forensic investigation of compression...

FORENSIC

INVESTIGATION

OF

COMPRESSION

FITTING

FAILURES

W.R. Shipway

September 2011

In partial fulfilment of the requirements

for the degree Master of Science in

Advanced Mechanical Engineering from

Brunel School of Engineering and Design

Copyright 2011 © W.R. Shipway

Forensic Investigation of Compression Fitting Failures W.R. Shipway

ii

Mr. W. Shipway

113a The Avenue

Ealing

London

W13 8JT

Tel. 07726020166

e-mail: [email protected]

Dr. Giulio Alfano - Lecturer School of Engineering & Design Brunel University, Uxbridge Middlesex, UB8 3PH, UK Tel: +44(0)1895 267062 Fax: +44(0)1895 256392 e-mail: [email protected]

Brunel University

Kingston Lane

Uxbridge

Middlesex

UB8 3PH

tel: +44 (0)1895 274000


iii

Acknowledgements

Thank you to Dr. G. Alfano who was the advisor on this project for his time and guidance

at all stages, particularly the late Friday evening‟s where by-right he should have been

elsewhere. Thank you to Costas Thantos who assisted with the set-up of the experimental

rig and Tony McCalla for materials procurement.


iv

Table of Contents

Acknowledgements .............................................................................................................. iii

List of Figures ...................................................................................................................... vii

List of Tables......................................................................................................................... ix

Nomenclature ......................................................................................................................... x

1. Introduction ........................................................................................................................ 1

1.1 Forensic Engineering ....................................................................................................... 1

1.2 Compression Fittings ....................................................................................................... 2

1.3 Aims and Objectives ........................................................................................................ 3

1.4 Methodology .................................................................................................................... 4

1.4.1 Experimental Testing .................................................................................................... 4

1.4.2 Computational Modelling ............................................................................................. 4

1.4.3 Analytical Engineering.................................................................................................. 4

2. Technical and Literature Review ....................................................................................... 5

2.1 Technical Review ............................................................................................................. 5

2.1.1 Coupling Requirements ................................................................................................. 5

2.1.2 Sealing Mechanism ..................................................... Error! Bookmark not defined.


v

2.1.3 Applications .................................................................................................................. 7

2.1.4 Materials ........................................................................................................................ 8

2.1.5 Failure mechanisms ..................................................................................................... 10

2.2 Literature Review and Analytical Analysis ................... Error! Bookmark not defined.

2.2.1 Continuum Analysis .................................................... Error! Bookmark not defined.

2.2.2 Microscopic Analysis .................................................................................................. 17

3. Methods ............................................................................................................................ 23

3.1 Experimental Testing ..................................................................................................... 23

3.1.1 Apparatus .................................................................................................................... 23

3.1.2 Experimental Procedure .............................................................................................. 23

3.2 Computational Modelling .............................................................................................. 27

3.2.1 Continuum Modelling ................................................................................................. 27

3.2.2 Materials - Introducing Strain Hardening ................................................................... 28

3.2.3 Elements - Axisymmetric Model of the Specimen ..................................................... 32

3.2.4 Model Parameters........................................................................................................ 33

3.2.5 Convergence ................................................................ Error! Bookmark not defined.

4. Results and Discussion ..................................................................................................... 40

4.1 Experimental Results and Discussion ............................................................................ 40


vi

4.2 FE Simulation Results and Discussion .......................................................................... 46

4.3 Comparison between FE simulation and Experimental Results – Direct Method ......... 52

4.4 Discussion on the Engineering Analysis ........................................................................ 54

4.5 General Discussion ........................................................ Error! Bookmark not defined.

4.6 Further Analysis ............................................................. Error! Bookmark not defined.

4.6.1 Experimental Investigation ......................................... Error! Bookmark not defined.

4.6.2 Engineering Analysis .................................................. Error! Bookmark not defined.

4.6.3 FE Simulation ............................................................................................................. 61

5. Conclusions and Recommendations ................................................................................ 62

6. References ........................................................................................................................ 65

7. Bibliography ..................................................................................................................... 70

8. Appendix .......................................................................................................................... 71

8.1 Results Data ................................................................................................................... 72

8.2 Project Management ...................................................... Error! Bookmark not defined.

8.3. Email correspondence with Conex ............................................................................... 77


vii

List of Figures

Figure 1.1. A typical type-A compression fitting for domestic use. ...................................... 3

Figure 2.1. Designs and working environment for compression fittings. .............................. 5

Figure 2.2. Type-A compression coupling. ............................................................................ 7

Figure 2.3. Type-B compression fitting, requiring a flared pipe end [12]. ............................ 7

Figure 2.4. A cylinder submitted to internal and external pressure. .................................... 13

Figure 2.5. Compound cylinder. .......................................... Error! Bookmark not defined.

Figure 2.6. A uniform band of pressure. .............................................................................. 15

Figure 2.7. Two contacting bodies with point‟s m and n touching upon the exchange of a

force...................................................................................................................................... 18

Figure 2.8. Example asperity population [48]. ..................................................................... 18

Figure 2.9. Possible fluid flow paths. ................................................................................... 22

Figure 3.1. Experimental setup for hydrostatic testing. ....................................................... 26

Figure 3.2. A simple elastic model....................................... Error! Bookmark not defined.

Figure 3.3. True stress-true strain curves for various elastic-perfectly plastic (EPP) and

strain hardened models. ........................................................................................................ 32

Figure 3.4. Finite element mesh, with parameters for various partitions. ............................ 33

Figure 3.5. Reference nodes and their constraints to the thread edges. ............................... 35

Figure 3.6. Interactions between the parts, highlighted. ...... Error! Bookmark not defined.


viii

Figure 3.7. Boundary conditions imposed on the model ..................................................... 38

Figure 3.8. The initial model of the von-Mises stress results, and highlighted areas of

interest. ................................................................................................................................. 38

Figure 3.9. Element inspection points for mesh convergence check. .................................. 38

Figure 3.10. Mesh convergence results, with no. of nodes used in the simulation

highlighted in red. ................................................................................................................ 39

Figure 4.1. Low leakage rate initiating at the top of the pipe from measurement 1. ........... 43

Figure 4.2. Measurement 10. The specimen ferrule popped of its seat, but maintained a

seal. Note visible ring of deformation creating the seat. ...................................................... 43

Figure 4.3. Pipe separation causing score marks as the ferrule moves axially along the pipe.

.............................................................................................................................................. 43

Figure 4.4. Incorrect ferrule positioning from measurement 11, causing a below average

failure pressure. .................................................................................................................... 44

Figure 4.5. Cut section of pipe at ferrule interaction, named as „deformation seat‟. ........... 44

Figure 4.6. Mini-max contact pressure inquiry locations. ................................................... 47

Figure 4.7. Experimental results of failure pressures and FE simulation of qcmax

. .............. 47

Figure 4.8. Coarse meshes producing inaccurate stress development deformation due to the

finite nature of FEM. ............................................................................................................ 50

Figure 4.9. Comparisons of the ferrule deformation. ........................................................... 53

Figure 4.10. Compression fitting designs which do not conform to the Kitemark. ............. 58

Figure 8.1. Above 2000psi the pipes structural integrity was compromised. ...................... 78


ix

List of Tables

Table 2.1. Mechanical properties of compression fittings used in water utilities .................. 9

Table 3.1. Constraint equation input in Abaqus ................... Error! Bookmark not defined.

Table 4.1 Observed failure mode at particular tests. ............ Error! Bookmark not defined.

Table 8.1. Steps in a forensic investigation outlined by S. Brown [1]. ................................ 71

Table 8.2. Uniaxial test parameters and results.................................................................... 72

Table 8.3. Experimental results. ........................................................................................... 72

Table 8.4. Statistical evaluation of results. .......................... Error! Bookmark not defined.

Table 8.5. Contact pressure at locations shown in Fig. 4.6. . Error! Bookmark not defined.

Table 8.6. Measured displacement-thread comparisons. ..... Error! Bookmark not defined.

Table 8.7. A Gantt chart timeline. ........................................................................................ 76


x

Nomenclature

Standard symbols are used throughout this report, and the reader must be aware of fluid or

solid mechanics conventions, as some symbols are conflicting.

Thermodynamics and Fluid Mechanics:

A – area, normal to force of fluid flow

∂ – partial change

d – infinitesimal change

∆ – incremental change

F – force

k – permeability of a porous material

KE – kinetic energy

L – length of pressure drop

p – pressure at point of inspection

PE – potential energy

Q – heat transfer, or leakage rate

V – volume

W – work

x – displacement

μ – viscosity


xi

Solid Mechanics:

A – area, normal to force

r – radii, varying (under consideration)

ε, γ – strain components: direct; shear

δ - difference

F – force

g – average asperity

E, G, K, n, UTS, μ, ν, Y – material constants: elastic modulus; shear modulus, strength

coefficient (for strain hardening) or spring stiffness constant; strain hardening exponent;

tensile strength, coefficient of friction; Poisson‟s ratio; yield strength

L – length

N – No. of cycles or No. of terms of a linear equation

q – intensity of a uniformly distributed load

S – compliance matrix

ζ, τ – stress components: direct; shear

t – thickness

u, v, w – displacement components, cylindrical coordinate system

x, y, z – displacement components, Cartesian coordinate system


xii

Superscripts and Subscripts:

x, y, z – Cartesian coordinates

1, 2 3 – principal coordinates

a – annealed (material)

atm – atmospheric

c – contact

col – collar

cp – compound

cz – brass (alloy of copper and zinc)

Cu – copper

θ – tangential

e – elastic

f – fracture

i – internal

int – interference

l – leakage

o – external

p – plastic or pre-strained (material)

PEEQ – equivalent plastic strain

r – radial

s – shaft (pipe)

vM – von Mises stress invariant

z – axial


1

Abstract

A preliminary examination of compression fitting failures is reported. Forensic analysis for litigation

purposes is conducted on one of the most relevant situations concerning their use – correct installation.

