Pipelines stress analysis Report [Mechatronics]

8/4/2019 Pipelines stress analysis Report [Mechatronics]

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MINISTY OF HIGHER EDUCATIONAINSHAMS UNIVERSITY

FACULTY OF ENGINEERING3RD YEAR MECHATRONICS DEPARTMENT

Mechatronics (1)

Pipelines stress analysis

Supervised By : Dr Wagdy El-Desouki Abdel-Ghany

Group: 5

Sec. 1 Sec. 1

Sec. 1 Sec. 2

Sec. 2 Sec. 2

Submitting date06/2011


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Contents

1 Introduction ............................................................................................................ 42 Objective and scope .............................................................................................. 63 Methodology ........................................................................................................... 7

3.1 Classification of loads and failure modes ........................................................................................................ 73.2 Primary Loads .................................................................................................................................................. 73.3 Secondary Loads .............................................................................................................................................. 83.4 Static vs. Dynamic loads ................................................................................................................................ 103.5 Sustained vs. Occasional loads ...................................................................................................................... 123.6 The Stresses ................................................................................................................................................... 12

4 Normal and Shear Stresses from Applied Load ................................................ 144.1 Axial Load ...................................................................................................................................................... 15 4.2 Internal / External Pressure ........................................................................................................................... 164.3 Bending Load ................................................................................................................................................. 184.4 Shear Load ..................................................................................................................................................... 224.5 Torsional Load ............................................................................................................................................... 22

5 Allowable stresses & theories of Failure ........................................................... 235.1 Maximum Stress Theory ................................................................................................................................ 255.2 Maximum Shear Theory ................................................................................................................................ 275.3 Octahedral Shear Theory ............................................................................................................................... 28

6 Design under Secondary Load ........................................................................... 287 Piping Codes: ....................................................................................................... 29

7.1 Limits of stresses set by code ANSI / ASME B 31.3 ........................................................................................ 30


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7.2 ASME B31.1 power piping ............................................................................................................................. 31 7.3 ASME B31.3 Process Piping ........................................................................................................................... 46 7.4 ASME B31.9 Building Services Piping............................................................................................................. 59

8 Flexibility analysis ............................................................................................... 608.1 Methods of Flexibility Analysis ...................................................................................................................... 61

9 Pipe supports ....................................................................................................... 639.1 Introduction ................................................................................................................................................... 639.2 Pipe supports standards ................................................................................................................................ 649.3 Types of supports .......................................................................................................................................... 649.4 Pipe system support designing ...................................................................................................................... 75

10 Buried Pipe Design .............................................................................................. 8310.1 Introduction and Overview ....................................................................................................................... 8310.2 Soil Mechanics .......................................................................................................................................... 8510.3 Strength of Materials ................................................................................................................................ 8610.4 Water Systems .......................................................................................................................................... 8910.5 Design for Value ....................................................................................................................................... 90

11 Conclusion.......................................................................................................... 105


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1 Introduction

Pipes are the most delicate components in any process plant. They are also the busiest

entities. They are subjected to almost all kinds of loads, intentional or unintentional. It is

very important to take note of all potential loads that a piping system would encounter

during operation as well as during other stages in the life cycle of a process plant.

Ignoring any such load while designing, erecting, hydro-testing, start-up shut-down,

normal operation, maintenance etc. can lead to inadequate design and engineering of a

piping system. The system may fail on the first occurrence of this overlooked load.

Failure of a piping system may trigger a Domino effect and cause a major disaster.

Stress analysis and safe design normally require appreciation of several related

concepts.

An approximate list of the steps that would be involved is as follows.

1. Identify potential loads that would come on to the pipe or piping system during its

entire life.

2. Relate each one of these loads to the stresses and strains that would be

developed in the crystals/grains of the Material of Construction (MoC) of the

piping system.

3. Decide the worst three dimensional stress state that the MoC can withstand

without failure

4. Get the cumulative effect of all the potential, loads on the 3-D stress scenario in

the piping system under consideration.

5. Alter piping system design to ensure that the stress pattern is within failure limits.

The goal of quantification and analysis of pipe stresses is to provide safe design

through the above steps. There could be several designs that could be safe. A piping

engineer would have a lot of scope to choose from such alternatives, the one which is

most economical, or most suitable etc. Good piping system design is always a mixture

of sound knowledge base in the basics and a lot of ingenuity.


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Pipes are required for carrying fluids. These fluids can be of various states of matter.

Gaseous fluids (like LP), Liquid Fluids (like Water) and Solid or Semi-solid (like plastic

pellets).

The pipes in Process Industry like in Reliance are used for transferring fluids at highertemperature and pressure. The various processes in a Process plant cause the liquids

to be pressurized and to be heated up. Thus the liquids passing through the pipes attain

a high pressure and/or a high temperature. When a metal is heated it expands. If this

metal of pipe is allowed to expand freely, there is no overstress in the same. But

suppose the free movement is restricted by any means, stress is introduced in the

system. The case becomes more complicated by considering weight of the pipe, the

insulation, weights of the valves, flanges and other fittings and the pressure of the fluids

that is flowing through the piping.

So the task of the Stress Engineer is

1) To select a piping layout with an adequate flexibility between points of anchorage to

absorb its thermal expansion without exceeding allowable material stress levels, also

reacting thrusts & moments at the points of anchorage must be kept below certain

limits.

2) To limit the additional stresses due to the dead weight of the piping by providing

suitable supporting system effective for cold as well as hot conditions. Piping systems

are not self supporting and hence they require pipe supports to prevent from collapsing.

Pipe supports are of different types like Rest, Guides, Line stops, Hangers, Snubbers,

and Struts. Each type of pipe support plays a vital role in supporting the pipe system.

