supports/saddles/rigid attachments

Thrust loading Axial and hoop moment loading

o Tees/nozzles Pressure loading Axial and hoop moment loading Thrust loading

The following sections describe the derivation of the hoop and axial repair thickness increase factors, fth,stress, for each component and loading situation. The following formulae are used throughout this section;

⎟⎟⎠

⎞⎜⎜⎝

⎛⎟⎟⎠

⎞⎜⎜⎝

⎛=

i

i

b

bt t

DDt

2

2λ

i

brr

ii

b

ttt

tDD

== ,2

ρ

( )

( ) ( )( ) 1.06.02

215.01

)()(65.02.0144.1

25.02.045.0092.0

alrtha

alhartal

CLogCLogCosCK

CCCCosCK

−−=

−−−−= −

( ) ( ) 3050

4

3

240181

2220

.al

rt.ha

alrt

)C(Log)C(LogCosC..K

C)C(LogCos.K

−⎟⎠⎞

⎜⎝⎛−=

⎟⎠⎞

⎜⎝⎛ −

=

LCC

DC

tDC

CCC a

ali

a

i

irt

a

hha

2,128,2

=⎟⎟⎠

⎞⎜⎜⎝

⎛==


15

End dome, main body connection The analysis of the repair thickness increase factor for main body, end dome connection is based on Timoshenko [1]. The stress intensification at the connection is caused by the difference in radial displacement under internal pressure between the main body and the end dome and is given by;

2

22

4 o

i

i

i

DD

EtPD

w =Δ


)(

EtQ i2

33

112 υδβ

−Δ= implying 2

2

21128

11

o

i

ii DD

)(tDPQ

υβ −=

Inserting this shear force, Q, in Equation (13) and locating the maximum values of the average hoop stress (which occurs at βx = 1.8574) and average axial stress (which occurs at βx = π/4), it can be shown that the hoop and axial repair thickness increase factors simplify to;

2

2

2

2

29301173871

0320118301

o

i

iip,a

a,ava,th

o

i

iip,h

h,avh,th

DD.

tDPQ.f

DD.

tDPQ.f

+=+==

+=+==

βσσ

βσσ

(16)

Supports/saddles/rigid attachments The analysis of the repair thickness increase factor for supports, saddles and rigid attachments is based on BS 5500 [17]. An empirical approach is followed based on graphical data presented in BS 5500 [17] to account for the stress intensity caused by rigid attachments to the cylinder tank body.

If cylinder attachment is off-centre, then L is replaced by Le where LxLLe

24−= and x is the distance

between the axial centre of the cylinder and the centre of the attachment. For a rigid attachment of a circular cross-sectional area then aah DCC 85.022 == . The average hoop, σav,h and axial stress, σav,a, within the component is given by [1];


16

2

2

64

62

i

a

i

a

i

ia,av

i

h

i

h

i

ih,av

tM

tN

tPD

tM

tN

tPD

++=

++=

σ

σ

Inserting the empirical relationships for the hoop tensile and moment loads resulting from the applied thrust, Q, from [17] implies that the hoop and axial repair thickness increase factors for an applied thrust load are given by;

( )

( )21

43

41

21

KKtDP

Qf

KKtDP

Qf

iia

a,ava,th

iih

h,avh,th

++==

++==

σσσ

σ

(17)

Combined thrust, axial and hoop moment loads applied to cylindrical vessel attachments are designed by assuming that the moment loads are equivalent to applied thrusts acting on the edge of the attachment.

h

h

a

aequivalent C

MCMQQ

25.1

25.1

++=

Using this result from the applied thrust load case implies the hoop and axial repair thickness increase factors for combined applied loads are;

( )

( )21

43

4251

25111

2251

25111

KKtDC

M.CM.Q

Pf

KKtDC

M.CM.Q

Pf

iih

h

a

aa,th

iih

h

a

ah,th

+⎟⎟⎠

⎞⎜⎜⎝

⎛+++=

+⎟⎟⎠

⎞⎜⎜⎝

⎛+++=

(18)

If the thrust or moment load is present but unknown then an approximation can be made by assuming the stress induced by the applied thrust or moment load is equivalent to that induced by internal pressure, i.e.;

4

2aD

PQ π

= where haa CC.D 352= .

Using this approximation implies that the repair thickness increase factor for rigid attachments to cylindrical tanks or vessels is independent of the applied loading type. The hoop and axial repair thickness increase factors simplify to;


17

( )

( )21

2

43

2

1

21

KKtD

Df

KKtD

Df

ii

aa,th

ii

ah,th

++=

++=

π

π

(19)

Tees/nozzles The repair thickness increase factor for tees is derived in this section and is based on BS5500 [17]. For tees and nozzles conservative assumptions are made for the applied moments and thrusts acting as although they will be present they will be difficult to quantify. The assumption made is that the stresses induced by the applied thrusts and moments are equivalent to those induced by internal pressure. The repair thickness increase factors for pressure loads are taken from BS 5500 [17]. It is empirically based and is valid for both hoop and axial directions. The hoop and axial repair thickness increase factors for pressure loading only are given by;

450

215225 .

ra,thh,th )t.(

.ff ρ+

== (20)

For thrust and moment loads a similar approach as derived in 0 is used to determine the repair increase factors The average hoop, σav,h, and axial σav,a, stress within the component is given by [17];

2

2

6

6

i

a

i

aa,av

i

h

i

hh,av

tM

tN

tM

tN

+=

+=

σ

σ

Inserting the empirical relationships for the in-plane tensile and moment loads resulting from the applied thrust from [4] implies that the hoop and axial repair thickness increase factors are given by;

( )

( )21

43

4

2

KKtDP

Qf

KKtDP

Qf

iia

a,ava,th

iih

h,avh,th

+==

+==

σσσ

σ

(21)

Combined thrust, axial and hoop moment loads applied to cylindrical vessels attachments are designed by assuming that the moment loads are equivalent to applied thrusts acting on the edge of the attachment.

b

h

b

aequivalent D.

M.D.

M.QQ85051

85051

++=


18

Using the results from the applied thrust load case implies the hoop and axial repair thickness increase factors for combined thrust and moment loads are;

( )

( )21

43

485051

850511

285051

850511

KKtDD.

M.D.

M.QP

f

KKtDD.

M.D.

M.QP

f

iib

h

b

a

a

a,ava,th

iib

h

b

a

h

h,avh,th

+⎟⎟⎠

⎞⎜⎜⎝

⎛++==

+⎟⎟⎠

⎞⎜⎜⎝

⎛++==

σσ

σσ

(22)

For combined loadings the repair increase factors are the summation of the individual repair increase factors corresponding to the applied loads i.e.

momentththrustthpressurethtotalth ffff ,,,, ++= (23)

As with most nozzles attached to cylindrical vessels, the only applied load that will be known with any degree of certainty is the internal pressure. All other applied loads will be difficult to define. Therefore, the applied moments are assumed to be related (conservatively) to the internal pressure by;

4

2bD

PQ π

=

This relationship comes the assumption that the axial stress induced by moments is the same as that induced by internal pressure. The hoop and axial repair thickness increase factors are therefore given by;

( )

( )21

2450

43

2450

215225

2215225

KKtD

D)t.(

.f

KKtD

D)t.(

.f

ii

b.

ra,th

ii

b.

rh,th

+++

=

+++

=

πρ

πρ

(24)


19

Spherical tank and vessel repair design The following connections and attachments and corresponding loads are considered for spherical tanks and pressure vessels;

o Supports/saddles/rigid attachments Thrust loading Moment loading

o Tees/nozzles Pressure loading Moment loading Thrust loading

The following sections describe the derivation of the hoop and axial repair thickness increase factors, fth,stress, for each component and loading situation. The following formulae are used throughout this section;

ii

a

tDDs 287.1

=

i

brr

ii

b

ttt

tDD

== ,2

ρ

( )