Specifically, the applied nut tightness and its consequent failure pressure. Analytical methods are described,

and used to convey the means of the sealing mechanism, via a metal to metal mechanical face seal, or static

seal. A geometric and material non-linear finite element (FE) simulation using Abaqus is constructed and

validated by experimental deformation measurements. Experimental data is produced in accordance with BS

EN1254-2:1998 – Plumbing fittings. The results are coupled with the FE simulation to suggest the contact

pressure and failure pressure at that nut tightness. Recommendations regarding the results for use in litigation

has been summarised.

Keywords: Forensic, compression fittings, failure, litigation, water utilities, copper pipes, splicing, leakage

rate, leakage model, interference fit, mechanical face seal, static seal

1. Introduction

1.1 Forensic Engineering

Forensic engineering and the understanding of forensic evidence is an important sector of

the engineering discipline as it provides clarification of the cause of failure (failure

analysis). This in turn allows insight into reducing the risk of reoccurrence, damage to

property or environment, personal injury (terminal and non-terminal), loss prevention, and

can provide suggestions for better design. Failure may arise from inadequate materials and

equipment not performing to specification, design, operation and misuse, lack of adequate

maintenance, and durability (wear and tear). The failed product or system may have

exceeded service life, it may not be fit for purpose, or may be of bad quality.

It is also valuable in deciding responsibility of the failure and can underpin negligence in

the particular roles of a company, management, and cost cutting. It provides the evidence


2

to support liability in litigation or arbitration. It is important to consider the location of the

system and the risks associated with the location, particularly when assessing damage.

Table 8.1 [1] highlights typical steps taken in forensic investigations, and is used in this

report (although not in chronological order).

1.2 Compression Fittings

Compression fittings (or couplings) connect similar or dissimilar piping used for fluid

transportation, particularly water and natural gas supplies [2][3], and are typically

manufactured from brass. High pressure couplings are typically manufactured from mild

carbon steel or stainless steel for use in the hydraulics, oil and gas, chemical, and nuclear

industries [4]. They provide a suitable sealing interface minimizing the leakage rate whilst

maintaining fluid viscosity, both minimising flow resistance and maintaining the average

laminar/turbulent fluid flow. Due to the seal mechanism it is important to ensure material

compatibility at the interface between the ferrule and joining pipe. The size of the

compression coupling is determined by the diameter of pipework. Domestic and industrial

pipe sizes for water utilities vary, and compression fitting applications provide a range

from 6mm to 108mm [5] and an anchor bracket may be needed for large pipes. A typical

compression fitting is shown in Fig. 1.1. The two types of compression fitting commonly

used are Type-A for above ground and general use, and Type-B for underground systems.

This report focuses on the litigation of building damage through the escapes of water,

therefore Type-A will be the subject of the investigation. It is important to understand the

failure of this type of coupling due to their abundance in domestic and industrial sectors.

The failure mechanism of the coupling is an important aspect of litigation, and a means of

providing proof of responsibility.


3

Figure 1.1. A typical type-A compression fitting for domestic use.

1.3 Aims and Objectives

The aim of this report is to computationally and experimentally investigate some of the

most frequent causes of failure of Type-A compression fittings, evaluate the conditions

which lead to the failure mode by under tightening of the nut, and produce a summary and

recommendations directly usable by forensic engineers investigating compression fitting

failures leading to leakages. In particular, the specific objectives of this project are,

a. Create a FE model of a compression fitting of common use within the water

utility industry.

b. Set-up an experimental test of the compression fitting under typical service

conditions.

c. Validate the numerical model for a suitable number of experimental tests.

d. Couple the FE simulation to the experimental results to suggest the minimum

sealing requirement.


4

e. Use the results of the experiment and simulation to produce guidelines

regarding the evidence to be traced by forensic engineers investigating

compression fitting failures and how to link such evidence to the causes of

failure.

1.4 Methodology

1.4.1 Experimental Testing

An experimental arrangement has been conducted in accordance with the standards related

to copper piping and copper alloy plumbing fittings. The tests discern the pressure and nut

tightness that may lead to leakage. Evidence was also acquired relating to the cause of

failure and is an important contribution to the overall aims of the project, and vital for

forensic investigation. Observations were taken in all instances to suggest any available

trace evidence, and used to analyse in detail the failure mechanism.

1.4.2 Computational Modelling

A macroscopic simulation was created to model the sealing and failure mechanisms for the

compression fitting. It is a finite element portrayal of the plastic deformations of the

ferrule, pipe, and the corresponding large scale deformation.

1.4.3 Analytical Engineering

The underlying principles of continuum deformation and microscopic evaluation of

leakage and contact mechanics was discussed and used to evaluate some of the

observations made throughout the project.


5

B&Q PlumbSure (IBP) 8mm

straight coupler. £3.99

B&Q PlumbSure (IBP) 15mm

straight coupler. £1.45

Norgren 18 and 36 series. Used for

any fluid (incl. compressed air).

2. Technical and Literature Review

2.1 Technical Review

2.1.1 Coupling Requirements

The majority of compression fittings in the domestic and light industry use are

manufactured by either Conex [6], a subsidiary of International Building Product (IBP);

Comap [7], a subsidiary of Group Alberties; and Kuterlite [8], a subsidiary of Pegler

Yorkshire. The pertaining qualities of the designs are the geometry of the ferrule and the

interfaces between the ferrule and the fitting body and nut. Examples of variations in

fitting design are shown in Fig 2.1.

Figure 2.1. Designs and working environment for compression fittings.


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7

Figure 2.2. Type-A compression coupling.

Figure 2.3. Type-B compression fitting, requiring a flared pipe end [12].

2.1.3 Applications

Compression fittings are used within the area of fluid transportation. This typically

includes water and gas utility services, the oil and gas sector, and may extend into the

chemical industry. However, due to their appropriate use in potable water pipe-work [13]

they may be found in numerous areas. Within the water utilities industry (both domestic

and industrial) the fittings are used in replacement of conventional joining methods like

soldering. These can be more convenient in hard to access locations where soldering is not

suitable. They can be manufactured to a number of configurations including directional

changes (commonly named elbows), multi-port (or splicing) to connect or add to a pipe

system, or a union with no change in direction.


8

However, due to the increasing costs of copper the water industry is more frequently using

cheaper materials for new water piping installations like plastic (PVC, PP, PE), as they

offer comparable corrosion resistance.

2.1.4 Materials

The material parameters can only be suggested as the exact properties of the fitting have

not been disclosed by the manufacturers. Duplex brass is typical, and used for Conex

compression fittings [14], as well as DZR (dezincification resistant) for soft water as

dezincification may occur. Fittings which are DZR have the CR conformance mark.

Various names and chemical compositions are found to hold the name Duplex brass,

including alpha-beta brass, Muntz metal and 60/40 brass. The approximate composition is

60% Cu and 40% Zn. Specific compositions are shown below. As experimental tests were

not conducted in order to determine the material properties careful selection from

published sources were used. Equally, as the properties required are varied and not

typically stated in material databases multiple sources were used.

C35330 – DZR brass (59.5-64% Cu, 15-35% Zn, 0.2-0.25% As)

C36000 – Free cutting brass (60-63% Cu, 25-35% Zn, 0.35% Fe, 0.2-0.25% As)

C37000 – free cutting Muntz metal (59-62% Cu, 8-15% Zn, 0.15% Fe)

It seems adequate to suggest 60/40 brass is appropriate for the fitting, with C35330,

C36000, C37000 all having approximately 60/40 chemical combinations. Experimental

data from F. Stachowicz [15] found that copper pipe 16mm in diameter and 0.95 mm wall-

thickness had material properties as described in Table 2.1. This conforms with the „half

hard‟ designation of copper pipe - R250, where the value represents the UTS of the pipe.

The elastic modulus was not stated, and was obtained from experimental data of N.W.

Murray‟s work on thin-walled pipes being bent in the plastic range [16]. The


9

experimentally tested specimens cited were comparable to the pipe used in this report, and

the hardening exponent is corroborated by D.W.A Rees [17], who suggests that n = 1/3 for

many alloys.

Analytical and FE fracture modelling of a compression fitting is complex and not

necessary in this pilot study as the lower region of plasticity (and elasticity) is the subject

of the forensic investigation. The value of deformation at fracture was only needed to

ensure that the models (analytical and FE) do not suggest stress and strain rates beyond

fracture. The generalized plastic strain at fracture quoted for brass is from an FE simulation

of the tensile test conducted, with the data obtained at a centre point in the specimen [18].

It is assumed that brass is a perfectly isotropic material and compressive deformation, as is

the case for compression fittings, will follow the same strain path.

Coefficients of friction were not readily available for the contact variations (brass-brass

self-contact and brass-copper), so values were taken from [19] for brass-brass, μcz-cz = 0.3,

using a value from self-contact, under dry air at 30oC and under fretting cycle N = 0. For

copper the coefficient varies greatly depending on surface contaminants. An average value

for brass-copper of μCu-cz = 0.3 was used [20].

Table 1.1. Mechanical properties of compression fittings used in water utilities

Property Symbol Brass Copper Ref.