Pipe supports are desirable to reduce the weight, wind and where possible, expansion

and transient effects, so that piping system stress range is not excessive for the

anticipating cycles of operation, us avoiding fatigue failure. Limiting the line movement

at specific locations may be desirable to protect sensitive equipment, to control vibration

or to resist external influences such as wind, earthquake, or shock loadings.

All these objectives are achieved by:-


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1) Limiting the sagging of the piping system within allowable limits ( i.e. In Sustain case

the max vertical movement should be less than 10mm ).

2) Directing the line movements so as protect sensitive equipments against overloading

(i.e. nozzle loads are always kept under the allowable nozzle loading provided by thevendor).

3) Resisting pipe system to collapse in case of earthquake, wind or shock loadings.

2 Objective and scope

With piping, as with other structures, the analysis of stresses may be carried to varying

degrees of refinement. Manual systems allow for the analysis of simple systems,

whereas there are methods like chart solutions (for three-dimensional routings) and

rules of thumb (for number and placement of supports) etc. involving long and tedious

computations and high expense. But these methods have a scope and value that

cannot be defined as their accuracy and reliability depends upon the experience and

skill of the user. All such methods may be classified as follows:

1. Approximate methods dealing only with special piping configurations of two-, three or

four-member systems having two terminals with complete fixity and the piping layoutusually restricted to square corners. Solutions are usually obtained from charts or

tables. The approximate methods falling into this category are limited in scope of direct

application, but they are sometimes usable as a rough guide on more complex

problems by assuming subdivisions of the model into anchored sections fitting the

contours of the previously solved cases.

2. Methods restricted to square-corner, single-plane systems with two fixed ends, but

without limit as to the number of members.

3. Methods adaptable to space configurations with square corners and two fixed ends.

4. Extensions of the previous methods to provide for the special properties of curved

pipe by indirect means, usually a virtual length correction factor.


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3 Methodology

3.1 Classification of loads and failure modes

Pressure design of piping or equipment uses one criterion for design. Under a steady

application of load (e.g. pressure), it ensures against failure of the system as perceived

by one of the failure theories. If a pipe designed for a certain pressure experiences a

much higher pressure, the pipe would rupture even if such load (pressure) is applied

only once.

The failure or rupture is sudden and complete. Such a failure is called catastrophic

failure.

It takes place only when the load exceeds far beyond the load for which design was

carried out. Over the years, it has been realized that systems, especially piping,

systems can fail even when the loads are always under the limits considered safe, but

the load application is cyclic (e.g. high pressure, low pressure, high pressure, ..). Such a

failure is not guarded against by conventional pressure design formula or compliance

with failure theories. For piping system design, it is well established that these two types

of loads must be treated separately and together guard against catastrophic and fatigue

failure.

The loads the piping system (or for that matter any structural part) faces are broadly

classified as primary loads and secondary loads.

The failure of the piping system may be sudden failure due to onetime loading or fatigue

failure due to cyclic loading. The sudden failure is attributed to primary loadings and the

fatigue failure to secondary loading.

3.2 Primary Loads

These are typically steady or sustained types of loads such as internal fluid pressure,

external pressure, gravitational forces acting on the pipe such as weight of pipe and


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fluid, forces due to relief or blow down pressure waves generated due to water hammer

effects.

The last two loads are not necessarily sustained loads. All these loads occur because of

forces created and acting on the pipe. In fact, primary loads have their origin in someforce acting on the pipe causing tension, compression, torsion etc leading to normal and

shear stresses. A large load of this type often leads to plastic deformation. The

deformation is limited only if the material shows strain hardening characteristics. If it has

no strain hardening property or if the load is so excessive that the plastic instability sets

in, the system would continue to deform till rupture. Primary loads are not self-limiting. It

means that the stresses continue to exist as long as the load persists and deformation

does not stop because the system has deformed into a no-stress condition but because

strain hardening has come into play.

In short

Primary loads are usually force driven (gravity pressure, spring forces, relief

valve, fluid hammer etc.)

Primary loads are not self-limiting. Once plastic deformation begins it continues

till the failure of the cross section results.

Allowable limits of primary stresses are related to ultimate tensile strength.

Primary loads are not cyclic in nature.

Design requirements due to primary loads are encompassed in minimum wall

thickness requirements.

3.3 Secondary Loads

Just as the primary loads have their origin in some force; secondary loads are caused

by displacement of some kind. For example, the pipe connected to a storage tank may

be under load if the tank nozzle to which it is connected moves down due to tank

settlement.

Similarly, pipe connected to a vessel is pulled upwards because the vessel nozzle

moves up due to vessel expansion. Also, a pipe may vibrate due to vibrations in the


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rotating equipment it is attached to. A pipe may experience expansion or contraction

once it is subjected to temperatures higher or lower respectively as compared to

temperature at which it was assembled.

The secondary loads are often cyclic but not always. For example load due to tanksettlement is not cyclic. The load due to vessel nozzle movement during operation is

cyclic because the displacement is withdrawn during shut-down and resurfaces again

after fresh start-up. A pipe subjected to a cycle of hot and cold fluid similarly undergoes

cyclic loads and deformation. Failure under such loads is often due to fatigue and not

catastrophic in nature.

Broadly speaking, catastrophic failure is because individual crystals or grains were

subjected to stresses which the chemistry and the physics of the solid could not

withstand.

Fatigue failure is often because the grains collectively failed because their collective

characteristics (for example entanglement with each other etc.) changed due to cyclic

load. Incremental damage done by each cycle to their collective texture accumulated to

such levels that the system failed. In other words, catastrophic failure is more at

microscopic level, whereas fatigue failure is at mesoscopic level if not at macroscopic

level.