( )5.02

35.01

2.286.4

4.22.1

sExpK

sExpK

−=

−=

( )

( )5.04

25.03

2.22.1

3.238.0

sExpK

sExpK

−=

−=

Supports/saddles/rigid attachments The analysis of repair thickness increase factors for supports, saddles and rigid attachments is based on BS 5500 [17]. An empirical approach is followed based on graphical data presented in BS 5500 [17] to account for the stress intensity caused by rigid attachments to the spherical tank body. The average hoop, σav,h and axial stress, σav,a, within the component is given by [1];


20

2

2

64

62

i

a

i

a

i

ia,av

i

h

i

h

i

ih,av

tM

tN

tPD

tM

tN

tPD

++=

++=

σ

σ

Inserting the empirical relationships for the in-plane tensile and moment loads resulting from the applied thrust, Q, from [17] implies that the hoop and axial repair thickness increase factors are given by;

( )

( )21

43

41

41

KKtDP

Qf

KKtDP

Qf

iia

a,ava,th

iih

h,avh,th

++==

++==

σσσ

σ

(25)

Combined thrust and moment loads applied to spherical vessel attachments are designed by assuming that the moment loads are equivalent to applied thrusts acting on the edge of the attachment.

ii

equivalent tDMQQ 4.1

+=

Using the results from the applied thrust load case implies the hoop and axial repair thickness increase factors for combined applied loads are;

( )

( )21

43

44111

44111

KKtDtD

M.QP

f

KKtDtD

M.QP

f

iiiia,th

iiiih,th

+⎟⎟⎠

⎞⎜⎜⎝

⎛++=

+⎟⎟⎠

⎞⎜⎜⎝

⎛++=

(26)

If the thrust or moment load is present but unknown then an approximation can be made by assuming the axial stress induced by the applied thrust or moment load is equivalent to that induced by internal pressure, i.e.;

4

2aD

PQ π

=

Using this approximation implies that the repair thickness increase factors for rigid attachments to spherical tanks or vessels are independent of the applied loading type. The hoop and axial repair thickness increase factors simplify to;

( )

( )21

2

43

2

1

1

KKtD

Df

KKtD

Df

ii

aa,th

ii

ah,th

++=

++=

π

π

(27)


21

Tees/nozzles The analysis of repair thickness increase factors for tees, nozzles is based on BS 5500 [17]. An empirical approach is followed based on graphical data presented in BS 5500 [17] to account for the stress intensity caused by nozzle attachments to the spherical tank body. In the following formula the repair thickness increase factor is independent of loading direction. The repair thickness increase factor for internal pressure loading is given by;

SCFfth = where ( )

31.0

88.095.075.01

2r

r

tt

SCF−+

=ρ

(28)

The repair thickness increase factor due to thrust loading is given by;

iib

th tDDPQSCFf

214

π= where ( ) ⎟⎟

⎠

⎞⎜⎜⎝

⎛−−= 42.0

3.0 2.0143.018.3ρrtSCF (29)

The repair thickness increase factor due to moment loading is given by;

iib

th tDDPMSCFf

2116

2π= where ( ) )sin(43.018.3 3.0 ρrtSCF −= (30)

For combined loadings the repair increase factor is the summation of the individual repair increase factors corresponding to the applied loads i.e.

momentththrustthpressurethtotalth ffff ,,,, ++= (31)

If the thrust or moment load is present but unknown then an approximation can be made by assuming the axial stress induced by the applied thrust or moment load is equivalent to that induced by internal pressure, i.e.;

4

2bD

PQ π

= and 16

3bD

PM π

=

Equating these formulas into the appropriate formulae above implies that the repair thickness increase factor is independent of loading type and is given by;

( )

ii

b

r

rth tD

Dt

tf

295.075.01

2 31.0

88.0−+

=ρ

(32)


22

Overlap lengths – repair thickness increase factors For pipework the design overlap (axial extent) length of the repair, Ldesign, is given by, AEAT - 57711;

iidesign tDL 2= (33)

For a cylindrical tank (main body only), the design overlap length (in any direction), Ldesign, is given by [17];

iidesign tDL 2= (34)

i.e. the same as for pipework. For a spherical tank or the end dome of a cylindrical tank, the design overlap length (in any direction), Ldesign, is given by [17];

iidesign tDL = (35)

i.e. one half that for pipework. For an available overlap length for the repair less than the design overlap length, Ldesign, the repair thickness increase factor, fth,overlap, is derived based on either the stress decay from the defect within the wall of the vessel or piping system component or within the adhesive bond line. Note, if sufficient overlap length is available then fth,overlap is unity. The following sections determine the theoretical stress decay for the following;

o Repair laminate o Adhesive bond o Circular hole within a flat plate

From this analysis the relevant stress intensity factors are derived. Also discussed within the last section is the limit of defect (circular hole) size for the energy release rate calculation for a straight pipe section.

Stress decay within the repair laminate For a cylindrical body under an axial moment, M0, the axial stress, σax, (as a function of axial distance from the applied moment, is given by (Timoshenko [1]));

)exp())cos()(sin(62 xxx

tM

repair

oax βββσ −+= where 22

24 )1(12

repairi tDυβ −

=

The average stress over an axial length, L, where βL >> 1 is given by;


23

Lt

M

repair

oaverageax β

σ 162, =

Therefore, the average stress is related to the overlap length and repair laminate thickness by;

2/3,1

repairaverageax Lt

∝σ

Assuming that for reduced overlay lengths, the average axial stress is not greater than if sufficient overlap length is available then; 2/3

,2/3

, availablerepairavailabledesignrepairdesign tLtL = Therefore, the thickness increase factor for limited available overlap length is given by;

3/2

,

,, ⎟⎟

⎠

⎞⎜⎜⎝

⎛==

available

design

designrepair

availablerepairoverlayth L

Lt

tf (36)

The stress decay from a defect within a spherical vessel is greater than that for a corresponding cylindrical vessel. Therefore, conservatively, the repair laminate thickness factor derived for a cylindrical vessel is taken also to apply to spherical vessels.

Stress decay within the adhesive bond line For a cylindrical body under an axial moment, M0, the axial stress, σax, (as a function of axial distance from the applied moment, is given by (Frost et. al. [5]);

)exp()cos(2 xxM oax βββσ −≈ where 34 1

repairt∝β

The average stress over an axial length, L, is given by;

)1)exp()sin()((cos(, −+∝ LLLL

M oaverageax ββββσ

Therefore, the average stress is related to the overlap length and repair laminate thickness by, assuming βL<1;

4/9,repair

averageax tL

∝σ

Assuming that for reduced overlay lengths, the average axial stress is not greater than if sufficient overlap length is available then;


24

4/9,

4/9, availablerepair

available

designrepair

design

tL

t

L=

Therefore, the thickness increase factor for limited available overlap length is given by;

9/4

,

,, ⎟⎟

⎠

⎞⎜⎜⎝

⎛==

available

design

designrepair

availablerepairoverlayth L

Lt

tf (37)

The decay of stresses analysis assumes that local to the defect the curvature from either the cylindrical or spherical vessel is negligible, i.e. the flat plate solution is valid.