Elastic modulus E (GPa) 115 92 [21][16]

Yield stress Y (MPa) 140 65 [21][15]

Tensile strength UTS (MPa) 651 240 [18][15]

Poisson‟s ratio ν 0.33 0.34 [22]

Hardening exponent n 0.3 0.32 [21][15]

Strength coefficient K (MPa) 340 390 [21][15]

Strain at fracture εf 0.435 0.4 [18]

Coefficient of

friction μ

0.3

(brass-brass)

0.3

(copper-brass) [19][20]


10

2.1.5 Failure mechanisms

In tightening of the nut the ferrule “bites” onto the pipe, by causing a homogeneous

indentation radially. It is expected that if the nut is not tightened adequately upon

installation a resultant overpressure, or possibly nominal pressure, may cause the pipe to

work axially along the fitting until it by-passes the ferrule, at which point leakage will

occur. The range of possible causes of failure specific to compression couplings is

extensive and can include,

- Incorrect fitting – ferrule placement, under-torquing the nut, over tightening

- Excessive pressure or water hammer

- Pressure transient and/or vibrations

- Thermal cycling (hot or cold)

- Material defect

- Inadequate design

- Service life exceeded or bad condition of fitting

- Post bending of the pipe. After the coupling has been installed the pipe may be bent

to fit it into place; this will produce lateral (radial) loading.

- Incorrect pipe preparation – the pipe must be cut square to sit in the fitting body.

- The use of pipe jointing compound. Note: The UK copper board state “Jointing

compound is not normally necessary or recommended on compression ends,

however a thin smear can be used on the jointing surfaces (not the threads) if a

slight weep occurs.” [23]. The compound may inhibit the ferrule from deforming,

and after solidification and material contraction from the polycrystallisation may

lead to leakage. Also Fernox, a jointing compound manufacturer, state their LS-X

product is “ideal for use on compression fittings” [24], and does not solidify.


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12

Although both materials are of copper alloy the compression fitting will typically have

slightly different material characteristics (see Table 2.1).

The expressions for the strain-displacement kinematic relationship in cylindrical (polar)

coordinates are [25]

In axisymmetric situations, as is the case for the compression fitting, the deformation

produced is symmetrical about the axis, and the general equations of 3.1a reduce to

εr = ∂u

∂r, εθ =

u

r, εz =

∂w

∂z, γrz =

∂u

∂z+

∂w

∂r (2.1b)

and in tensor notation

𝛆 =1

2

2εx γxy γxz

γyx 2εy γyz

γzx γzy 2εz

=1

2

2εr γrθ γrz

γθr 2εθ γθz

γzr γzθ 2εz

=1

2 2εr 0 γrz

0 2εθ 0γzr 0 2εz

(2.2)

The Cauchy stress tensor is equally reduced due to axisymmetry. A cylinder subjected to

internal and external pressure as shown in Fig. 2.4 will produce no shear stress due to

symmetry [26], and so the stress tensor, in cylindrical form, becomes

𝛔 =

ζx τxy τxz

τyx ζy τyz

τzx τzy ζz

=

ζr τrθ τrz

τθr ζθ τθz

τzr τzθ ζz

= ζr 0 τrz

0 ζθ 0τzr 0 ζz

(2.3)

The radial and tangential (θ) stresses in terms of the strain deformations produced in a

cylinder can primarily be described from S. Timoshenko [27]. At the geometrical limits of

(2.1a)


13

the cylinder the components of the stress tensor were first described by G. Lamé [28] and

known as the Lamé equations. They are,

ζr =ri

2pi − ro2po

ro2 − ri

2 −(pi − po)ri

2ro2

r2 ro2 − ri

2

ζθ =ri

2pi − ro2po

ro2 − ri

2 +(pi − po)ri

2ro2

r2 ro2 − ri

2

(2.4)

Figure 2.4. A cylinder submitted to internal and external pressure.

Where a – inner radius, b – outer radius, r – radius under inspection, pi or po – internal or

external pressure respectively. For external or internal pressures only Eq. (2.5) reduces by

setting either, pi = 0 or po = 0 respectively. For compound cylinders (where an external

cylinder is shrunk onto an internal cylinder by thermally increasing its diameter) the

reduced equations combine to describe the stresses produced by a shrinkage fit or

interference fit. This type of problem is shown in Fig. 2.5.

pi ri

ro

po


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15

Figure 2.5. A uniform band of pressure.

Solutions to the governing equations stated in Eq. (2.1) for hollow cylinders as a load

carrying structure are modelled by either thin shell theory [29] or elastic theory [30][31].

Comparisons on the applicability of each have been suggested in the referenced material

[32][33]. Hollow cylinders and bands of pressure are common in many structural

environments, including vessels for the nuclear, petrochemical, aerospace and marine

industries [34] (for example submersibles [35]). The governing equations for a uniform

band of pressure have been solved for an infinite and finite cylinder, with variations in the

model according to the required condition that the moments are zero at the free ends (for a

finite length) [31]. An infinite length is defined by St. Venant‟s principle, which indicates

that the stresses decay to zero at a distance two to three times the pipe outer radius [36].

Whilst the compression fitting connects to the cut end of the pipe at a distance less than

this, the inner surface is restrained by the enclosing fitting - hence it is not a free end. For

thin shell theory they involve the stress function of A.E.H. Love [37]. G.P Steven [38]

suggests a solution for the interference pressure fit of a collar on a hollow shaft as

pint = δ

ro,col2 + 1

ro,col2 − 1

+ νcol

2Gcol 1 + νcol +

ri,s2 + 1

1 − ri,s2 + νs

2Gs 1 + νs

(2.6)

Where in this instance the radius at the interface between the pipe and collar is named ri,col

and ro,s, and as with the compound cylinders δ is the lack of initial fit. There are variations


16

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17

ςCOLL =Pr

t+

Et

r 3(1 − ν2) (2.7)

and comparisons can be made with thin cylinder theory. This is only a single variation of

many forms for the collapse load. Cylindrical shells under axial compression have been

extensively researched as part of Thin Walled Structures, and it would not be appropriate

to suggest any one solution over the other without further analysis. In particular it is not

known whether Eq. (2.7) holds for comparatively short cylinders. It is possible that

collapse may occur on the ferrule; there are restrictive boundary conditions imposed on

approximately 1/3 the length of the ring but the rest is free to deform and potentially

collapse. The alterations of Eq. (2.7) and its variations when part of the axial structure is

restricted radially would be interesting to study in itself.

2.2.2 Microscopic Analysis

The aforementioned continuum model describes the mechanism of compressively

deforming the ferrule onto the pipe. The sealing mechanism however occurs at a

microscopic level, as a function of the contact between the interfering parts. If two bodies

are in contact as Fig. 2.4a and a force, F, is exchanged normal to the tangent of the origin,

O, so that two points, m and n, touch (as per Fig. 2.4b), then the distance m and n have

travelled vertically, called w1 and w2, can be described by [46],

w1 + w2 = 1 − ν1

2

πE1+

1 − ν22

πE2

qcdA

r (2.8)

The applied force, F, is described above as the intensity of the uniformly distributed load

across the contacting surfaces, qc. Other subscripts denote the material properties of the

two bodies, and dA is an infinitesimal area of the contact surface.


18

a) Initial contact of two bodies b) Local deformation of the two bodies

Figure 2.6. Two contacting bodies with point‟s m and n touching upon the exchange of a

force.

Eq. (2.8) was initially solved by H. Hertz [47] as,

qc =3

2

F

πra rb (2.9)

Where, ra and rb are the major and minor radii of the semi-ellipsoid; and the maximum

contact pressure, qc, acts at the centre of the contact surface. Consequently Hertz‟s

formulations based on Eq. (2.9) became known as Hertzian contact mechanics, and can be

applied to individual asperities. Modelling of the individual asperities in this way is

unreasonable and requires exacting geometry. In 1966 J. Greenwood and J. Williamson

[48] described a new theory of elastic contact and built an instrument to measure the

surface topography. This was used to obtain an asperity population with a statistical height

distribution. The radius of curvature described in Eq. (2.8) was defined as an average

value. Fig. 2.5 shows an extract of the asperity population.

Figure 2.7. Example asperity population [48].

The fluid flow that constitutes the leakage rate can be initially derived from the Navier-

Stokes equations which govern the motion of a fluid. For the axisymmetric situation [49],


19

For steady flow,

∂ur

∂t=

∂uz

∂t= 0 (2.11)

and Eq. (2.10) is reduced accordingly. Following a parabolic velocity profile, an exact

solution is obtained by the Poisseuille Flow Law [50],

uz =r2

8μ

dp

L (2.12)

Where uz is described by Eq. (2.10), the flow in the axial direction, and μ is the fluid

viscosity. Whilst the flow is not axisymmetric due to the asperity contact varying not only

radially but also tangentially (hence θ ≠ 0) it is assumed that the average gap, and therefore

flow rate, is constant. As such, the situation in one cylindrical half line is not

distinguishable from any other.

There is a large body of research conducted on mechanical interface seals (see B.

Tournerie and J. Frene‟s summary on the principal research areas [51], and A.O. Lebeck‟s

book on mechanical face seals [52]). The majority of work pertaining to the leakage of

mechanical face seals revolves around rotating or reciprocating shafts, due to their

abundance in machines and engines. In these cases the required solution is based on a thin

fluid film, which provides a fully lubricated seal. This fluid film is paramount in deriving

(2.10)


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21

adjusting A.O. Lebeck‟s version of the simplified hydrostatic model [54][55]. Eq. (2.12)

can be coupled to Eq. (2.14), where uz = Q and is adjusted for the radial direction. It is

rewritten as a function of the pressure and radius only, so that

Q = −2πrg3

12μ

dp

dr (2.16)

and the total flow is constant across any radius.

The integration of Equ, 2.14 is performed separately, and the notation is the same as the

Lamé equations of 3.5 except that the integration is not over the entire radius, but that of

the contacting radius,

Q 1

rg3

rc

ric

dr =−2π

μ . dp

p

pi

(2.17)

A primitive solution is

p − pi =Qμ

−2πg3ln

rc

ric (2.18)

which can provide the pressure variation across the seal face for a given leakage.