In short

Secondary loads are usually displacement driven (Thermal expansion,

Settlement, Vibration etc.)

Secondary loads are self-limiting i.e. the loads tends to dissipate as the system

deforms through yielding.

Allowable loads for secondary stresses are based upon fatigue failure modes. Secondary loads are cyclic in nature (expect settlement).

Secondary application of load never produces sudden failure and sudden failure

occurs after a number of applications of load.


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3.4 Static vs. Dynamic loads

Static loads are those loads applied on to the piping system so slowly that the system

has time to respond, react and also to disturb the load. Hence, the system remains in

equilibrium. The examples of such loadings are the thermal expansion, weight etc.

The dynamic load changes so quickly with time that the system will have no time to

distribute the load. Hence the system develops unbalanced forces.

The examples of Dynamic loadings are wind load, earthquake, fluid hammer etc. these

can be categorized in to mainly three types:-

Random:

In this type of loading the load changes unpredictably with time. The major loads

covered under this type are :-

Wind load:

In most of the cases analysis is done using static equivalent of dynamic model. This is

achieved by increasing the static loading by a factor to account for the dynamic effects.

Earthquake:

Here again the analysis is done using static equivalent of a dynamic loading model. This

is again is approximate.

Harmonic:

In harmonic type of profile, the load changes in magnitude and direction in a sine profile.

The major loads covered under this are:-

Equipment Vibration:

This is mainly caused by the eccentricity of the equipment drive shaft of the rotating

type of equipment connected to the piping.


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Acoustic Vibration:

This is mainly caused by change of fluid flow condition within pipe i.e.from laminar to

turbulent e.g. Flow through orifice. Mostly these vibrations follow harmonic patterns with

predictable frequencies based on flow conditions.

Pulsation:

This type of loading occurs due to flow from reciprocating pumps, compressors etc. if

this type of profile the loading starts from zero to some value, remains there for certain

period of time and then comes back to zero. The major types of loads covered under

this are:-

Relief valve outlet:

When the relief valve opens the flow raises from zero to full value over the opening time

of the valve. This causes a jet forces and this remains until the full venting is achieved

to overcome the over pressure situation and then valve closes bringing down the force

over the closing time to valve.

Fluid hammer:

If the flow of fluid is suddenly stopped due to pump trip or sudden closing of valve, there

will compression of fluid at one side and relaxation at the other side. This wave

propagates causing pulsation flow.

Slug flow:

This happens mainly due to multi phase flow. In general when fluid changes direction in

a piping system, it is balanced by net force in the elbow. This force is equal to change in

momentum with respect to time. Normally this force is constant and can be absorbed

through tension in pipe wall, to be passed on to adjacent elbow which may have equal

and opposite load and gets nullified. Hence, these are normally ignored.

However, if density of fluid velocity changes with time similar to slug of liquid in a gas

system, this momentum load will change with time as well leading to dynamic load.


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3.5 Sustained vs. Occasional loads

The loads on the piping system which are steady and developed due to internal

pressure, external pressure, weight etc. affecting the structure design of the piping

component are called the sustained loading. These loadings develop longitudinal, shear

or hoop stresses in the pipe wall. These could be either tensile or compressive in

nature.

3.6 The Stresses

The MoC of any piping system is the most tortured non-living being right from its birth.

Leaving the furnace in the molten state, the metal solidifies within seconds. It is a very

hurried crystallization process. The crystals could be of various lattice structural patterns

such as BCC, FCC, HCP etc. depending on the material and the process. The grains,

crystals of the material have no time or chance to orient themselves in any particular

fashion. They are thus frozen in all random orientations in the cold harmless pipe or

structural member that we see.

When we calculate stresses, we choose a set of orthogonal directions and define the

stresses in this co-ordinate system. For example, in a pipe subjected to internal

pressure or any other load, the most used choice of co-ordinate system is the one

comprising of axial or longitudinal direction (L), circumferential (or Hoope's) direction (H)

and radial direction (R) as shown in figure. Stresses in the pipe wall are expressed as

axial (SL), Hoope's (SH) and radial (SR). These stresses which stretch or compress a

grain/crystal are called normal stresses because they are normal to the surface of the

crystal.

But, all grains are not oriented as the grain in Figure 1. In fact the grains would have

been oriented in the pipe wall in all possible orientations. The above stresses would

also have stress components in direction normal to the faces of such randomly oriented


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crystal. Each crystal thus does face normal stresses. One of these orientations must be

such that it maximizes one of the normal stresses.

Figure 1: Hoope's, radial and axial stresses

The mechanics of solids state that it would also be orientation which minimizes some

other normal stress. Normal stresses for such orientation (maximum normal stress

orientation) are called principal stresses, and are designated S1 (maximum), S2 and S3

(minimum). Solid mechanics also states that the sum of the three normal stresses for all

orientation is always the same for any given external load.

That is

In addition to the normal stresses, a grain can be subjected to shear stresses as well.

These acts parallel to the crystal surfaces as against perpendicular direction applicable

for normal stresses. Shear stresses occur if the pipe is subjected to torsion, bending

etc. Just as there is an orientation for which normal stresses are maximum, there is an

orientation which maximizes shear stress. The maximum shear stress in a 3-D state of

stress can be shown to be

i.e. half of the difference between the maximum and minimum principal stresses. The

maximum shear stress is important to calculate because failure may occur or may be

deemed to occur due to shear stress also. A failure perception may stipulate that

maximum shear stress should not cross certain threshold value. It is therefore


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necessary to take the worst-case scenario for shear stresses also as above and ensure

against failure.

It is easy to define stresses in the co-ordinate system such as axial-Hoope s-radial (L-H-

R) that are defined for a pipe. The load bearing cross-section is then well defined andstress components are calculated as ratio of load to load bearing cross-section.