Stress decay from a circular hole within a flat plate The stress decay from a circular hole (radius a) under a uni-axial load, S is given by (Timoshenko [6]);

)2sin(3212

)2cos(312

12

)2cos(3412

12

4

4

2

2

4

4

2

2

4

4

2

2

2

2

θτ

θσ

θσ

θ

⎟⎟⎠

⎞⎜⎜⎝

⎛−+−=

⎟⎟⎠

⎞⎜⎜⎝

⎛+−⎟⎟

⎠

⎞⎜⎜⎝

⎛+=

⎟⎟⎠

⎞⎜⎜⎝

⎛+−+⎟⎟

⎠

⎞⎜⎜⎝

⎛−=

ra

raS

raS

raS

ra

raS

raS

r

where, θ = the angle with respect to the load direction r = distance from the hole centre σr = radial stress σθ = circumferential stress τ = in-place shear stress In this analysis, the interest is in the calculation of the stress field around a circular hole under a tensile loading of 2S in the circumferential direction and S in the axial direction. Basically, this model of a defect and its stress perturbation is appropriate for large diameter vessels or pipes where the effect of curvature is negligible. Through the principle of superposition the required stress distribution around a circular hole under a 2:1 tensile loading is given by;


25

)2sin(3212

)2cos(312

12

3

)2cos(3412

12

3

4

4

2

2

4

4

2

2

4

4

2

2

2

2

θτ

θσ

θσ

θ

⎟⎟⎠

⎞⎜⎜⎝

⎛−+=

⎟⎟⎠

⎞⎜⎜⎝

⎛++⎟⎟

⎠

⎞⎜⎜⎝

⎛+=

⎟⎟⎠

⎞⎜⎜⎝

⎛+−−⎟⎟

⎠

⎞⎜⎜⎝

⎛−=

ra

raS

raS

raS

ra

raS

raS

r

The principal stresses around the hole are given by;

22

2,1 22τ

σσσσσ ττ +⎟

⎠⎞

⎜⎝⎛ −

±+

= rr

Inserting the stress profile around the hole in the principal stress equation and limiting the profile to the specific case of θ=0 implies;

)

23

211(

)23

232(

4

4

2

2

2

4

4

2

2

1

ra

raS

ra

raS

−+=

++=

σ

σ (38)

The principal stress decay from around the hole is tabulated in Table 1;

a/r σ1 σ2 1.00 5.00 0.00 0.90 4.20 0.42 0.80 3.57 0.71 0.70 3.10 0.88 0.60 2.73 0.99 0.50 2.47 1.03 0.40 2.28 1.04 0.30 2.15 1.03 0.14 2.03 1.01 0.10 2.02 1.00 0.05 2.00 1.00 0.01 2.00 1.00

Table 1: Stress (principal) decay from a circular hole in a flat plate

Using principal stress, σ1, to define an acceptable stress perturbation, δ, then Equation (38) can be re-written as;

043

43

4

4

2

2=−+ δ

ra

ra


26

Solving for a/r in terms of δ gives (assuming δ<<1);

δ

δ

34

3161

21

21

=

++−=ra

Defining as negligible the stress perturbation of less than 1% implies that the minimum overlap length for a repair over a hole in a flat plate (or large diameter pipe or vessel) is given by;

drimplyingra 42.0 ≥≤ (39)

i.e. the minimum overlay length measured from the centre of the defect should be at least 4 times that of the hole or defect diameter, d. The solution for the energy release rate for a circular defect is based on the deflection of the repair laminate from a flat plate. The solution in AEAT - 57711 does not take account of the curvature of the substrate. The following discussion defines the maximum defect size (circular hole) diameter for which the solution in AEAT - 57711 is valid. The flat plate solution for the energy release rate is based on the theory of pure bending of plates implying that the middle surface of the plate remains at the neutral surface. For this condition the maximum bending strain (in-plane) is given by;

Dti=maxbend,ε

To estimate the maximum curvature of the substrate as a function of defect size before the pure bending flat plate solution is no longer valid the following simple approach is adopted. Consider the following figure, Figure 1. The analysis is based on calculating the equivalent bending strain that is equivalent to the initial curvature based on applied moments at the end of the defect, radius, a.

bM My0 ti

R

θ

a

Figure 1: Schematic diagram of circular defect within a pipe


27

The in-plane bending strain is given by;

θ

θθεRRR

aab

bendingsin−

=−

=

which for small θ becomes

2

22

66 Ra

bending ==θε

Therefore the flat plate solution for the energy release rate is valid when the following inequality is met;

Dt

Ddimplying i

bending << 2

2

6maxbend,εε (40)

Pressure area method for composite repairs to tees and nozzles The fundamental criterion of pressure area method, BS 5500, Annex F, which limits the design pressure for a given repair laminate thickness is given by;

fbfmp A)P.f(A)P.f(PA 5050 −+−≤ (41)

where f = nominal design stress of the repair laminate (given by Ehεc) Ap = pressure loaded area (see Figure 2) Afm = cross-sectional area of repair laminate - main body Afb = cross-sectional area of repair laminate - branch


28

Figure 2: Area definitions for a tee or nozzle

To simplify the analysis of the pressure area method for repairs, the following assumptions are made;

o Only repair material is considered in the analysis (underlying steel carries no load) o Nominal design stress is much greater than one half the internal pressure (induced stress)

Equation (41) simplifies to;

p

fbfm

AAA

fP+

≤ (42)

In effect Equation (42) places an upper limit on the internal pressure carrying capacity of the repair (non-leaking situation). From Figure 2, the following definitions of area apply;

)tL(t)eL(eA

LtLeA

)tL(D)DL(D)eL(d)dL(DA

repairbrepairmbbfb

irepairmmfm

repairbbb

ii

mbio

mi

p

+=+=

==

+++=+++=222222

(43)

where in Equation (43) the first equation is written using the nomenclature of BS 5000 and the second and third equation are written using the nomenclature used throughout this report. For most practical applications;


29

2/DLtL birepairb >>>> and Therefore Equation (43) simplifies to;

brepairfb

irepairfm

bb

ii

p

LtA

LtA

LDLDA

=

=

+=22

(44)

The length of the repair in the branch or main body (assuming unrestricted access) is given by, respectively;

iibi tDLL 2== (45)

where ti is the wall thickness of the main body of the tee. Note, the axial length of the repair is based on the dimensions of the main body and applies to both the main body and the branch. Therefore inserting Equations (44) and (45) into (42) gives

bbii

birepaircc LDLD

LLtEP++

≤ ε2 (46)

For the case on unrestricted access for the repair and branch then Equation (46) simplifies further to;

bi

repaircc DDtEP

+≤

12 ε (47)

Equations (46) or (47) provide an upper limit to the internal pressure that can be applied based on the dimensions of the tee or nozzle and the thickness of the repair in question. Two examples are presented to demonstrate the replacement area method.

Tee Unrestricted access for a tee with dimensions, Di = 273 mm, Db = 168.3 mm, with respective wall thickness, ti = 9.3 and tb = 7.1 mm. Repair thickness, trepair, is 8 mm and hoop modulus, Eh = 20 GPa and allowable strain, εc = 0.0025. Maximum allowable pressure is given by Equation (47).

bar118103168273

800250202 4 .*.

*.**P =+

≤

The repair thickness increase factor, fth, for this tee is given by;


30

462273

39173168

2141

2141

25022502

.....

Dt

tD

.f..

b

bth =⎟

⎟⎠

⎞⎜⎜⎝

⎛⎥⎦⎤

⎢⎣⎡=⎟

⎟

⎠

⎞

⎜⎜

⎝

⎛⎥⎦

⎤⎢⎣

⎡=

For a repair thickness of 8 mm, the design pressure for a leaking repair is 14 bar, assuming a practical value of γ = 100 J/m2. Therefore the pressure area method for limiting the maximum allowable pressure of the repair is consistent with the other design rules outlined in this report.

Vessel nozzle Restricted access for a vessel nozzle on a vessel with dimensions, Di = 1000 mm, Db = 200 mm with respective wall thickness, ti = 12.5 and tb = 7.1 mm. Li=2(Diti)0.5 = 158.1 mm and Lb = 110 mm. Repair thickness, trepair, is 10 mm and hoop modulus, Eh = 20 GPa and allowable strain, εc = 0.0025. Maximum allowable pressure is given by Equation (46).

bar9141011020011581000

11011581000250202 4 .**.*

.*.**P =++

≤

The repair thickness increase factor, fth, for this vessel nozzle is given by;

47351210002

200

5121721

5225221

5225450450

..**

...

.tD

d

tt.

.f..