Alternatively, the leakage rate can be found by using the extremities of the seal geometry

and the pressure inside and outside of the seal. Rearranging Eq. (2.18) for these parameters

gives,

Q =−∆p2πg3

μ lnro

c

ric

=ri

c

roc exp

−∆p2πg3

μ (2.19)

Of note is that if the internal pressure is greater than external, pi > po (i.e. it is not a

vacuum), then the integration is across the negative radial direction, and

∆p = po − pi (2.20)


22

So that Eq. (2.20) is negative. If pi < po, in the case of a vacuum, the leakage would move

into the pipe.

The above descriptions apply as an approximation for a certain amount of leakage (based

on assumptions stated by A.O. Lebeck [54]). However, at particular liquid pressures there

is no leakage. An expression must be found for the internal and external pressures as a

function of hydrostatic equilibrium. The body of water may follow two paths, assuming

that the thread contact pressure is greater than that of the ferrule, and is shown in Fig. 2.6.

Resolving forces for axial equilibrium,

pir. drro

ri

− por. drro

ri

= 0 (2.21)

Where the radii varies as Fig. 2.6, and is described as

ro − ri = g (2.22)

So that again the average asperity gap is presumed.

a) A potential flow path under the ferrule

b) A potential flow path over the ferrule

Figure 2.8. Possible fluid flow paths.

ri ro

ri ro

po

po

pi

pi


23

3. Methods

3.1 Experimental Testing

3.1.1 Apparatus

1. Dead Weight pressure gauge tester

2. Type-B fitting adaptor

3. Capillary copper pipe, 500mm

4. Capillary end adaptor

5. Ball valve

6. Support blocks

7. Clear polycarbonate protection screen

3.1.2 Experimental Procedure

An experimental arrangement was conducted in accordance with BS EN1254-2:1998 [56],

under hydrostatic pressure – Annex A of EN1254-2. Of importance is the plastic

deformation of the ferrule and the pipe, hence these were replaced after each test to ensure

experimental consistency. The couple was disassembled, and, referring to Fig. 2.2, parts 1

(fitting body) and 2 (nut) were inspected for damage. Part 3 (ferrule) was replaced, and

part 4 (pipe) was cut down by 17mm to remove the compressive deformation. The cut-off

was recorded according to its failure pressure, and stored for closer inspection. The pipe

length exceeded 100mm in length, and the setup is shown in Fig. 3.1a, with both

requirements corresponding to BS EN1254-2:1998 [57]. The actual experiment is shown in

Fig. 3.1b and, specifically, the procedure was as follows,


24

1) The capillary end adaptor is brazed onto the capillary pipe, and adapts the capillary

pipe to the 15mm adapter.

2) The capillary pipe is bent to adequate geometry to allow the pipe to rest on the

support blocks horizontally. The dead weight pressure gauge tester uses oil as the

medium, but the test uses water under hydrostatic pressure. Therefore, there must

be an interface between water and oil, which occurs at the end of the capillary

adaptor (see step 4).

3) Item 2 is attached to the pump.

4) Oil is pushed to the capillary end via the manual wheel on the pump.

5) The ball valve and pipe assembly is attached. The ball valve is lightly supported,

but it should be ensured that axial movement of the pipe is NOT restricted.

6) One end of the straight coupling is attached to the ball valve end according to the

manufacturers recommendations (1 1/4 turns for a 15mm coupling).

7) The other side of the sample is attached to the capillary pipe connection.

8) The sample is tightened according to Table 8.3. The first sections of Table 8.3 are

completed.

9) Water is added via the ball valve until it is full.

10) The ball valve is shut.

11) Item 7 is attached to protect the operator from any sudden bursts.

12) The pressure is gradually increased. The sample is observed for any leaks

13) Once a leak occurs pressure is reduced to ambient by turning the wheel until the

10psi level is unloaded. Note: do not reduce p beyond ambient as this may

contaminate the pump with water. If this occurs the pump needs to be flushed.

14) The specimen is unscrewed on the test side, water drained, and sample examined.

Data is collected according to Table 8.3.


25

15) The fitting body and nut is inspected on the thread, pipe abutment (for fitting body),

and interface with ferrule. The ferrule and pipe compression end is inspected for

any obvious damage and evidence.

16) The pipe and ferrule end is cut off ensuring a square cut for a new specimen and

stored according to failure pressure and tightness. The new pipe end is de-burred.

17) The ferrule is replaced. The nut and fitting body are replaced if damaged.

18) Steps 7 through 17 are repeated.

Note: all precautions were taken in case of an unwanted failure. However, the risk of injury

was small when compared to using a compressible fluid (gas) for the experiment, as the

volumetric elastic energy is comparatively insignificant. Also, Type-B connectors would

have been tested for a gas system.

The fittings were tested to 20 bar, as this is the manufactured rating of the ball valve. Upon

this pressure being reached, all connections were inspected for leaks during the test. The

test was then repeated from the other end of the tightness variable (i.e. over-tightening),

converging onto the manufacturers recommended tightness. This was repeated several

times for experimental reliability. Once this was completed the higher pressure failure tests

were conducted, with inspection of all joint connections at each pressure level.

The conversion from the number of turns to axial displacement of the nut was found by the

pitch of the straight coupling thread, and measured at 1.5mm pitch. This is a standard pitch

of major thread diameter of an M20, but is coarse and the measurement may not have been

accurate. As multiple 15mm fittings were used to construct the experiment the distance

was measured directly at various thread tightness, and is shown along with the results of

the tests in Section 9.


26

a) Apparatus arrangement

b) Actual setup (without protective screen)

Figure 3.1. Experimental setup for hydrostatic testing.

Of note is the ball valve height. The hydrostatic pressure on the specimen is more than the

pressure at the ball valve due to the force of gravity. Reducing the First Law of

Thermodynamics where there is no fluid flow, V = 0,

p h = ρ h g h h1

h0

dh (3.1a)

Adapter from pump

to copper pipe

Pressure gauge

Specimen End block

≥100mm ≥100mm

Capillary end

(adaptor)

Dead weight pressure

gauge tester Specimen Ball valve


27

As there is no variation of gravity (the change in height is small) and the fluid is

incompressible (water), the static pressure head becomes,

p = ρg∆h = 784.8 Pa (3.1b)

Which is Pascal‟s Law. Values of the variables are ρ = 1000 kg/m3, g = 9.81 m/s

2, and the

change in height as measured from the top of the valve to the bottom of the specimen ∆h =

80mm. This pressure change is only 0.772% different from that of atmospheric, patm =

0.101325 MPa. The same variation applies to the additional pressure on the specimen due

to the weight of the liquid in the capillary pipe (from the pressure tester). They are both

considered negligible additions to the measured pressure.

3.2 Computational Modelling

3.2.1 Continuum Modelling

The continuum computational model will be a static stress/displacement non-linear

analysis, to include plasticity and large strain evolution. Because of plasticity and large

strains geometric non-linearity has been accounted for in the computational model (so

named “Nlgeom” in Abaqus). For this parameter the state of the model at the end of a step

provides the initial state for the start of the next step [58]. Therefore, because the current

step geometry is used it follows a true strain evolution. Also, the kinematic equations to

compute the strains are different. Because strains in the ferrule are large this will have a

great effect.

The model correlates the amount of plasticity of the ferrule and pipe with an experimental

leakage. From Eqs. (2.8) and (2.9) (the Hertzian solution) the contact pressure qc is the

principal effect, and from the preceding discussions on leakage it is obvious that this

contact pressure plays the pivotal role in maintaining an adequate seal. Therefore, the FE

simulation uses CPRESS, the contact pressure, to correlate the sealing mechanism at


28

varying tightness. Varying the horizontal translations of the nut according to the

experimental variable (the nut tightness) the maximum asperity contact can be suggested to

minimize leakage. Also, the FE simulation can be verified by measuring the radial and

axial displacement in Abaqus and compared with the experimental deformation.

3.2.2 Materials - Introducing Strain Hardening

Because there was large plasticity in the ferrule a sensitivity analysis has been conducted

on strain hardening of brass to suggest its importance on the Abaqus model. A uniaxial test

model has been created in Abaqus, and a comparison has been made between an elastic-

perfectly plastic model and a strain hardened model. By comparing the simulation of a

uniaxial test with published experimental results it also confirmed whether Abaqus

accurately models the strain hardening phenomena for the configured parameters.

As the stress produced in the model depends on the strain increment only (for example it is

not time or temperature rate dependent) the strain hardening exponent model (power law)

suggested by J.H. Holloman [59] has been used as the constitutive description [60],

𝛔 = K𝛆 pn (3.2)

The strain hardening exponent can be calculated empirically. For example by use of a

double compression test of the material under consideration. One of the specimens to be

simultaneously compressed is in the annealed condition and the other has a determined

amount of strain already present (pre-strain). Then by using,

n = 𝛆 a − 𝛆 p

ln( εo − 𝛆 p)

𝛆 a

(3.3)

Where the subscripts denote annealed (a) and pre-strained (p) conditions. See R.

Ebrahimi‟s treatise on the experimental setup [61]. From Table 2.1 the strain at fracture is


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30

𝛆 =

ε1

ε2

∆L

L000

=

ε1

ε2x(e)

L000

=

00

x(e) mm

100mm000

(3.8)

with the compliance tensor for isotropic materials

𝐒 =1

E

1 −ν −ν 0 0 0−ν 1 −ν 0 0 0−ν −ν 1 0 0 00 0 0 2(1 + ν) 0 00 0 0 0 2(1 + ν) 00 0 0 0 0 2(1 + ν)

(3.9)

and the constitutive relation

𝛆 = 𝐒 𝛔 (3.10)

Eq. (3.10) reduces accordingly for a uniaxial test, and upon rearranging and using brass

60/40 produces a displacement of

x(e) = ∆L =ςe

EL =

100

115000× 100 = 0.087mm (3.11)

This situation was simulated in Abaqus and shown in Fig. 3.2. The model parameters were

- Elements: Hex 3D structured, linear, full integration. 1 element.