Similarly, it is possible to calculate shear stress in a particular plane given the torsional

or bending load. What are required for testing failure - safe nature of design are,

however, principal stresses and maximum shear stress. These can be calculated from

the normal stresses and shear stresses available in any convenient orthogonal co-

ordinate system. In most pipe design cases of interest, the radial component of normal

stresses (SR) is negligible as compared to the other two components (SH and SL). The

3-D state of stress thus can be simplified to 2-D state of stress. Use of Mohr's circle

then allows calculating the two principle stresses and maximum shear stress as follows:

The third principle stress (minimum i.e. S3) is zero.

All failure theories state that these principle or maximum shear stresses or some

combination of them should be within allowable limits for the MoC under consideration.

To check for compliance of the design would then involve relating the applied load to

get the net SH, SL, and then calculate S1, S2and max and some combination of them.

4

Normal and Shear Stresses from Applied LoadAs said earlier, a pipe is subjected to all kinds of loads. These need to be identified.

Each such load would induce in the pipe wall, normal and shear stresses. These need

to be calculated from standard relations. The net normal and shear stresses resulting in

actual and potential loads are then arrived at and principle and maximum shear


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stresses calculated. Some potential loads faced by a pipe and their relationships to

stresses are summarized here in brief

4.1 Axial Load

A pipe may face an axial force (FL) as shown in Figure 2. It could be tensile or

compressive.

Figure 2: Pipe under axial load

What is shown is a tensile load. It would lead to normal stress in the axial direction (SL).

The load bearing cross-section is the cross-sectional area of the pipe wall normal to the

load direction, Am. The stress can then be calculated as

The load bearing cross-section may be calculated rigorously or approximately as

follows:

The axial load may be caused due to several reasons. The simplest case is a tall

column.


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The metal cross-section at the base of the column is under the weight of the column

section above it including the weight of other column accessories such as insulation,

trays, ladders etc. Another example is that of cold spring. Many times a pipeline is

intentionally cut a little short than the end-to-end length required. It is then connected to

the end nozzles by forcibly stretching it. The pipe, as assembled, is under axial tension.

When the hot fluid starts moving through the pipe, the pipe expands and compressive

stresses are generated. The cold tensile stresses are thus nullified. The thermal

expansion stresses are thus taken care of through appropriate assembly-time

measures.

4.2 Internal / External Pressure

A pipe used for transporting fluid would be under internal pressure load. A pipe such as

a jacketed pipe core or tubes in a Shell & Tube exchanger etc. may be under net

external pressure. Internal or external pressure induces stresses in the axial as well as

circumferential (Hoope s) directions. The pressure also induces stresses in the radial

direction, but as argued earlier, these are often neglected.

The internal pressure exerts an axial force equal to pressure times the internal cross-

section of pipe.

This then induces axial stress calculated as earlier. If outer pipe diameter is used for

calculating approximate metal cross section as well as pipe cross- section, the axial

stress can often be approximated as follows:

The internal pressure also induces stresses in the circumferential direction as shown in

Figure 3.


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Figure 3: Hoope's stresses

The stresses are maximum for grains situated at the inner radius and minimum for

those situated at the outer radius. The Hoope's stress at any in between radial position

(r) is given as follows (Lame's equation)

For thin walled pipes, the radial stress variation can be neglected. From membrane

theory, SH may then be approximated as follows.

Radial stresses are also induced due to internal pressure as can be seen in Figure 4.


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Figure 4: Radial stresses due to internal pressure

At the outer skin, the radial stress is compressive and equal to atmospheric pressure

(Patm ) or external pressure (Pext) on the pipe. At inner radius, it is also compressive

but equal to absolute fluid pressure (Pabs). In between, it varies. As mentioned earlier,

the radial component is often neglected.

4.3 Bending Load

A pipe can face sustained loads causing bending. The bending moment can be related

to normal and shear stresses. Pipe bending is caused mainly due to two reasons:

Uniform weight load and concentrated weight load. A pipe span supported at two ends

would sag between these supports due to its own weight and the weight of insulation (if

any) when not in operation. It may sag due to its weight and weight of hydrostatic test

fluid it contains during hydrostatic test. It may sag due to its own weight, insulationweight and the weight of fluid it is carrying during operation.

All these weights are distributed uniformly across the unsupported span, and lead to

maximum bending moment either at the centre of the span or at the end points of the

span (support location) depending upon the type of the support used.

Let the total weight of the pipe, insulation and fluid be W and the length of the

unsupported span be L (see Figure 5).


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Figure 5: Pipe under bending load

The weight per unit length, w, is then calculated (w = W/L). The maximum bending

moment, Mmax

, which occurs at the centre for the pinned support is then given by the

beam theory as follows:

For Fixed Supports, the maximum bending moment occurs at the ends and is given by

beam theory as follows:

The pipe configuration and support types used in process industry do not confirm to any

of these ideal support types and can be best considered as somewhere in between. Asa result, a common practice is to use the following average formula to calculate bending

moment for practical pipe configurations, as follows:

.


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Also, the maximum bending moment in the case of actual supports would occur

somewhere between the ends and the middle of the span.

Another load that the pipe span would face is the concentrated load. A good example is

a valve on a pipe run (see Figure 6):

Figure 6 : Point load

The load is then approximated as acting at the centre of gravity of the valve and the

maximum bending moment occurs at the point of loading for pinned supports and is

given as:

For rigid supports, the maximum bending moment occurs at the end nearer to the

pointed load and is given as

a is to be taken as the longer of the two arms (a and b) in using the above formula.