Di

i

Dd

th =⎟⎟⎠

⎞⎜⎜⎝

⎛

⎟⎠⎞

⎜⎝⎛ +

=⎟⎟⎠

⎞⎜⎜⎝

⎛

⎟⎠⎞⎜

⎝⎛ +

=

The repair thickness increase factor for reduced access on the branch is given by;

441110

1158 ..LL

favailable

designth ===

For a repair thickness of 10 mm, the design pressure for a leaking repair is 8 bar, assuming a practical value of γ = 100 J/m2. Therefore the pressure area method for limiting the maximum allowable pressure of the repair is consistent with the other design rules outlined in this report.


31

Summary of design method and design rules The previous sections have outlines the theory for estimating the increase in loads due to either pipe or vessel component geometry or limited available overlap length. The theory developed is relatively complicated and therefore in these sections it is in an inappropriate form for inclusion into a design guideline. Therefore, the following simplifications have been made in order to reduce the complexity of theory. To repeat the definition of repair thickness for piping system or vessel components, it is given by the product of repair thickness increase factors times the repair thickness for the equivalent straight pipe section, i.e.;


For piping system components, the simplified repair thickness increase factors for design are given by Table 2 with relevant comments;

Piping system

component Repair thickness increase

factor, fth,stress Comment

Bend 1.2

Equations (8) and (9) are simplified to a single equation with an additional factor (20%) added to account for axial loads, from Equation (9)

Tee

2502

2141

.

b

b

Dt

tD. ⎟

⎟

⎠

⎞

⎜⎜

⎝

⎛⎥⎦

⎤⎢⎣

⎡

Minimum value, fth,stress = 1.2 Maximum value, fth,stress = 3 Equation (12) is simplified to a single equation within a minimum value set to 1.2

Flange ⎟⎟⎠

⎞⎜⎜⎝

⎛−+

ftt. 106401

tf > t where tf is the wall thickness of the flange In general pressure loads dominate the performance therefore only the hoop component is used from Equation (14)

Reducer ⎟⎟⎠

⎞⎜⎜⎝

⎛−+ 2

2

106401DD. r

D > Dr where Dr is the (smaller) diameter of the reducer In general pressure loads dominate the performance therefore only the hoop component is used from Equation (14)

Table 2: Repair thickness increase factors for piping system components For the repair of piping components, the maximum allowable design pressure for the repair laminate design thickness, tdesign,component, is restricted to;


32

b

component,designcc

DDtE

P+

≤ε2

(49)

using the pressure area method, Section 0. For cylindrical vessel components, the simplified repair thickness increase factors for design are given by Table 6, with relevant comments;

Cylindrical vessel component

Repair thickness increase factor,

fth,stress Comment

End dome, main body connection 2

203201

dDD.+

D > Dd where Dd is the (smaller) diameter of the dome end In general pressure loads dominate the performance therefore only the hoop component is used from Equation (16)

Supports/saddles/rigid attachments ( )43

2

21 KK

DtDa ++

π

Da is equivalent area of attachment and

( ) ( ) 30504

3

240181

2220

.al

rt.ha

alrt

)C(Log)C(LogCosC..K

C)C(LogCos.K

−⎟⎠⎞

⎜⎝⎛−=

⎟⎠⎞

⎜⎝⎛ −

=

LCC

DC

tDC

CCC a

ali

a

i

irt

a

hha

2,128,2

=⎟⎟⎠

⎞⎜⎜⎝

⎛==

haa CC.D 352= Note: Minimum value, fth,stress = 1.2, maximum value, fth,stress = 3 In general pressure loads dominate the performance therefore only the hoop component is used from Equation (19). Minimum and maximum values are set to be consistent with piping system tees.

Tees/nozzles 450

215225 .

r )t.(. ρ

+

ttt,

DtD b

rb ==

2ρ

Note: Minimum value, fth,stress = 1.2, maximum value, fth,stress = 3 In general pressure loads dominate the performance therefore only the hoop component is used from Equation (20). Minimum and maximum values are set to be consistent with piping system tees.

Table 3: Repair thickness increase factors for cylindrical vessel components


33

Spherical vessel

component Repair thickness increase

factor, fth,stress Comment

Supports/saddles/rigid attachments ( )43

21 KK

DtDa ++

π

Da is equivalent area of attachment and

⎟⎟

⎠

⎞

⎜⎜

⎝

⎛⎥⎦

⎤⎢⎣

⎡−=

⎟⎟

⎠

⎞

⎜⎜

⎝

⎛⎥⎦

⎤⎢⎣

⎡−=

50

4

250

3

28712221

287132380

.a

.a

DtD..Exp.K

DtD..Exp.K


Tees/nozzles ( )

⎥⎥

⎦

⎤

⎢⎢

⎣

⎡ −+310

8809507501

2 .r

.r

tt.. ρ

ttt,

DtD b

rb ==

2ρ


Table 4: Repair thickness increase factors for spherical vessel components

For the repair of either a cylindrical or spherical vessel component, the maximum allowable design pressure for the repair laminate design thickness, tdesign,component, is restricted to;

b

component,designcc

DDtE

P+

≤ε2

(50)

using the pressure area method, Section 0. To account for the limited axial extent of available substrate (Lavailable), which is defined as 0.05 m, the repair thickness increase factor, fth,overlay, is taken as;


34

3/2

, ⎟⎟⎠

⎞⎜⎜⎝

⎛=

available

overoverlayth L

Lf (51)

This formula relates to the stress decay within the repair laminate (Section 0). The stress decay value is taken as the relevant length scale as the stress decay within the laminate is less than within the adhesive layer (Section 0) between the laminate and the substrate.


35

FEA VERIFICATION OF DESIGN RULES The predictions of the repair increase thickness factors from the design formulae presented in Section 0 are compared with finite element stress analysis.

Cylindrical vessel

Figure 3: Finite element mesh of a cylindrical vessel

Figure 3 displays the finite element mesh used for the verification of the design rules for a cylindrical vessel. The dimensions of the vessel and attachments are as follows: Diameter (vessel), Di = 2 m Diameter (branch), Db = 200 mm Wall thickness, ti = 10 mm Internal pressure, P = 1 MPa Length, L = 4m


36

Figure 4: FE results for cylindrical vessel under internal pressure load

To check that the mesh was correctly defined, the average stress within the wall of the vessel was checked against the membrane stress. The membrane stress is given by;

MPa 0012

==i

i

tPD

σ .

Figure 4 plots the average stress within the vessel confirming the membrane stress calculation.

Figure 5: FE predictions of the principal stress around a rigid attachment under internal pressure

The first verification example is the calculation of the stress intensification at the rigid attachment under internal pressure. From Figure 5 the maximum stress intensity or repair thickness increase factor


37

(actual stress divided by the membrane stress) is 132/100 = 1.32. The design formula, Equation (17) predicts a repair thickness increase factor of 2.1.

Figure 6: FE predictions of the principal stress around a nozzle under internal pressure

The second verification example is the calculation of the stress intensification at the nozzle attachment. From Figure 6 the maximum stress intensity or repair thickness increase factor (actual stress divided by the membrane stress) is 256/100 = 2.56. The design formula, Equation (20) predicts a repair thickness increase factor of 2.51.

Figure 7: FE predictions of the principal stress around the main body, end dome attachment

The third verification example is the calculation of the stress intensification at the cylinder body, end dome attachment. From Figure 7 the maximum stress intensity or repair thickness factor is 65/50 =1.3 (based on axial stresses). The design formula, Equation (16) predicts a repair thickness factor of 1.29.