- No geometric nonlinearity (it is elastic).

- 100 x 20 x 20mm specimen geometry.

The results are the same as Eq. (3.15), as it is an exact solution (hence the use of a single

element). The model also follows St. Venant‟s principle due to the homogeneity of the

stress distribution (i.e. the model has a uniform stress state throughout, and no localised

concentrations at the boundary conditions).


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32

Figure 3.2. True stress-true strain curves for various elastic-perfectly plastic (EPP) and

strain hardened models.

3.2.3 Elements - Axisymmetric Model of the Specimen

Geometric modelling was constructed from Drawing D00001 (see Appendix), with half of

a straight coupling being modelled. In this way both the axial line of symmetry (along the

pipe axis) and symmetry on the plane normal to this are used. Because an axisymmetric

model was employed it was not possible to model the thread as the independent variable.

The major and minor diameter of the thread on both the nut and the body were measured

and used to create a solid interface as is seen in the geometry of Fig. 3.5 or Fig. 3.6.

The computational model follows similar parameters to the uniaxial test to ensure accurate

material and geometric simulation. Geometric non-linearity was applied due to the large

strain evolution. As with the simulation of Section 3.3.2 and of Table 8.2, second-order

(quadratic) elements were used. These elements avoid the potential for shear locking, as

first order elements cannot produce curved shapes [63][64]. This leads to shearing stresses

for the case of pure bending, and is inaccurate. Bending stresses in the model was a likely

situation because of the moments induced in the ferrule. The element types were more

0

50

100

150

200

250

300

350

400

450

0.00 0.10 0.20 0.30 0.40

ζ (

MP

a)

ε

Plasticity Modelling

Strain hardened model (brass)

EPP model (brass)

Strain hardened model (copper)

EPP model (copper)

Abaqus FE simulation (brass-

hardened)Abaqus FE simulation (brass-EPP)

Abaqus FE simulation (copper-

hardened)Abaqus FE simulation (copper-EPP)


33

complex, owing to the complex geometry of the model, and are shown in Fig. 3.4. Whilst

second-order reduced integration may be useable because of the mesh quality (i.e. a fine

mesh), it provided no great reduction in computational processing time. This was because

of the relatively small geometry. That is, the areas of interest in the model were not

disproportionately small compared with the rest of the model. If the ferrule were, say, 10

times smaller than the fitting body, re-evaluation of the element types would have been

necessary. Therefore, full integration was used, and also eliminated the potential for zero-

energy modes („hour-glassing‟).

Figure 3.3. Finite element mesh, with parameters for various partitions.

3.2.4 Model Parameters

The independent variable that defines the sealing mechanism is the torque applied to the

nut, which places a linear transformation onto the ferrule, compressing it. Hence, the

mechanical interactions between the four parts are paramount to an accurate simulation.

Within Abaqus “the mere physical proximity of two surfaces in an assembly is not enough

to indicate any type of interaction between the surfaces” [65] and so interactions needed to

be setup for the simulation, shown in Fig. 3.6. Specifically, the they were 1) between the

inner chamfer of the body and inner chamfer of the nut with the outer part of the ferrule, 2)

the threaded part of the nut and the body, 3) the outer surface of the pipe with the inner

Element Shape: quad, structured

Geometric Order: quad

Integration: full

Element Shape: tri, structured

Geometric Order: quad

Integration: full


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35

the nut is imposed on the model with the linear transformation by relating the nut thread

(horizontal) edge to the body thread (horizontal) edge. However, this situation creates

inconsistencies when relating two nodal degrees of freedom (DOF) [69]. The DOF of each

node set (on the body and nut parts) is independent of each other as they are separate parts.

The Abaqus Manual suggests a „linear constraint equation‟ is used to resolve the

inconsistency. This was set up by creating three reference nodes: a node for the thread on

the nut and body, and a dummy node for the prescribed displacement. The reference nodes

were constrained to their respective thread edge in the usual manner, along the axial

direction. The reference nodes were then constrained to each other by using the linear

constraint equation. The constraint equation must be of the form that the combination of

nodal variables is equal to zero, such that

A1uiP + A2uj

Q + ⋯ + AN ukR = 0 (3.13)

Where uiP is a nodal variable at node P, DOF i, and AN is a coefficient that defines the

relative motion [70]. For the tightening of the nut relative to the body the displacement

equation can be found from analysis of Fig. 3.5.

Figure 3.4. Reference nodes and their constraints to the thread edges.

z2 − z1 = z3 (3.14a)

Standardising Eq. (3.14a) to zero and re-writing in the form of Eq. (3.13),

z1 z2 z3


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38

Figure 3.5. Boundary conditions imposed on the model

Figure 3.6. The initial model of the von-Mises stress results, and highlighted areas of

interest.

Figure 3.7. Element inspection points for mesh convergence check.


39

Figure 3.8. Mesh convergence results, with no. of nodes used in the simulation highlighted

in red.

0

2

4

6

8

10

12

0

50

100

150

200

250

300

350

400

0 100 200 300 400 500 600 700

ζvM

%w

arnin

g

(MP

a)

ζvM

(MP

a)

No. of nodes

body

ferrule

pipe

nut

ζvM % warning (MPa)


40

4. Results and Discussion

Three systematic scientific approaches were conducted and have been presented in this

section, with corroborating and correlating analysis.

4.1 Experimental Results and Discussion

As the empirical evidence is considered the most valuable of the investigation it is

presented first. The tests were conducted multiple times to obtain a statistical significance

and Fig. 4.7 illustrates the statistical mean pressure to cause leakage at variations of nut

tightness. It can be clearly seen that an exponential growth is developing from the

experimental results. Although it was anticipated that a normal or skewed normal

distribution would arise with a peak at 1 1/4 turns, it was not expected that the required

failure pressures would be so high for such comparatively low contact pressures between

the component parts. As such the experimental methodology of accumulating data sets at

low pressures had to be rejected. Some of the components of the experimental rig had

ratings up to 2MPa, so the nut tightness was incrementally increased up to this point, and

then over tightening of the nut was to be tested. The designated pressure limit of 2MPa was

apt considering working pressures of in house water mains is typically between

0.1~0.5MPa, and by restricting to this method the rig integrity would be maintained.† It

quickly became apparent that this would not suffice for many nut tightness and indeed 50%

of the results were above this pressure. Furthermore, the compression fittings tested would

not allow a tightness beyond one and a half turns, and failure pressures from nut over-

tightening was therefore not tested. The design of the 15mm fitting tested was such that

significant over-tightening was impossible (only 1/4 turns past the recommended tightness

† Of note is that since the privatisation of the water industry in 1989 there has been no governing body to

enforce minimum and maximum working pressures [73], so exact pressure extrema could not be quoted.


41


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43

Figure 4.1. Low leakage rate initiating at the top of the pipe from measurement 1.

Figure 4.2. Measurement 10. The specimen ferrule popped of its seat, but maintained a

seal. Note visible ring of deformation creating the seat.

a) Light score marks of measurement 16 at 1/4 turns and b) More visible score marks of

measurement 17 at 1/3 turns

Figure 4.3. Pipe separation causing score marks as the ferrule moves axially along the

pipe.

a b


44

Figure 4.4. Incorrect ferrule positioning from measurement 11, causing a below average

failure pressure.

Figure 4.5. Cut section of pipe at ferrule interaction, named as „deformation seat‟.

Fig. 4.1 shows the failure leakage of a low flow rate at hand tight, initiating at the top of

the pipe. This is due to the specimen both deflecting the pipe by the fittings own weight

and also supporting the pipe at the top of the nut and the bottom of the edge of the pipe.

This thereby created a gap between the pipe and the internal fitting components, allowing

leakage.

Fig. 4.2 and Fig. 4.3a represent the same no. of turns, and the visible score marks are

consistent with each other. However, there is a discrepancy of 86% between the failure

pressures. The mode of failure is also not consistent between the two as test 10 did not

show that the pipe immediately separated from the specimen, even though it failed at a

higher pressure than test 16. Comparing with the mean failure pressure of 3.2MPa for 1/4

turns it is possible that either,

Axial tilt of the ferrule


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46

observed in tests 2, 8, 14, and 18 where by the tightness was 1/12 of a turn, the minimum

amount tested.

It was noted that whilst conducting test 21 the pressure was raised comparatively slowly.

The pipe loudly „popped‟ three times, where-by the pipe worked axially along the

specimen, but did not separate completely until the third pop.

It was also noted that after each test when the pipe section was disassembled, occasionally

when the water was drained in the receptacle there was oil visible, which came from the

pressure tester section (i.e. the pump). It became obvious that the oil may contaminate the

specimen and the pipe section that inserted into the specimen. This severely spoiled initial

results as it would act as a lubricant to assist separation of the pipe from the specimen.

These results were disregarded. Whilst all attempts were made to remove the oil with a

degreasing agent, and a new section of pipe was used (as dictated by the testing procedure)

it was not possible to eliminate the possibility that residual oil may have affected some test

results.

4.2 FE Simulation Results and Discussion

As has been described in Sections 2.2.2 – Microscopic Analysis and 3.2.1 – Continuum

Modelling, the contact pressure plays a fundamental role in adequately preventing a

leakage occurrence. Therefore the maximum contact pressure, qcmax

, (CPRESS) is

measured across the potential leakage paths of Fig. 2.6, and the results are shown in Fig.