As can be seen, the bending moment can be reduced to zero by making either a or b

zero,


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I.e. by locating one of the supports right at the point where the load is acting. In actual

practice, it would mean supporting the valve itself. As that is difficult, it is a common

practice to locate one support as close to the valve (or any other pointed and significant

load) as possible. With that done, the bending moment due to pointed load is minimal

and can be neglected.

Whenever the pipe bends, the skin of the pipe wall experiences both tensile and

compressive stresses in the axial direction as shown in Figure 7:

Figure 7: Stresses on pipe cross section due to bending

The axial stress changes from maximum tensile on one side of the pipe to maximum

compressive on the other side. Obviously, there is a neutral axis along which the

bending moment does not induce any axial stresses. This is also the axis of the pipe.

The axial tensile stress for a bending moment of M, at any location c as measured from

the neutral axis is given as follows.

I is the moment of inertia of the pipe cross-section. For a circular cross-section pipe, I is

given as


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The maximum tensile stress occurs where c is equal to the outer radius of the pipe and

is given as follows.

Where Z (= I/ro) is the section modulus of the pipe.

4.4 Shear Load

Shear load causes shear stresses. Shear load may be of different types. One common

load is the shear force (V) acting on the cross-section of the pipe as shown in Figure 8:

Figure 8: Shear force acting on a pipe

It causes shear stresses which are maximum along the pipe axis and minimum along

the outer skin of the pipe. This being exactly opposite of the axial stress pattern caused

by bending moment and also because these stresses are small in magnitude, these are

often not taken in account in pipe stress analysis. If necessary, these are calculated as

Where Q is the shear form factor and Am is the metal cross-section.

4.5 Torsional Load

This load (see Figure 9) also causes shear stresses.


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Figure 9: Torsional load on a pipe

The shear stress caused due to torsion is maximum at outer pipe radius. And it is given

in terms of the torsional moment and pipe dimensions as follows.

RT is the torsional resistance (= twice the moment of inertia).

All known loads on the pipe should be used to calculate contributions to S L, SH and t.

These then are used to calculate the principal stresses and maximum shear stress.

These derived quantities are then used to check whether the pipe system design is

adequate based on one or more theories of failure.

5 Allowable stresses & theories of Failure

Allowable stresses as specified in the various codes are based on the material

properties. These can be classified in two categories as below.


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Time Independent stresses

Time independent allowable is based on either yield stress or the ultimate tensile

strength measured in a simple tensile test. The yield stress is the elastic limit and that is

the value below which the stresses are proportional to strain and when the load isremoved, there is no permanent distortion. The tensile strength is the highest load,

which the specimen can be subject to without failure.

The code ANSI / ASME B 31.1 permits smaller of of the tensile strength or 5/8 of the

yield strength. ANSI / ASME B 31.3 uses lower of 1/3 of the tensile strength or 2/3 of the

yield strength.

Time dependent stresses

The time dependent allowable is related to Creep rupture strength at high

temperature. This is best explained for a piping system as follows.

Pipe running between two equipments expands as it gets heated up. The increased

length can be accommodated only by straining the pipe as its ends are not free to

move.

This straining induces stress in the pipe. However when the line is cooled during

shutdown to ambient temperature the expansion returns to zero, the straining no longer

exists and hence stress also disappears. Every time the plant starts from a stress free

condition i.e. cold condition and soon gets to stressed with maximum at operating

conditions from cold get stressed with stress reaching maximum at operating condition

and then reducing to zero when operating stops and system cools down.

The actual performances of the piping system do not exactly follow the above path. The

piping system can absorb large displacement without returning to exactly to previousconfiguration. Relaxation to the sustaining level of material will tend to establish a

condition of stability in few cycles, each cycle lowering the upper limit of hot stress until

a state of equilibrium is reached in which the system is completely relaxed and capable

of maintaining constant level of stress. The stress at which the material is relieved due

to relaxation appears as stress in opposite sign. Thus the system which originally was


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stress less could within a few cycles accommodate stress in the cold condition and

spring itself without the application of external load. This phenomenon is called Self

springing. This is also called the Elastic shake down.

Here the maximum stress range is set to 2 Sy or more accurately the sum of hot andthe cold yield stresses in order to ensure eventual elastic cycling.

The degree of self springing will depend upon the magnitude of the initial hot stresses

and temperature, so that while hot stresses will gradually decrease with time, the sum of

the hot and cold stress will stay the same. This sum is called the Expansion Stress

range. This concepts lead to the selection of an allowable stress range.

For materials below the creep range the allowable stresses are 62.5% of the yield

stress, so that bending stress at which plastic flow starts at elevated temperature is 1.6

Sh and by same reasoning 1.6 Sc will be stress at which flow would take place at

minimum temperature. Hence, the sum of this could make the maximum stress the

system could be subjected to without flow occurring in either the hot or cold condition.

Therefore, Smax = 1.6(Sc+Sh).

A piping system in particular or a structural part in general is deemed to fail when a

stipulated function of various stresses and strains in the system or structural part

crosses a certain threshold value. It is a normal practice to define failure as occurring

when this function in the actual system crosses the value of a similar function in a solid

rod specimen at the point of yield. There are various theories of failure that have been

put forth. These theories differ only in the way the above mentioned function is defined.

Important theories in common use are considered here.

5.1 Maximum Stress Theory

This is also called Rankine Theory. According to this theory, failure occurs when the

maximum principle stress in a system (S1) is greater than the maximum tensile principle

stress at yield in a specimen subjected to uni-axial tension test.


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Uni-axial tension test is the most common test carried out for any MoC. The tensile

stress in a constant cross-section specimen at yield is what is reported as yield stress

(Sy) for any material and is normally available. In uni-axial test, the applied load gives

rise only to axial stress (SL) and SH and SR as well as shear stresses are absent. SL is

thus also the principle normal stress (i.e. S1). That is, in a specimen under uni-axial

tension test, at yield, the following holds.