38

Further comparisons between FE analysis and the design rules are listed in Table 5;

Loading FE analysis Design rules Axial tension on nozzle (no pressure) Q=62.8 kN 5 6 (Equation (21))

Bending moment (hoop) on nozzle (no pressure) M= 31.4 kNm 22.6 33 (Equation (22))

Bending moment (axial) on nozzle (no pressure) M= 31.4 kNm 14 23 (Equation (22))

Hydrostatic loading (tank full of water) 1.5 1.2 (Equation (19))

Table 5: Further comparisons of repair thickness increase factors between FE analysis and design rules

Spherical vessel

Figure 8: Finite element mesh of a spherical vessel

Figure 8 displays the finite element mesh used for the verification of the design rule for a spherical vessel. The dimensions of the vessel and attachments are as follows: Diameter (vessel), Di = 2 m Diameter (branch), Db = 200 mm Wall thickness, ti = 10 mm Internal pressure, P = 1 MPa


39

Figure 9: FE results for spherical vessel under internal pressure load

To check that the mesh was correctly defined the average stress within the wall of the vessel was checked against the membrane stress. The membrane stress is given by;

MPa 504

==i

i

tPD

σ

Figure 9 plots the average stress within the vessel confirming the membrane stress calculation.


40

Figure 10: FE predictions of the principal stress around a rigid attachment under internal pressure

The first verification example is the calculation of the stress intensification at the rigid attachment. From Figure 10 the maximum stress intensity or repair thickness increase factor (actual stress divided by the membrane stress) is 81.3/50 = 1.63. The design formula calculation, Equation (25) predicts a repair thickness increase factor of 1.7.


41

Figure 11: FE predictions of the principal stress around a nozzle under internal pressure

The second verification example is the calculation of the stress intensification at the nozzle attachment. From Figure 11 the maximum stress intensity or repair thickness increase factor (actual stress divided by the membrane stress) is 92.3/50 = 1.85. The design formula calculation, Equation (28) predicts a repair thickness increase factor of 2.11 Further comparisons between FE analysis and the design rules are listed in the following Table 6;

Loading FE analysis Design rules Axial tension on nozzle (no pressure) Q=31.4 kN 1.8 1.7 (Equation (29))

Bending moment on nozzle (no pressure) M= 31.4 kNm 30.6 36.4 (Equation (30))

Hydrostatic loading (tank full of water) 0.6 0.8 (Equation (25))

Table 6: Further comparisons of repair thickness increase factors between FE analysis and design rules


42

Overlay length

FE analysis was performed on repairs of different overlay lengths for both 2D and 3D through wall defects. The 2D defect is 20 mm wide and circumferential in extent. The 3D defect is a 20 mm circular through wall defect. The maximum stresses within the substrate from the FE analysis are compared with the repair thickness increase factors as predicted by Section 0 in Figure 12. Also plotted is the predicted stress decay from a circular hole in a flat plate under a 2:1 applied tensile load, Section 0

0

1

2

3

4

5

6

0 10 20 30 40 50 60 70

Overlay length (mm)

Rep

air t

hick

ness

incr

ease

fa

ctor

Design calculations 2D FE calculations3D FE calculations Stress decay from hole

Figure 12: Comparison between FE predictions and design rules

The predictions from the stress decay (in-plane) agree with the FE predictions. The design calculations based on the stress decay within the wall of the vessel are conservative.

Summary of conclusions of FEA analysis The design formulae presented in Sections 0, 0 and 0 at best can only be considered empirical in their derivation. Despite this the comparison with FE predictions is surprisingly good confirming the design approach adopted for the repair of tanks and vessels using composite overwraps. The design rules for increased repair thickness for reduced overlay length are very conservative for 3D defects but are sensibly conservative for 2D defects. The results from the predicted stress decay from a circular hole within a flat plate under an applied tensile load, 2 hoop to 1 axial, is close, as expected to the 2-D FE prediction.


43

EXPERIMENTAL VERIFICATION OF DESIGN RULES To verify the design rules for repairs to pipework components, tanks and vessels, a series of experimental tests on repairs taken to failure were carried out. The experimental verification consisted of two test specimens with repairs to defects at critical locations. The first test specimen was a 150 mm diameter Schedule 80 carbon steel (galvanised) spool (11 mm wall thickness) donated by Walker Technical Resources, Figure 13. It comprised a 2 meter straight pipe section together with two tee connections and two welded elbow or bend sections (a 90º and a 45º).

Figure 13: Pipe test spool

The second test article was a 1 meter diameter welded steel pressure vessel, originally an air receiver, Figure 14. The end closure plates and the inner rubber liner were removed from this vessel.


44

Figure 14: Pressure vessel test spool

Test arrangement The pipe test spool is shown in Figure 13. To close the test spool, three blind flanges were obtained from Linvic Engineering Ltd with the fourth flange specially machined and fitted by, Excel Precision Engineering Ltd. The pressure vessel shown in Figure 14. Surface preparation for all tests was carried out by Hydroblast – Industrial & Scientific Cleaning to specification, SA2½. On-site pressure testing was carried out by Johnson Controls Ltd (Harwell Testing). The pressure delivery system was initially a hand pump (max capacity 40 bar) followed by a compressed air powered Hydratron pump (max capacity 160 bar). The pressurizing fluid for all tests was water.


45

Figure 15: On-site pressure testing carried out by Johnson Controls Ltd (Harwell Testing)

Defects Within the pipe spool, a total of five defects were introduced. The pipe was drilled with 5 mm holes at selected locations and each hole covered with a 30 mm diameter PTFE disc. The defects were located at the extrados of the 45 and 90 bends, at the base of the end flange and at the base of the tee. A reference defect was also placed mid-way in the straight pipe section. Figure 16 presents photographs of the defect location details. Within the vessel, a total of four defects were introduced. The vessel was drilled with 1/8 inch (3.175 mm) NPT mm (threaded) holes at selected locations and each hole covered with a 30 mm diameter PTFE disc. The defects were located at both end dome, main body attachments and a distance of 30 mm from both end flanges. Figure 17 presents photographs of the defect location details.

Defect 1 (450 bend)

Defect 2 (900 bend)

Reference defect

Defect 3 (flange) Defect 4 (tee)


46

Figure 16: Defect details and location for the pipe spool

Repair design The design of all supplied repairs was performed by ESR Technology. The through wall defect design case was based on mean energy releases rates derived from the Repair Suppliers qualification test data. Pressure testing For the pipe spool, the pressure was raised until the first repair failed. This repair was removed and the defect plugged. The spool was then re-pressurized until the next repair failed. The test was stopped on failure of the second repair. For the vessel, only blow-off tests on each repair were performed. The vessel itself was not pressurized. Both short and medium term duration tests were performed. In the short-term tests, the pressure was raised in 5 bar increments with a hold time of 2 minutes. In the medium term tests, the pressure was raised from an initial pressure of between 5 to 10 bar by a rate of 2 bar/day, with the aim that the repairs failed after approximately 1000 hours. This type of test is termed a low speed loading rate test. The details of this test procedure and analysis of the results is contained within AEAT/57394.

Figure 17: Defect details and location for the vessel

Defect 3 (flange)

Defect 4 (tee)

200 mm

Defect 1 (450 bend)

Defect 2 (900 bend)

Reference defect

Defect 230 mm

Defect 1 50 mm Defect 3

50 mm

Defect 4 30 mm


47

Repair Supplier 1 – test results and analysis Short term test results and analysis The design details of each Repair Supplier 1 repair for the pipe spool are given in Table 7 including thickness, repair thickness increase factor and axial extent. Also included in Table 7 are the predicted and measured failure pressures. All pressures are normalised to the predicted mean failure pressure of the thinnest repair (4.48 mm) with a repair thickness increase factor of unity.