4.7. Because an axisymmetric model has been utilised the contact pressure is shown as

contact lines in Abaqus (rather than surfaces). Specifically, referring to Fig. 4.6, the

minimum contact pressure is found across the possible leakage paths from the pseudocode

process shown in Algorithm 1,


47

Algorithm 1. Procedure for qcmin

if ∈1 < ∈2 and ∈3 < ∈2

qc = max {∈1, ∈3};

else if ∈2 < ∈1 and ∈4 < ∈1

qc = max {∈2, ∈4};

else if (∈1 < ∈2 and ∈3 > ∈2) or (∈2 < ∈1 and ∈4 > ∈1)

qc = min {max {∈1, ∈3}, max {∈2, ∈4}};

The second statement is in the event that the top (bottom) path fills up but the next contact

prohibits leakage so that the bottom (top) path then fills up and leakage occurs along this

path.

Figure 4.6. Mini-max contact pressure inquiry locations.

Figure 4.7. Experimental results of failure pressures and FE simulation of qcmax

.


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50

Figure 4.8. Coarse meshes producing inaccurate stress development deformation due to

the finite nature of FEM.

The model was verified to an approximate extend by measuring the deformation directly.

The Finite Element Method employed in Abaqus/Standard develops a solution for the

displacements firstly, and then provides the strains/stresses/etc. via kinematic and

constitutive relationships, which may introduce approximations. Hence comparing the

displacements provides a direct comparison.

The simulation was set up so that the hydrostatic pressure was impressed onto the internal

surface of the pipe after the horizontal translation had completed. The tested compression

fittings have a rating of 1.5 times the maximum operating pressure of the fitting [57], at 2.4

MPa. Completing a simulation at 2.4MPa and at 0.101MPa and employing Algorithm 1

Inaccurate stress

development Inaccurate

deformation


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52

As can be seen, this expression, along with the strength coefficient K, can expand a

mathematical model to include strain hardening, using just the frequently referred material

properties of yield (Y) and tensile stress (UTS). Finding the equivalent expression for the

system under inspection may pose its own problems though, and so the default Hardening

Power Law was imposed. Furthermore, the hardening model of the uniaxial test applied a

displacement as a function of the equivalent plastic strain form the initial compression

fitting model, which in turn may be affected by the sensitivity analysis on strain hardening

(and therefore the applied displacement), and so is a self-perpetuating cycle.

Finally, of note, the simulation assumes that the classic isotropic Coulomb friction model

is an acceptable model for the friction between the moving parts. Whilst this is true, as the

dry friction created is independent of the speed of the moving body (i.e. how quickly the

nut is tightened), and the area of contact, it is in fact irrelevant. The model applies

interactions at the interfaces to actually define the contact surfaces. If no interactions were

created the parts would behave independently of each other. The fact that the nut is

tightened to a specific amount of turns (or horizontal displacement) is independent of the

friction resisting the displacement, as this displacement must be enforced. However,

physically, the torque applied is a function of the friction, and in particular instances of

installing a compression fitting this may prohibit the required tightness for adequate

sealing. The torque was not measured specifically as this varies according to the size of the

fitting, and whether a lubricant was used on the thread.

4.3 Comparison between FE simulation and Experimental Results – Direct Method

The deformation of the ferrule at the manufacturers instructed tightness of 1 1/4 turns is

compared with the FE simulation, and shown in Fig. 4.9. The radial and axial discrepancy

is 0.2% and 3.2% respectively.


53

a) FE simulation measured node coordinates – (r, θ, z)

b) Experimental measured deformation (example)

Figure 4.9. Comparisons of the ferrule deformation.

node 622

original coord.: 8.47, 0, 4.57

deformed coord.: 8.60, 0, 5.68

node 103



node 6




54

4.4 Discussion on the Engineering Analysis

A multi-scale modelling approach was adopted. The analytical descriptions are used as a

means of portraying the sealage and later the leakage mechanism so that the fundamental

principles and relationships are expressed, rather than explicitly evolving an analytical

leakage model. As such they provide the understanding to relate the physical observations

and results with both the leakage mechanism and the ability to gather quantitative data

rather than qualitative data from the FE simulation. Most importantly from the

developments shown in Fig. 4.7 an exponential trend is empirically forming from the data

gathered. This is in accordance with the primitive leakage solution of Eq. (2.19).

Eqs. (2.1) through (2.11) describe a continuum representation of the sealing mechanism

through the classical theory from mechanics of materials. Whilst the solutions have been

well proven (see S. Timoshenko [46][27]) their application to compression fittings is

approximate. Particularly, it suggests that the end of the cut pipe is restricted due to the

enclosing fitting. However, this is a simplification. It is indeed restricted in the radial

direction, but not necessarily in the axial direction. Initially the end is free in this direction,

but as the ferrule compresses onto the pipe it inhibits further separation. This in turn may

invoke a greater force applied the other end of the fitting, which may or may not be free.

Equally, the continuum analysis is based on expressing solutions for the purpose of stress

distribution. But, through backward analysis and by use of the kinematic and constitutive

relations it would not be too involving to evaluate the governing equations as a function of

the displacement and forces (hence contact pressure and load intensity). The modelling

shown as Eqs. (2.8) through (2.22) is at micro-scale due to the asperity distribution. This is

important in understanding the leakage rates, and the surface interactions of the component

parts. However, data was not available on the surface roughness of the fittings, and


55

therefore could not be used to suggest the critical gap and its relation to the contact

pressure.

There is an expansive amount of research relating to mechanical face seals conducted in

the 70‟s and 80‟s owing to their use in machinery, within the area of tribology. The

analytics described in Section 2.2 express the main relationships between contact pressure,

asperity gap, and leakage rate; in particular Equ‟s 2.11 through 2.18. This provided the

necessary details to couple the FE simulation with the experimental results. However, the

derivations above are not accurate, as g(r), the asperity gap is a function the radius. A.O.

Lebeck [54] showed that in the case of tapered non-contacting or contacting surfaces

where g = h (the film thickness) and h = f(θ) where θ is the angle of the tapered fluid film.

In this case r = rm, the mean face radius, and Eq. (2.17) can be solved. This approach could

be taken for Eq. (2.17) through 2.19, by stating g = gm. However, the solutions of A.O.

Lebeck were based on the assumption that the ratio of the face width to the mean face

radius is < 0.1. This is not the case for the compression fittings as the contact area is large.

It may also be apparent that the model described is for a vertical seal, such that the contact

face is directly proportional to the radius, and does not take into account the internal taper

of the body and nut for the compression model. If Eq. (2.16) were adequately applied to

this case it would provide the leakage rate, Q = f(∆p,r).

Whilst Eqs. (2.10) through (2.20) derive a simple model as a function of the leakage rate

and pressure, it does not immediately explain the observed phenomena of the pipe

“popping” off the fitting by the ferrule being removed from its deformation seat, and are

described later.


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57

1) The measuring device (Vernier caliper) and location of the instrument onto the

ferrule. The measured radial and axial deformation varied from 17.3~17.18mm and

6.82mm~6.63mm respectively. The mean value was compared.

2) The elasticity of the material upon unloading (i.e. removing pipe from the nut

and body to access the ferrule).

3) Exacting geometry of the FE model.

The elastic behaviour can be accounted for by use of Table 2.1, where Y/E = εe =

1217μstrain. Converting the true strain into the engineering strain,

εtrue = ln 1 + εeng → εeng = eεtrue − 1 = 0.492 5.2

Comparing with εe shows a 0.25% elastic offset strain, and is negligible in comparison to

the discrepancy between the measured and simulated results.

There are issues with transmitting the required torque to some fitting bodies. In some

instances, it is not possible to achieve this and still conform to the manufacturers assembly

recommendations. For example by holding both threads of an elbow joint and torquing will

induce forces normal to the pipe, and can cause the pipe to drift from its correct position.

This is against the manufacturers assembly requirements. Examples of a lack of design to

allow torque transmission are shown in Fig. 4.10. This is detailed in BS EN1254-2:1998-

Section 4.4.5 [76], but is not enforced in the designs listed. There are also many examples

of combined fitting types, with one end being a compression design. The same design

recommendations are described for these types of fittings in EN1254-4:1998-Section 4.4.5

[77].


58

Figure 4.10. Compression fitting designs which do not conform to the Kitemark.

There are many variables to the Abaqus simulation that are defined but not definite. Most

are related to the material properties and exacting geometry. For example, the strain

hardening of brass and copper has been used and compared with published works, but the

material coefficients and constants are not necessarily exact to the materials used in the

manufacturing of the components. Even Duplex and DRZ brass, as specified by

compression fitting manufacturers, have a wide spectrum of properties, and specific data

was not obtainable from them. This was also the case for brass-brass and brass-copper

coefficients‟ of friction. More importantly, the metallurgical treatment of the brass and

copper pipe will have a great effect on the ability of the ferrule to compress onto the pipe

and form an adequate seal. For example, the soft copper pipe R220 (annealed) is applicable

for water utilities [78] and is used in circumstances where tight bends or complex pipe

routing is necessary. It is not exceptional to suggest that compression fittings would also be

used in such circumstances as they are designed to be used in inaccessible areas where the

typical method of soldering is unavailable. But, to use a compression fitting for R220

designation of pipe an internal sleeve must be used inside the pipe for the ferrule to

compress onto. If a fitting were installed by someone who did not realize the pipe was of

the R220 variety rather than the R250 they may be oblivious to the need for an internal

sleeve. This circumstance is likely, due to various installations and retrofitting after the

initial plumbing system has been fitted.


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61

where clearly there is a large amount of plasticity occurring in the fittings tested. There are

a large amount of plasticity models available, and adjusting the solution to include

plasticity would be possible.

4.6.3 FE Simulation

The computational model studies the sealing mechanism of a compression fitting. It does

not however model the method of leakage, and as such it follows that the next logical step

would be to couple the Abaqus model with that of a CFD model. It is noted however that

in the simulation many of the set parameters are to enforce relations on the respective

parts, for example the coupling constraint on the pipe and body parts. In failure it was

observed that frequently the pipe detached itself from the body. In fact, all leakage

mechanisms must start at this location, as it is the origin of the water-pipe-fitting interface.