SL = SY, SH = 0, SR = 0

S1 = SY, S2 = 0 and S3 = 0.

The maximum tensile principle stress at yield is thus equal to the conventionally

reported yield stress (load at yield/ cross-sectional area of specimen).

The Rankine theory thus just says that failure occurs when the maximum principle

stress in a system (S1) is more than the yield stress of the material (Sy).

The maximum principle stress in the system should be calculated as earlier. It is

interesting to check the implication of this theory on the case when a cylinder (or pipe) is

subjected to internal pressure.

As per the membrane theory for pressure design of cylinder, as long as the Hoope's

stress is less than the yield stress of the MoC, the design is safe. It is also known that

Hoope's stress (SH) induced by external pressure is twice the axial stress (SL). The

stresses in the cylinder as per the earlier given formula would be:

The maximum principle stress in this case is S2 (=SH). The Rankine theory and the

design criterion used in the membrane theory are thus compatible.


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This theory is widely used for pressure thickness calculation for pressure vessels and

piping design uses Rankine theory as a criterion for failure.

5.2 Maximum Shear Theory

This is also called Tresca theory. According to this theory, failure occurs when the

maximum shear stress in a system max is greater than the maximum shear stress at

yield in a specimen subjected to uni-axial tension test. Note that it is similar in wording,

to the statement of the earlier theory except that maximum shear stress is used as

criterion for comparison as against maximum principle stress used in the Rankine

theory.

In uniaxial test, the maximum shear stress at yield condition of maximum shear testgiven earlier is

The Tresca theory thus just says that failure occurs when the maximum shear stress in

a system is more than half the yield stress of the material (Sy). The maximum shear

stress in the system should be calculated as earlier.

It should also be interesting to check the implication of this theory on the case when a

cylinder (or pipe) is subjected to internal pressure.

As the Hoope's stress induced by internal pressure (SH) is twice the axial stress (SL)

and the shear stress is not induced directly ( = 0) the maximum shear stress in the

cylinder as per the earlier given formula would be


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This should be less than 0.5Sy, as per Tresca theory for safe design. This leads to a

different criterion that Hoope's stress in a cylinder should be less than twice the yield

stress. The Tresca theory and the design criterion used in the membrane theory for

cylinder are thus incompatible.

5.3 Octahedral Shear Theory

This is also called Von Mises theory. According to this theory, failure occurs when the

octahedral shear stress in a system is greater than the octahedral shear stress at yield

in a specimen subjected to uniaxial tension test. It is similar in wording to the statement

of the earlier two theories except that octahedral shear stress is used as criterion for

comparison as against maximum principle stress used in the Rankine theory or

maximum shear stress used in Tresca theory.

The octahedral shear stress is defined in terms of the three principle stresses as

follows.

In view of the principle stresses defined for a specimen under uni-axial load earlier, the

octahedral shear stress at yield in the specimen can be shown to be as follows.

The Von Mises theory thus states that failure occurs in a system when octahedral shear

stress in the system exceeds 2 Sy / 3.

For stress analysis related calculations, most of the present day piping codes uses a

modified version of Tresca theory.

6 Design under Secondary Load


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As pointed earlier, a pipe designed to withstand primary loads and to avoid catastrophic

failure may fall after a sufficient amount of time due to secondary cyclic load causing,

fatigue failure. The secondary loads are often cyclic in nature. The number of cycles to

failure is a property of the material of construction just as yield stress is. While yield

stress is cardinal to the design under primary sustained loads, this number of cycles to

failure is the corresponding material property important in design under cyclic loads aim

at ensuring that the failure does not take place within a certain period for which the

system is to be designed.

While yield stress is measured by subjecting a specimen to uni-axial tensile load,

fatigue test is carried out on a similar specimen subjected to cycles of uni-axial tensile

and compressive loads of certain amplitude, i.e. magnitude of the tensile and

compressive loads. Normally the tests are carried out with zero mean loads. This

means, that the specimen is subjected to a gradually increasing load leading to a

maximum tensile load of W, then the load is removed gradually till it passes through

zero and becomes gradually a compressive load of W (i.e. a load of W), then a tensile

load of W and so on. Time averaged load is thus zero. The cycles to failure are then

measured; the experiments are repeated with different amplitudes of load.

7 Piping Codes:

There are many different piping codes for process and high-pressure lines in use

throughout the world. Some countries, like Canada, take advantage of the considerable

body of knowledge contained in the U.S. codes.

In the U.S. we follow the ASME Code for Pressure Piping, B31. The code was first

published in 1935 by the ASA (American Standards Association, now known as ANSI,

the American National Standards Institute). The responsibility for developing the code

was assigned to the ASME (American Society of Mechanical Engineers).

The ASME Code is so extensive that it was more convenient to break it up into several

separate documents, which represent various industries. The code now consists of:


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B31.1 Power Piping

B31.3 Process Piping

B31.4 Pipeline Transportation Systems for Liquid Hydrocarbons and Other

Liquids

B31.5 Refrigeration Piping

B31.8 Gas Transportation and Distribution Piping

B31.9 Building Services Piping

B31.11 Slurry Transportation Piping Systems.

Usually, by the time you get involved in a project, most of the piping specifications are

written, and the codes to be used have been laid out in the specifications. But how did

the engineer who wrote the specs know which codes to apply? Much of the time the

answer lies in the codes themselves. The codes will explain what their intended scope

is. But the codes are often applied to piping systems that are outside their scope.

This sound like it might be a big problem, but the intelligent application of a piping code

outside of its scope is not necessarily bad.