Defect number

Repair

thickness increase

factor, f th

Repair thickness

(mm)

Axial Extent (mm)

Predicted mean

failure pressure

Predicted UCL

failure pressure

Measured failure

pressure

1 - 45º elbow 1.2 Equation (9) 4.48 100 1/1.2 =

0.83 1.04/1.2 =

0.87 1.31

2 - 90º elbow 1.2 Equation (9) 4.48 100 1/1.2 =

0.83 1.04/1.2 =

0.87 Did not fail

3 - End flange 1.2 Equation (36) 5.6 100 1.19/1.2

= 0.99 1.24/1.2 = 1.03 Did not fail

4 - Tee 1.53 Equation (11) 6.72 100 1.34/1.53

= 0.88 1.39/1.53

= 0.91 Did not fail

5 - Reference 1 4.48

100 (hoop

direction also)

1 1.04 1.55

Table 7: Repair details and test results for pipe spool – short-term tests

Predicted failure pressures are based on both mean and upper 95% confidence level energy release rates. The measured failure pressures were higher than anticipated. The predicted failure pressures were 0.83 (450 elbow) and 1 (reference defect). The measured failure pressures were 1.31 (450 elbow) and 1.55 (reference defect). The failure mode of both repairs was leakage from the edge of the repair at or close to the interface through the tapered end, implying that the mode of failure is interfacial delamination. The design details of each Repair Supplier 1 repair for the vessel are given in Table 8 including thickness, repair thickness increase factor and axial extent. Also included in Table 8 are the predicted and measured failure pressures, with the same normalization as Table 7.


48

Defect number

Repair

thickness increase

factor, f th

Repair thickness

(mm)

Axial extent (mm)

Predicted mean

failure pressure

Predicted UCL

failure pressure

Measured failure

pressure

1 – Cylinder - top 1 4.48 100 1 1.04

Stripped thread, no

result

2 – Dome - top

2 Equation (36) 6.72

100 (except up to

flange face, 30 mm)

1.34/2 = 0.67

1.39/2 = 0.69 1.82

3 – Cylinder - base 1 4.48 100 1 1.04 3.23

4 – Dome - base

2.5 Equation (36) 7.84 30 1.47/2.5

= 0.59 1.53/2.5 = 0.61 1.82

Table 8: Repair details and test results for vessel – short-term tests

Predicted failure pressures are based on both mean and upper 95% confidence level energy release rates. The measured failure pressures were significantly higher than predicted, approximately 3 times the expected value. The predicted failure pressures were 0.67 (defect 2 – dome-top), 1 (defect 3 – cylinder-base) and 0.59 (defect 4 – dome-base). Measured failure pressures were 1.82, 3.23, 1.82 respectively. The failure mode of all repairs was leakage from the edge of the repair, i.e. interfacial delamination.

Medium term test results and analysis The design details of each Repair Supplier 1 repair for the pipe spool are given in Table 7. The results of the medium term test are presented in Table 9 and Figure 18. The spool was pressurised at a rate of 2 bar/day with an initial pressure of 10 bar. The results are normalised with respect to the predicted short-term failure pressure for that particular repair. In estimating the performance of the repair it is assumed that the regression behaviour of the repair is independent of the geometry of the test specimens. Therefore the regression curve measured from the plain pipe spools is also assumed to apply for the pipe spool tests.


49

Defect number

Repair thickness

increase factor, f th

Time to failure (hours)

Predicted mean failure

pressure

Measured failure

pressure

1 - 45º elbow 1.2 Equation (9) Did not fail - -

2 - 90º elbow 1.2 Equation (9) 480 0.66 1.03

3 - End flange 1.2 Equation (36) Did not fail - -

4 - Tee 1.53 Equation (11) 480 0.66 1.06

5 - Reference 1 Did not fail - -

Table 9: Repair details for pipe test spool – medium-term tests

In Figure 18, the regression curve measured for the plain pipe spool is shown as the solid black line (taken from AEAT/573894 [5]) and the normalised data used to generate that regression curve (solid black triangles).

00.10.20.30.40.50.60.70.80.9

1

0 0.5 1 1.5 2 2.5 3 3.5 4 4.5

Log (hours)

Nor

mal

ised

Pres

sure

Normalised regression curve from previous LSL testsNormalised data from original LSL tests1st defect to fail - 90 bend - assumed RTIF = 1.22nd defect to fail - tee - assumed RTIF = 1.53

Figure 18: Medium-term test results

The first repair to fail was at the 900 bend. The second repair to fail was at the tee. Failure times were after 480 hours. For both repairs the measured failure pressures, 1.03 and 1.06 were higher than predicted, 0.66 and 0.66, respectively. The failure mode of both repairs was leakage from the edge of the repair implying that the mode of failure is interfacial delamination. Figure 19 presents a photograph of the failure mode.


50

Figure 19: Failure mode of repair at 900 bend

The design details of each Repair Supplier 1 repair for the vessel are given in Table 8. Only defect numbers 1 and 2 were used in these tests. The results of the medium term tests are presented in Table 10 and Figure 20. The repair was pressurized at a rate of 2 bar/day with an initial pressure of 10 bar. The results are normalized with respect to the short-term failure pressure for that particular repair and it is assumed that the regression curve measured for the plain pipe spool applies.

Defect number

Repair

thickness increase factor,

f th


Predicted mean

failure pressure

Measured failure

pressure

1 – Cylinder –

top 1 720 0.64 1.07

2 – Dome – top

2 Equation (36) 336 0.68 0.91

Table 10: Repair details for vessel – medium-term tests


51

In Figure 20, the regression curve measured for the plain pipe spool is shown as the solid black line (taken from AEAT/573894 [5]) and the normalized data used to generate that regression curve (solid black triangles).

00.10.20.30.40.50.60.70.80.9

1

0 0.5 1 1.5 2 2.5 3 3.5 4 4.5

Log (hours)

Nor

mal

ised

Pre

ssur

e

Normalised regression curve from previous LSL testsNormalised data from original LSL tests1st defect to fail - defect 2 - assumed RTIF = 22nd defect to fail - defect 1 - assumed RTIF = 1

Figure 20: Medium-term test results

The first repair to fail was at defect 2 – dome-top after 336 hours. The second repair to fail was at defect 1 – cylinder-top after 720 hours. For both repairs the measured failure pressures, 0.68 and 0.64 were higher than predicted, 0.91 and 1.07. The failure mode of both repairs was leakage from the edge of the repair, i.e. interfacial delamination. Figure 21 presents a photograph of the failure mode.


52

Figure 21: Failure mode of repair at defect 2

Repair Supplier 2 – test results and analysis Short term test results and analysis The design details of each Repair Supplier 2 repair for the pipe spool are given in Table 11 including thickness, repair thickness increase factor and axial extent. Also included in Table 11 are the predicted and measured failure pressures. All pressures are normalized to the predicted mean failure pressure of the thinnest repair (3.2 mm) with a repair thickness increase factor of unity.


53

Defect number

Repair

thickness increase

factor, f th

Repair thickness

(mm)

Axial Extent (mm)

Predicted mean

failure pressure

Predicted UCL

failure pressure

Measured failure

pressure

1 - 45º elbow

1.2 Equation (9) 3.2 100 1/1.2 = 0.83 1.3/1.2 =

1.08 1.45

2 - 90º elbow

1.2 Equation (9) 3.2 100 1/1.2 = 0.83 1.3/1.2 =

1.08 Did not fail

at 2.9 3 - End flange

1.2 Equation (36) 4.0 100 1.26/1.2

= 1.05 1.65/1.2 =

1.375 Did not fail

at 2.9

4 – Tee 1.53 Equation (11) 4.0 100 1.26/1.53

= 0.81 1.65/1.53 =

1.08 Did not fail

at 2.9

5 – Reference 1 3.2

100 (hoop direction

also) 1 0.7 Did not fail

at 2.9

Table 11: Repair details for pipe test spool – short-term tests

Predicted failure pressures are based on both mean and upper 95% confidence level energy release rates. The measured failure pressure of the 450 elbow, 1.45 is greater than the predicted mean failure pressure of 0.83. The applied pressure was further increased to the maximum allowable of the test spool, 2.9. All the remaining repairs survived this pressure. The failure mode of the repair was interfacial delamination. The design details of each Repair Supplier 2 repair for the vessel are given in Table 12 including thickness, repair thickness increase factor and axial extent. Also included in Table 12 are the predicted and measured failure pressures, with the same normalization as Table 11.