Therefore, if a coupled analysis were conducted the constraint imposed would have to be

re-organised or removed in the proceeding analysis to allow leakage to occur.

From Section 2.2.2 it shown that the microscopic analysis is fundamental to the leakage

mechanism. Therefore a multiscale modelling approach using asperity contact for the

microscopic region would be valuable.

Algorithm 1 was used to define the potential leakage path and coupled to the contact

pressure. But, this is only an approximation as the contact pressure would change

according to the change in deformation due to the initial liquid penetration.

The FE simulation could be adjusted for varying tapers to suggest its influence. This is

however more suggestive of a re-design approach rather than a forensic investigation.


62

5. Conclusions and Recommendations

In this project experimental, numerical, and analytical techniques have been presented to

forensically evaluate the potential failures of compression fittings by incorrect installation.

The main observations and analysis of forensic evidence are summarised as follows,

By correlating the contact pressure of the FE simulation with the experimental

hydrostatic failure pressure and also observations it results that the likely failure

mechanism is due to the initial liquid penetration under the ferrule. This serves as a

lubricant to drastically reduce the coefficient of friction. The pipe and compression

fitting then transiently separate under the lubricants influence, rather than

maintaining a steady state leakage rate. The theory is supported by observing

„pops‟ as the pipe works axially along the fitting, initiated by the ferrule being

displaced from its deformation seat. The pops are time dependent, and if there is a

sudden increase in pressure there would be immediate complete separation.

It can be clearly seen that an exponential leakage pressure growth is

developing from the experimental results of Fig. 4.7. This is supported by the

basic leakage solution of Eq. (2.19).

An axial tilt of the ferrule, causing it to sit in the incorrect position, produced an

ability to operate at 64% of mean leakage pressure.

Often in cases of under-tightening the nut trace evidence shows that score marks

are left on the end of the pipe when separation and leakage occur. For a 15mm pipe

diameter the score marks are within 17mm of the end of the pipe. The score marks

are much more defined on 1/3 turns than 1/4 turns (see Fig. 4.3). On 1/12 and 1/6

turns the score marks were not visible. This is because the initial ferrule


63

deformation at serves to create a homogeneous contact with the fitting body and nut

and rather than an interference fit with the pipe.

When the pipe separates the fitting remains on the other side of the pipe. It has been

observed that the nut on the separated end remains intact, with the ferrule contained

within, and empirically shown that the nut tightness is consistent with the original

assembly. If physical evidence is available in a litigation case, it would be simple to

measure the horizontal distance and/or the amount of turns, and compare with a

control value obtained from another specimen. Comparing this with the

manufacturers recommended installation would immediately and undisputedly

support liability.

It was impossible to test over tightening of the specimens because of the limit in

thread depth as imposed by design.

Some compression fitting designs do not conform to BS EN1254-2:1998, even

though they are Kitemarked. Specifically they do not give the ability to allow

torque transmission. Not allowing a separate fixed position may induce forces

normal to the pipe, and can cause the pipe to drift from its correct position. In

litigation it may prove that the manufacturers are at fault by design, but normal

forces and their twisting effect would have to be corroborated in causing a water

escape. Examples of this design are shown in Fig. 4.10.

The failure pressure to cause leakage is very high in comparison to normal

operating pressures. The empirical results showed that under-tightening of the nut

by as much as 87% still produced an adequate seal to prevent leakage below

0.9MPa. However the tests eliminated all other potential failure variables and as is

often the case with engineering failures multiple factors can influence the final

failure mechanism. Some of these factors are described in Section 2.1.5.


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65

6. References

1. Brown, S., 2007, Forensic engineering: Reduction of risk and improving technology (for all things great

and small), Engineering Failure Analysis, V14, I6, pp. 1019-1037

2. Conex, 2011, Compression system technical specifications and product range, IBP

3. BES, 2011, Pipe Fittings – Compression Fittings, [online] Available at:

<www.bes.co.uk/products/140.asp?ref=MMS-compressionfittings.org.uk> [Accessed 07.2011]

4. Betabite Hydraulics, 2011, Useful Data - General, [online] Availible at:

<www.betabite.co.uk/technical.htm> [Accessed 13.07.2011]

5. BS EN 1254-2, 2007, Draft for public comment, Copper and copper alloys - Plumbing fittings, Part 2:

Fittings with compression ends for use with copper tubes, V7, British Standards Publishing, p. 6

6. Conex, 2011, www.ibpconex.co.uk, IBP, UK

7. Comap, 2011, www.comapindia.com, Groupe Alberties, India

8. Kuterlite, 2011, www.pegler.co.uk/EN/Products/Kuterlite/Specials, Pegler-Yorkshire, UK

9. BS EN 1254-2, 1998, Copper and copper alloys - Plumbing fittings, Part 2: Fittings with compression

ends for use with copper tubes, British Standards Publishing, p. 3

10. Ibid. [2], p. 7

11. Ibid. p. 5

12. Kuterlite, 2011, Compression Fittings, Pegler Yorkshire Group Ltd., p. 16

13. Ibid. [5], p. 5

14. Conex, 2011, Conex Product Guide, IBP, p. 5

15. Stachowicz, F., 2000, Bending with upsetting of copper tube elbows, J. of Materials Processing

Technology, V100, I1-3, pp. 236-240


66

16. Murray, N.W., Bilston, P., 1992, Local buckling of thin-walled pipes being bent in the plastic range,

Thin-Walled Structures, V14, I5, p. 430

17. Rees, D.W.A., 1987, Application of classical plasticity theory to non-radial loading paths, Proceedings of

the Royal Society London, A410, p. 443

18. Sljapic, V., et. al., 2002, Observations on fracture in axi-symmetric and three-dimensional cold upsetting

of brass, J. of Materials Processing Technology, V125-126, pp. 267-274

19. Goto, H., Ashida, M., 1988, Friction and wear of brass during fretting corrosion under various

environmental conditions, Tribology International, V21, I4, p. 185-Fig. 4

20. Liu, T., 1964, Sliding friction of copper, Wear, V7, I2, pp. 163-174

21. Megahed, M.M., 1990, Elastic-plastic behaviour of a thick-walled tube with general nonlinear hardening

properties, International J. of Mechanical Sciences, V32, I7, p. 554

22. Rees, D.W.A., 2000, Mechanics of Solids and Structures, London: Imperial College Press, p. 46

23. UK Copper Board, 2011, Jointing Copper Tubes - Compression joints, Copper Development Association,

p.13

24. Fenox, 2011, LS-X Product, Cookson Electronics, [online] Available at:

<www.fernox.com/products/traditional+plumbing+products/jointing+compounds/ls-x> Accessed [07.2011]

25. Timoshenko, S., 1951, Theory of Elasticity, PA: McGraw-Hill, p. 305

26. Timoshenko, S., 1940, Strength of Materials - Part I, 2nd Ed., London: Macmillian and Co., p. 205

27. Ibid., p. 52

28. Lamé, G., Clapeyron, B., 1833, Mémoire on the Interior Balance of Homogeneous Solid Bodies, Mem.

Académie des Sciences, Paris, Vol. 4

29. Timoshenko, S., 1959, Theory of plates and shells, NY: MacGraw-Hill

30. Burton, P., 1962, Shrink-fits ofmoderately long bands on thin walled cylinders, J. of Engineering for

Industry, V84, I3, pp. 338-342


67

31. Singh, Sadhu, 1994, On the shrink-fit problem of a thin cylinder, International J. of Pressure Vessels and

Piping, V60, I2, pp. 167-175

32. Klosner, J.M., Kempner, J., 1963, Comparison of elasticity and shell theory solutions, American Institute

of Aeronautics and Astronautics, V1, pp. 627–630

33. Klosner, J.M., Levine, H. S., 1966, Further comparison of elasticity and shell theory solutions, American

Institute of Aeronautics and Astronautics, V4, pp. 467–480

34. Iyengar, K.T., Sebastian, V. K., 1972, Comparison of elasticity and shell-theory solutions for finite

circular cylindrical shells, Nuclear Engineering and Design, V21, I1, pp. 137-157

35. McVee, J.D., 1994, The axisymmetric deformation of anisotropic cylindrical shells, Marine Structures,

V7, I2-5, pp. 257-305

36. Steven, G.P., 1977, A non-axisymmetric cylindrical contact problem, International J. of engineering

Science, V15, I2, , pp. 95-103

37. Love, A.E.H, 1944, A Treatise on the mathematical Theory of Elasticity, NY: Dover Press

38. Steven, G.P., 1975, The shrink-fit problem with both components being elastic, International J. of

Engineering Science, V13, I7-8, pp. 663-673

39. Flügge, W., 1973, Stresses in shells, Berlin: Springer

40. Hasan S, et. al., 2001, Determination of friction coefficient by employing the ring compression test, J. of

Engineering Materials Technology, V123, pp.338–48

41. Robinson, T., 2004, Study on ring compression test using physical modelling and FE simulation, J. of

Materials Processing Technology, V153-154, pp. 54-59

42. Zhu., Y, et. al., 2011, Determination of the friction factor of Ti-6Al-4V titanium alloy in hot forging by

means of ing-compression test using FEM, Tribology International, doi:10.1016/j.triboint