Of the seven B31 codes, most design engineers find that they spend most of their time

dealing with B31.1, B31.3, and B31.9. The remainder is confined to an overview of

these codes.

7.1 Limits of stresses set by code ANSI / ASME B 31.3

Limits of Calculated stresses due to Occasional loads

ANSI / ASME B 31.3 in clause 302.3.6specifies that the sum of longitudinal stresses

due to pressure, weight and other sustained loadings and of the stresses due to

produced by occasional loads such as wind or earthquake, may be as much as 1.33

times the basic allowable stress. Wind and earthquake forces need not be considered


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as acting concurrently. When the piping system is tested, it is not necessary to consider

other occasional loads such as wind and earthquake as occurring concurrently with test

loads.

Limits of calculated stresses due to Sustained loads

ANSI / ASME B 31.3 in clause 302.3.5specifies that the sum of longitudinal stresses,

SL in any component in a piping system due to pressure, weight and other sustained

loadings shall not exceed the allowable stress at the design temperature. The thickness

of pipe used in calculating the SL shall be the nominal thickness less the allowable due

to corrosion and erosion.

Limits of Displacement stress range

ANSI / ASME B 31.3 limits the allowable stress range to 78% of the maximum stress

the system could be subjected to without flow occurring either in hot or cold condition.

7.2 ASME B31.1 power piping

Power piping in this case means the piping that is used around boilers. It is called

power piping because often a boiler is used to make steam for power generation. This is

done by converting pressure and temperature energy into kinetic energy in a turbine,

which then produces electrical energy.

Mechanical engineering is the technical field associated with transforming energy into

work. A steam generator is a perfect example of that. This code relates particularly to

piping that would be found in electrical power plants, commercial and institutional

plants, geothermal plants, and central heating and cooling plants.

This code is primarily concerned with the effects of temperature and pressure on the

piping components.

Scope


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Around a boiler, the scope of B31.1 begins where the boiler proper ends, at either:

1. The first circumferential weld joint

2. The face of the first flange

3. The first threaded joint

This piping is collectively referred to as boiler external piping, since it is not considered

part of the boiler.

Power piping may include steam, water, oil, gas, and air services. But it is not limited to

these, and as mentioned before, there is nothing that says you cannot apply B31.1 to

other piping systems unrelated to boilers or power generation, as long as they would not

be better classified as within other piping codes, which may be more stringent.

Not in Scope

B31.1 does NOT apply to:

1. Components covered by the ASME Boiler and Pressure Vessel Code.

2. Building heating and distribution steam and condensate piping if it designed for 15

psig or less.

3. Building heating and distribution hot water piping if it is designed for 30 psi or less.

4. Piping for hydraulic or pneumatic tools (and all of their components downstream of

the first block valve off of the system header).

5. Piping for marine or other installations under federal control

6. Structural components

7. Tanks

8. Mechanical equipment

9. Instrumentation


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Pressure Design of Straight Pipe

This section contains several extremely useful formulas for determining either the

design pressure of a particular pipe or the required wall thickness of a pipe operating at

a certain pressure. These formulas are:

Where

tm = Minimum required wall thickness[in. or mm]

P = Internal design gage pressure [psi or kPa]

The pressure is either given or solved for in the equations.

Do = Outside diameter of the pipe [in or mm]

The outside diameter will be the OD of a commercially available pipe.

d = Inside diameter of the pipe [in or mm]

The inside diameter will be the ID of a commercially available pipe.

S = Maximum allowable stress values in tension for the material at the design

temperature [psi or kPa]

E = Weld joint efficiency, shown in ASME B31.1, Appendix A. These values

depend on the material used and the method of manufacture. Naturally, if the


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material is a cast product, there is no weld. In that case the casting factor F is

used.

F = Casting factor shown in ASME B31.1, Appendix A. Where a cast material is

used .

A = Additional thickness [in or mm].

Piping is generally purchased based on commercially available schedules or wall

thicknesses (unless specially ordered, which is usually prohibitively expensive). These

thicknesses must take into account the mill tolerance, which may be as much as 12.5

percent less than the nominal thickness.

These values are tabulated in ASME B31.1, Appendix A. Note that they are dependent

on the temperature to which the material will be exposed. This temperature is the metal

temperature. This would normally be the temperature of the fluid in the pipe, but if a

pipe was to be exposed to a high temperature externally, it would be the fluid

temperature outside the pipe.

Note that the values tabulated in Appendix A include the Weld Joint Efficiencies and the

Casting Factors. Therefore, the tabulated values are the values of S, SE, or SF. See

General Note (f) at the end of each table.


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The power piping Code ANSI B 31.1 specifies that the developed stresses due to

the sustained, occasional and expansion stresses be calculated in the following

manner.

Limits for Sustained and Displacement Stresses

This section addresses cyclic stresses among other things.

Note that the Stress Range Reduction Factors apply only to thermal cycling and NOT to

pressure cycles.

where,


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Ss = Sustained stress.

i = Stress Intensification factor.

MA = Resultant moment due to primary loads

= ( Mx + My + Mz ) 0.5

Sh = Basic allowable stress at the operating temperature

Z = Section modulus.

Limits for occasional stresses

where,

So = Occasional stress.

MB = Resultant range of moments due to occasional loads

= ( Mx + My + Mz ) 0.5

K = Occasional load factor

= 1.2 for loads occurring less than 1% of the time

= 1.15 for loads occurring less than 10% of the time.

Limits for expansion

Where,

Mc = Resultant range of moments due to Expansion (secondary) loads

= ( Mx + My + Mz ) 0.5

SA = Allowable expansion stress range


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104.3 Branch Connections

Branch connections are often made using fittings designed for the application. Such

fittings are manufactured according to the standards listed in ASME B31.1, Table 126.1.