Defect number

Repair

thickness increase

factor, f th

Repair thickness

(mm)

Axial extent (mm)

Predicted mean

failure pressure

Predicted UCL

failure pressure

Measured failure

pressure

1 – Cylinder –

top 1 3.2 100 1 1.3

0

2 – Dome – top

2 Equation (36) 3.2

100 (except up to

flange face, 30 mm)

1/2 = 0.5

1.3/2 = 0.65 1.22

3 – Cylinder –

base

1.8 Equation (36) 4.8

100 (except up to

flange face, 50 mm)

1.49/1.8 = 0.82

1.94/1.8 = 1.08

0

4 – Dome – base

2.5 Equation (36) 6.4 30 1.85/2.5

= 0.74 2.41/2.5 = 0.96 1.16

Table 12: Repair details for vessel – short-term tests


54

Predicted failure pressures are based on both mean and upper 95% confidence level energy release rates. The measured failure pressures for defect 4 – dome-base and defect 2- dome-top were greater than predicted, between 1.6 and 2.5 times. The predicted failure pressures were 0.5 (defect 2 – dome-top) and 0.74 (defect 4 – dome-base). Measured failure pressures were 1.22 and 1.16 respectively. The failure mode of both repairs was leakage from the edge of the repair, i.e. interfacial delamination. However for repairs to defects 1 and 3, which include the external weld bead between the cylinder main body and the dome ends, both repairs failed on immediate pressurization, interfacially debonding along the weld line. The weld bead protrusion was the cause of the problem.

Medium term test results and analysis The design details of each Repair Supplier 2 repair for the pipe spool are given in Table 11. The results of the medium term test are presented in Table 13 and Figure 20. The spool was pressurized at a rate of 2 bar/day with an initial pressure of 10 bar. The results are normalized with respect to the short-term failure pressure for that particular repair. In estimating the performance of the repair it is assumed that the regression behavior of the repair is independent of the geometry of the test specimens. Therefore the regression curve measured from the plain pipe spools is also assumed to apply for the pipe spool tests. In Figure 20, the regression curve measured for the plain pipe spool is shown as the solid black line (taken from AEAT/573894 [5]) and the normalized data used to generate that regression curve (solid black triangles). The first repair to fail was at the reference defect after 432 hours. The second repair to fail was at the tee after 960 hours. For both repairs the measured failure pressures, 0.76 and 1.28 were equal to or higher than predicted, 0.76 and 0.73, respectively.

Defect number

Repair thickness




pressure

Measured failure

pressure



3 - End flange 1.2 Equation (36) Did not fail - -

4 - Tee 1.53 Equation (11) 960 0.73 1.28

5 - Reference 1 432 0.76 0.76

Table 13: Repair details for pipe test spool – medium term tests


55

00.10.20.30.40.50.60.70.80.9

1

0 0.5 1 1.5 2 2.5 3 3.5 4 4.5

Log (hours)

Nor

mal

ised

Pre

ssur

e

Normalised regression curve from previous LSL testsNormalised data from original LSL tests1st defect to fail - reference defect - assumed RTIF = 12nd defect to fail - tee - assumed RTIF = 1.5

Figure 22: Medium term test results

The failure mode was leakage from the edge of the repair implying that the mode of failure is interfacial delamination. Figure 23 presents a photograph of the failure mode.

Figure 23: Failure mode of repair of reference defect


56

The tests on repairs to the vessel were unable to be performed as the threads on the pressure delivery attachment fittings would not seal.

Repair Supplier 3 – test results and analysis Short term test results and analysis The design details of each Repair Supplier 3 repair for the pipe spool are given in Table 14 including thickness, repair thickness increase factor and axial extent. Also included in Table 14 are the predicted and measured failure pressures. All pressures are normalized to the predicted mean failure pressure of the thinnest repair (3 mm) with a repair thickness increase factor of unity.

Defect

number

Repair

thickness increase

factor, f th

Repair thickness

(mm)

Axial Extent (mm)

Predicted mean

failure pressure

Predicted mean UCL

pressure

Measured failure

pressure

1 - 45º elbow

1.2 Equation (9) 3 100 1/1.2 =

0.83 1.04/1.2 =

0.87 2.18

2 - 90º elbow

1.2 Equation (9) 3 100 1/1.2 =

0.83 1.04/1.2 =

0.87 Did not fail

3 - End flange

1.27 Equation (36) 4 100 1.27/1.27

= 1 1.32/1.27

= 1.04 Did not fail

4 - Tee 1.53 Equation (11) 6 100 1.79/1.53

= 1.17 1.86/1.53

= 1.21 Did not fail

5 - Reference 1 3

100 (hoop direction

also) 1 1.04 Did not fail

Table 14: Repair details for pipe test spool – short-term tests

Predicted failure pressures are based on both mean and upper 95% confidence level energy release rates. The measured failure pressure of the 450 elbow, 2.18 is greater than the predicted mean failure pressure of 0.83. The applied pressure was further increased to the maximum allowable of the test spool, 3.2. All the remaining repairs survived this pressure. The failure mode of the repair was interfacial delamination. The design details of each Repair Supplier 3 repair for the vessel are given in Table 15 including thickness, repair thickness increase factor and axial extent. Also included in Table 15 are the predicted and measured failure pressures, with the same normalization as Table 14.


57

Defect number

Repair

thickness increase

factor, f th

Repair thickness

(mm)

Axial extent (mm)

Predicted mean

failure pressure

Predicted UCL

failure pressure

Measured failure

pressure

1 – Cylinder - top 1 3 100 1 1.04 0.6

2 – Dome - top

2 Equation (36) 6

100 (except up to

flange face, 30 mm)

1.68/2 = 0.84

1.75/2 = 0.87 No repair

3 – Cylinder - base 1 3 100 1 1.04 1.2

4 – Dome - base

2.5 Equation (36) 6 30 1.68/2.5

= 0.67 1.75/2.5

= 0.7 No repair

Table 15: Repair details for vessel – short-term tests

Predicted failure pressures are based on both mean and upper 95% confidence level energy release rates. Repairs were not applied to defect 2 – dome-top and defect 4 – dome-base as when Repair Supplier 3 arrived on site, the surface preparation was only available for defect 1 – cylinder-top and defect 3 – cylinder base. The measured failure pressures for defect 3 – cylinder-base was greater than predicted, by a factor of 1.15. The predicted failure pressure was 1 compared to the measured failure pressure of 1.2. The failure mode was interfacial delamination. However for defect 1 – cylinder top, which includes the external weld bead between the cylinder main body and the dome end, the repair failed at a pressure of 0.6 compared to a predicted value of 1.04. The repair failure was interfacial debonding along the weld line. The weld bead protrusion was the cause of the problem.

Medium term test results and analysis The design details of each Repair Supplier 3 repair for the pipe spool are given in Table 14. The results of the medium term test are presented in Table 16 and Figure 24. The spool was pressurised at a rate of 2 bar/day with an initial pressure of 5 bar. The results are normalised with respect to the short-term failure pressure for that particular repair. It should be noted that the reinforcement used in these repairs was an organic fibre whereas in the short-term tests glass fibres were used. Results were scaled accordingly to account for this change. In estimating the performance of the repair it is assumed that the regression behaviour of the repair is independent of the geometry of the test specimens. Therefore the regression curve measured from the plain pipe spools is also assumed to apply for the pipe spool tests.


58

In Figure 24, the regression curve measured for the plain pipe spool is shown as the solid black line (taken from AEAT/573894 [5]) and the normalised data used to generate that regression curve (solid black triangles).