43. Carter, W.T., Lee, D., 1985, A finite element analysis of cylinder and ring compression and its

experimental verification, Computers and Structures, V21, I1-2, pp. 1-19


68

44. Hutchinson, J.W., 1965, Axial buckling of pressurized imperfect cylindrical shells, American Institute of

Aeronautics and Astronautics, V3, pp.1461–1466

45. Mathon, C., Limam, A., 2006, Experimental collapse of thin cylindrical shells submitted to internal

pressure and pure bending, Thin-Walled Structures, V44, I1, pp. 39-50

46. Ibid. [25], p. 378

47. Hertz, H., 1881, On the contact of Elastic Solids, J. of Pure and Applied Mathematics, V92, pp. 156-171

48. Greenwood, J.A., Williamson, J.B.P., 1966, Contact of nominally flat surfaces, Proceeedings of the Royal

Society London, A 295, pp. 300–319

49. Acheson, D. J., 1990, Elementary Fluid Dynamics, Oxford: Oxford University Press

50. White, F. M., 2003, Fluid Mechanics, 5th Ed., NY: McGraw-Hill

51. Tournerie, B., Frene, J., 1984, Principal research areas on mechanical face-seals, Tribology International,

V17, I4, pp. 179-184

52. Lebeck, A., 1991, Principles and Designs of Mechanical Face Seals, NY: John-Wiley

53. Metcalfe, R., 1978, Predicted effects of sealing gap convergence on performance of plain end face seals,

American Society of Lubrication Engineers, V21, I2, pp. 177-185

54. Lebeck, A. O., 1988, Contacting mechanical seal design using a simplified hydrostatic model, Tribology

International, V21, I1, pp. 2-14

55. Lebeck, A. O., 1999, Proceedings of the Institution of Mechanical Engineers, Part J, J. of Engineering

Tribology, pp. 163-175

56. Ibid. [9]

57. Ibid., p. 10

58. Abaqus, 2008, Abaqus/CAE Users Manual, Linear and Nonlinear Procedures, Section 14.3.2, V6.9,

Rhode Island: Dassault Systèmes

59. Hollomon, J.H., 1945, Tensile deformation, Metals Technology, V12, pp. 268–290


69

60. Sing, W. M., Rao, K. P., 1997, Role of strain-hardening laws in the prediction of forming limit curves, J.

of Materials Processing Technology, V63, I1-3, p. 106

61. Ebrahimi, R., Pardis, N., 2009, Determination of strain-hardening exponent using double compression

test, Materials Science and Engineering: A, V518, I1-2, pp. 56-60

62. Ludwik, P., 1909, Elemente der Technologischen Mechanic, Verl. Julius Springer

63. Sun, E. Q., 2006, Shear locking and hourglassing in msc nastran, Abaqus, and Ansys, technical report,

MSC Software, Virtual Product Development Conference (VPD2006), Paper No. 27, CA

64. Prathap, G., 2005, Finite Element Analysis as Computation, India

65.Ibid. [58] - The Interaction Module, Section 15.1

66. Ibid. - Understanding Interactions, Section 15.3

67. Ibid. - Defining Contact Pairs in Abaqus/Standard - Defining contact between two separate surfaces,

Section 32.3.1

68. Abaqus, 2008, Abaqus/CAE Theory Manual, Interface Modelling, Coloumb Friction, Section 5.2.3, V6.9,

Rhode Island: Dassault Systèmes

69. Abaqus, 2008, Abaqus/CAE Analysis Users Manual, Using Substructures - Defining kinematic

constraints and transformations, Section 10.1.1, V6.9, Rhode Island: Dassault Systèmes

70. Ibid. - Linear constraint equations, Section 31.2.1

71. BS EN 1057, 2006, Copper and copper alloys -Seamless, round copper tubes for water and gas in

sanitary and heating applications, British Standards Publishing, p. 27

72. Ibid. [58] - Top-down meshing, Section 17.3.4

73. Lobina, E., 2001, UK Water Privatisation – a Briefing, Public Services International Research Unit, p. 4

74. Hertele, S., et. al., 2011, A generic stress-strain model for metallic materials with two-stage strain

hardening behaviour, International J. NonLinear Mechanics, V46, I3, pp. 519-531

75. Zhu, X.K., Leis, B.N., 2005, Influence of yield-to-tensile strength ratio on failure assessment of corroded

pipelines, J. Pressure Vessel Technology, V127, pp. 436–442


70

76. Ibid. [9], p. 11

77. BS EN 1254-4, 1998, Copper and copper alloys - Plumbing fittings, Part 4: Fittings combining other end

connections with capillary or compression ends, British Standards Publishing, p. 4

78. Ibid. [71], p. 12

7. Bibliography

[1] Moran, M. J., et. al., 2003, Introduction to Thermal Systems Engineering, Wiley and Sons


71

8. Appendix

Table 8.1. Steps in a forensic investigation outlined by S. Brown [1].

1.0 Preliminaries/action list (iterative)

2.0 Background (data collection)

2.1 Document incident/visit scene of accident/examine physical evidence at scene

2.2 Eyewitness and cognizant personnel accounts

2.3 Data recorded during incident and after incident (reports, photos, tapes, digital, etc.)

2.4 Product, construction, fabrication specifications

2.5 Service and maintenance history

2.6 Codes and Standards

2.7 Case histories, historical data

2.8 Public and private records

3.0 Lab analysis (NDE & DE examination & testing of specimen, part, component, system)

4.0 Determination of mode and mechanism and hazards of failure.

5.0 Simulations (theoretical and/or experimental tests)

6.0 Reconstruction (the sequence of events)

7.0 Causality/areas of responsibility for the incident consequences

7.1 Engineering design

7.2 Material selection

7.3 Material defect

7.4 Manufacturing/fabrication

7.5 Operation/use

7.6 Inspection

7.7 Maintenance

Causal factors for the incident

(a) strength/performance

(b) code compliance

(c) human factors

(d) warnings

(e) risks

(f) cost

(g) reliability

(h) acceptable use

(i) means of determining unfitness

(j) operation & use instructions

(k) inspection instructions

(l) maintenance instructions

(m) allowable defects

(n) material performance

(o) tests & analysis

(p) transportation

(q) packaging

(r) advertisement

(s) protection

(t) hazards identification

(u) specifications

8.0 Recommendations (to improve safety and loss prevention)


72

8.1 Results Data

Table 8.2. Uniaxial test parameters and results

Material Material

Parameter

Nlgeom ζvM (MPa) εmax

PEEQ

(εp %) On Off

Brass

Elastic x 50030 0.4350 0

x

41090 0.3612* 0

Elastic-perfectly

plastic

x 140 0.4350 0.4338

x

140 0.3612* 0.3561

Strain hardened

curve

x 404.2 0.4350 0.4315

x

389.0 0.3612* 0.3539

Copper

Elastic x 36800 0.4 0

x 30670 0.3365* 0

Elastic-perfectly

plastic

x 65 0.4 0.3993

x 65 0.3365* 0.3326

Strain hardened

curve

x 354.9 0.4 0.3961

x 338.4 0.3365* 0.3297

* Principal strain component εmax = lnΔL

L

Table 8.3. NOT SHOWN IN THIS PREVIEW


73

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Forensic Investigation of Compression Fitting Failures W. Shipway

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76

Table 8.4. A Gantt chart timeline.

Month JUNE JULY AUG SEPT

Weeks 1 2 3 4 1 2 3 4 1 2 3 4 1 2 3 4

PROJECT

Project choice

Report writing

Background

(research)

Proof reading

Print and submit

SETUP

FE simulation

Engineering

Analysis

Equipment

procurement

Materials

procurement

Experiment

RESULTS

Experiment

FE simulation

expected actual


77

8.3. Email correspondence with Conex

from: Warwick Shipway

to: [email protected]

date: 14 July 2011 21:49

subject: Compression fitting data

Hi,

Could you possibly provide me with some dimensions for the brass compression fittings

you manufacture. I have the Compression system technical specifications and product

range data sheet, but it doesn't have all of the dimensions required.

It‟s for type-A fittings with the nut, olive, and fitting body with external thread. I believe

they are made for pipe sizes 6, 8, 12, 15, 18, 22, 28, 35, 42, 54 (mm)?

Thanks,

Warwick

--

Warwick Shipway

BEng Hons (2008), Mechanical Engineer

07726020166

from: James (bob) McCunn <[email protected]

to: Warwick Shipway

date: 15 July 2011 08:09

subject: RE: Compression fitting data

Warwick,

Thank you for your enquiry.

The Conex compression fittings are designed for use with copper tube, carbon steel, and

stainless steel in sizes 6mm to 54mm inclusive. Additional sizes to those you have listed

are 10mm, 14mm, & 16mm.

Copper tube is to BS EN 1057, stainless steel tube to EN 10312, and carbon steel tube (not

for potable water) to EN 10305.

It is not company policy to issue drawings with critical dimensions, as these are

confidential.

Regards,

Bob. Bob McCunn - Senior Laboratory Engineer Conex Universal Limited, Whitehall Road, Tipton, West Midlands, UK, DY4 7JU

Direct Line: +44 (0)121 521 2902 Ext: 232 Fax: +44 (0)121 557 7936

Email: [email protected]

Web: www.IBPGroup.com | www.IBPConex.co.uk

Conex Bänninger- DOING MORE since 1909

IBP E-Mail Disclaimer:- This E-mail is solely for the intended addressee, is strictly confidential and may also be

legally privileged. If you are not the addressee please do not read, print, re-transmit, store

or act in reliance on it or any attachments. Instead, please E-mail it back to the sender and

then immediately permanently delete it. Please ensure you have appropriate virus

protection before opening any attachment.

Conex Universal Limited, Whitehall Road, Tipton, West Midlands, DY4 7JU.

Registered in England and Wales No. 07563347

tel:%2B44%20%280%29121%20521%C2%A02902

tel:%2B44%20%280%29121%20557%C2%A07936

http://www.ibpgroup.com/

http://www.ibpconex.co.uk/


78

Figure 8.1. Above 2000psi the pipes structural integrity was compromised.

forensic investigation of compression...

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