Defect number

Repair thickness




pressure

Measured failure

pressure



3 - End flange 1.2 Equation (36) 264 0.81 0.77

4 - Tee 1.53 Equation (11) Did not fail - -

5 - Reference 1 432 0.8 0.82

Table 16: Repair details for pipe test spool – medium term tests

00.10.20.30.40.50.60.70.80.9

1

0 0.5 1 1.5 2 2.5 3 3.5 4 4.5

Log (hours)

Nor

mal

ised

Pres

sure

Normalised regression curve from previous LSL testsNormalised data from original LSL tests1st defect to fail - flange - assumed RTIF = 1.22nd defect to fail - reference defect - assumed RTIF = 1

Figure 24: Medium term test results

The first repair to fail was at the end flange after 264 hours. The second repair to fail was at the reference defect at 432 hours. For both repairs the measured failure pressures, 0.77 and 0.82 were equal to predicted, 0.81 and 0.8, respectively. The failure mode of both repairs was interfacial delamination. Figure 25 presents a photograph of the failure mode.


59

Figure 25: Failure mode of repair to flange

The design details of each Repair Supplier 3 repair for the vessel are given in Table 15. Only defect numbers 1 and 2 were used in these tests. The results of the medium term test are presented in Table 17 and Figure 26. The repair was pressurized at a rate of 2 bar/day with an initial pressure of 5 bar. The results are normalized with respect to the short-term failure pressure for that particular repair and it is assumed that the regression curve measured for the plain pipe spool applies. In Figure 26, the regression curve measured for the plain pipe spool is shown as the solid black line (taken from AEAT/573894 [5]) and the normalized data used to generate that regression curve (solid black triangles).

Defect number

Repair thickness




pressure

Measured failure pressure

1 – Cylinder - top 1 504 0.80 0.88

2 – Dome - top 2 Equation (36) 360 0.81 1.14

Table 17: Repair details for vessel – medium term tests


60

00.10.20.30.40.50.60.70.80.9

11.1

0 0.5 1 1.5 2 2.5 3 3.5 4 4.5

Log (hours)

Nor

mal

ised

Pre

ssur

e

Normalised regression curve from previous LSL testsNormalised data from original LSL tests1st defect to fail - defect 2 - assumed RTIF = 22nd defect to fail - defect 1 - assumed RTIF = 1

Figure 26: Medium term test results The first repair to fail was at defect number 2 after 360 hours. The second repair to fail was at defect number 1 after 500 hours. For both repairs the measured failure pressures, 1.14 and 0.88 were equal to or higher than predicted, 0.81 and 0.8, respectively. The failure mode was interfacial delamination. Figure 27 presents a photograph of the failure mode.


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Figure 27: Failure mode of repair of defect 1

Generic conclusions Three repair suppliers were invited to carry out repairs on test spools containing defects, which are representative of engineering equipment found in service offshore and in process plant. Glass, polymeric and carbon fiber repair systems were investigated. A series of pressure tests were carried out on a 6” Schedule 80 pipe spool containing 5 defects including a 45º and 90º elbow, flange and a T-connector and a pressure vessel containing 4 defects. The pressure tests included both short-term and medium-term tests to failure. For the piping spool, the pressure was increased until one of the five repairs failed. This failed repair was removed, the defect plugged and the pressure test repeated until the next repair failed. Results from both short-term and medium-term tests are presented in a normalized form, the normalization with respect to the short-term predicted failure pressure of the repair. Overall the measured failure pressures were higher than predicted; suggesting the theory for the design or repairs using repair thickness increase factors is conservative. This statement is independent of the repair type. Also, it


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was confirmed that the regression behavior (i.e. reduction in strength as a function of time) of the repairs to pipe components and vessel attachments is comparable to that on plain pipe spools. It is also relevant to note that the reference defect on the pipe spool was repaired using a saddle type of repair rather than a standard fully circumferential repair. The test results show that the measured failure pressures for the saddle type of repair were no different to those from a fully circumferential repair. The tests on the vessel repairs were blow off tests, i.e. only pressure was applied to the repair until failure and not to the vessel itself. Results are presented in a normalized form, the normalization as for the piping spool repairs. Overall the measured failure pressures were higher than predicted suggesting the theory for the design of repairs is conservative. Again this conclusion is independent of repair type. Also, the regression behavior of the repairs was comparable to those repairs on the piping spool. The failure mode of all tests was interfacial delamination with leakage occurring at the edge of the repair typically a surface distance of up to 100 mm from the defect. In general, the conclusion of these tests is that the theory developed for the design of composite repairs to piping components and vessels is conservative. This is confirmed by test results from both short-term and medium-term tests. From the tests conducted 2 further conclusions can be made; 1) Limited axial length of repair (i.e. less than 50 mm) does not cause premature failure of the

repair. Tests confirmed that the design approach using repair thickness increase factors is valid. 2) The external protrusion of the weld bead in some tests caused premature failures of the repairs.

Grinding off or faring the repair (if grinding not practical) of these external protrusions should be considered prior to repair application.


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GENERAL CONCLUSIONS AND RECOMMENDATIONS This report has summarized the development work for extending the design guidelines for composite repairs for pipework to cover also piping system components, e.g. tees, bends etc., and tanks and vessels. For tanks and vessels, the focus is on nozzles and attachments rather then the main body. The development work consisted of 4 phases

• Definition of repair thickness increase factors • Finite element stress analysis (FEA) • Experimental results • Implications on design approach

The design development involved introducing the concept of repair thickness increase factors. These factors can be considered comparable to stress intensity factors as used in piping system design. The idea behind the design process is that the repair thickness is calculated for a straight pipe section, same operating conditions. Repair thickness increase factors are calculated based on the simplified geometry of the component. The final repair thickness is the reference repair thickness times the repair thickness increase factor. Predictions of the repair thickness increase factor were compared with a range of predictions from finite element stress analysis. In all cases studied for piping systems and vessel attachments the design predictions were always greater than or equal to the FEA predictions, i.e. the proposed design approach is conservative. Repair suppliers were invited to apply repairs to defects on both a piping spool and a pressure vessel. Defects (30 mm diameter through wall) were located at various locations within both spools. Failure pressures were measured and compared with predictions for both short and medium term duration tests. In all cases predictions were less than measurements providing confirmation that the design approach is conservative. Based on a comparison with both FEA predictions and experimental measurements the design approach of using repair thickness increase factors can be considered conservative. This approach for the design of repairs to piping system components and vessel attachments and nozzles has been implemented within the design guidelines for overwrap repairs (AEAT – 57711).


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BIBLIOGRAPHY Background standards and references are given in the following sections. Reference Standards

1) Theory of plates and shells, S. Timoshenko and S. Woinowsky-Krieger, McGraw-Hill 1959. 2) ISO 14692 – Petroleum and Natural Gas Industries – Glass Re-inforced Plastics (GRP) piping 3) ISO 15649, Petroleum and natural gas industries -- Piping 4) ISO 13623, Petroleum and natural gas industries -- Pipeline transportation systems 5) Structural integrity of beams strengthened with FRP plates – Analysis of the adhesive layer, S

Frost, R. Lee, V. Thompson, Structural Faults and Repairs, 2003. 6) Theory of elasticity, S. Timoshenko and J.N. Goodier, McGraw-Hill 1951. 7) prEN 13121, GRP tanks and vessels for use above ground. 8) API 579, Recommended practice for fitness for service. 9) BS 7910, Guide on methods for assessing the acceptability of flaws in fusion welded

structures 10) BS 8010, Code of practice for pipelines. Pipelines on land 11) ASME B31.1, Power Piping 12) ASME B31.3, Chemical plant and refinery piping 13) ASME B31.4, Pipeline Transportation Systems for Liquid Hydrocarbons and Other Liquids 14) ASME B31.8, Gas Transmission and Distribution Piping Systems 15) ASME Pressure Vessel and Boiler Code 16) BS 4994, Specification for design and construction of vessels and tanks in reinforced plastics 17) BS 5500, Specification for unfired fusion welded pressure vessels

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