But branch connections are often made in other ways. Especially where large bore

piping is concerned, some branch connections are made without the use of a

manufactured fitting. It might be helpful now to make a distinction between

manufactured and fabricated. A manufactured fitting is one that could be purchased

from a supplier. The supplier would sell it to you just as he received it from a factory. It

would be (or should be) made to a specification; perhaps one of the specifications listed

in ASME B31.1s Table 126.1. A fabricated fitting would be made in a shop, or in the

field, with pieces of pipe and/or plate.

There are a few reasons why it may be advantageous to use a fabricated fitting instead

of a manufactured fitting:

Lack of availability of the specific fitting. For instance, a 20 in 304SS lateral may not be

available.

The cost of the manufactured fitting is very high compared to the ability to fabricate it.

The cost of making the welds on a manufactured fitting is higher than on a fabricated

fitting

Consider an 18 diameter run of piping, and a 14 in diameter branch connection. There

are several ways that this could be accomplished.

One method might be to insert a full-size butt-welding tee, with an 18 in x 14 in reducer

exiting the branch, as shown in Figure 10.

Another method might be to use a reducing tee as shown in Figure 4.2.

Figure 11shows still a third method, in which the 14 in branch pipe is stubbed directly

into the 18 in run pipe. The type of cuts required to join pipes like this is called a

fishmouth by pipefitters.


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Figure 10: Branch connection using full size tee and reducing tee

Figure 11: A branch connection using a nozzle weld without a fitting.


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Obviously, the types of fabrications one could concoct are limited only by ones

imagination. In order to achieve safe designs that can withstand the same pressures as

other pipe configurations, these fabrications are subject to the requirements of this code

section.

Whenever a run pipe as shown in Figure 11is cut to accommodate a branch connection,

the strength of the run pipe is compromised due to the material that is removed. The

larger the branch the more material removed, and the worse the situation.

If a manufactured fitting is chosen (for example in Figure 10) to provide the branch

connection, and if the manufacturers specification is listed in Table 126.1 in the ASME

B31.1, then no further analysis is required. A situation similar to Figure 11 above might

require additional engineering analysis, and Section 104.3 provides the guidance to

perform that analysis.

The first thing to note is that fittings manufactured in accordance with the ASTM

standards listed in ASMEs Table 126.1 are satisfactory in meeting the code (Refer to

Section 104.3.1 [B.1]). For example, if your specification designated that all tees had to

meet ASTM A234 Piping Fittings of Wrought Carbon Steel and Alloy Steel for Moderate

and Elevated Temperatures, then you would be covered. No further qualifications

would be necessary for those tees (other than to establish that they were installed in the

piping system correctly5).

If, on the other hand, you decided that you really didnt want to spend the money on an

18 in diameter tee manufactured fitting, but would rather fabricate the fitting, then in

order to satisfy the code, you would have to follow the requirement set forth in Section

104.3.1(D). These requirements specify the extent to which branch connections must be

reinforced.

In practice, it is unusual to use a fabricated fitting when the branch diameter is the same

size as the run diameter. The reason for this is the cut and weld for the branch would

have to extend to the centerline of the pipe (halfway around the circumference). This

would require extensive cutting, welding, and reinforcement, and would not be an


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economical choice. Most pipe specs indicate that full-size branch connections be made

with a manufactured fitting (a tee).

For a single size reduction of the branch pipe, a reducing tee is often used. Below one

size reduction, the branches are most often fabricated until the branch size is smallenough that a manufactured welding fitting such as a Weld-o-let may be used. These

fittings are described in 104.3.1 (B.2). We will discuss these connections later in the

chapter covering fittings.

Extruded outlets are described in Sections 104.3.1 (B.3) and 104.3.1 (G). They are

manufactured by pulling a die through the wall of a pipe. Due to the custom nature of

these fittings, they are unusual in general industrial applications, but can be found in the

power industry.

This section pertains to branch connections where the axes of the main run and the

branch intersect, and the angle formed by the main run of the pipe and the branch is

between 45 and 90.

If both of these conditions do not exist, additional tests or analysis must be performed to

ensure that adequate strength is provided.

Section 104.3.1 (D) relates to branch connections subject to internal pressure. The code

designates a region surrounding the intersection known as the reinforcement zone.

The reinforcement zone bounds the region of concern at the branch with a

parallelogram. All of the analysis is confined to this zone, and any required

reinforcement must fall within this zone.

Imagine a plane passing through the intersecting axes of the branch connection. See

Figure 12. The discussion of areas in this code section refers to the cross-sectional

areas that appear in this imaginary section.


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Figure 12: ASME B31.1-2001 figure 104.3.1 (D) Example B, showing the various reinforcement areas for abranch connection.

The area of the material that is removed when the hole is cut in the run pipe must be

offset by material that is present in other components within the reinforcement zone.


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Credit is given for what is referred to as excess pipe wall. Consider that piping

systems are rarely operated at the maximum design pressure calculated by Formula (4)

in Section 104.1.2. For one thing, the piping used is the commercially available pipe

wall. This means that there is usually some inherent excess of pipe wall available in

either the run pipe or the branch pipe or both. The code allows you to take this excess

pipe wall into account when determining the need for additional reinforcement. This

excess pipe wall would be the difference between the nominal pipe wall minus the mill

tolerance, minus any additional thickness allowance, minus the wall thickness required

by Formula (3) or (3A) in paragraph 104.1.2(A). In other words,

Excess pipe wall = (nominal wall thickness x 0.875) - (corrosion, erosion, threading, or

grooving allowance) - (tm, as calculated by Formula (3) or (3A))

The 0.875 term accounts for the Mill tolerance (12.5 percent).

Therefore, one place that can make up the amount of material removed by the hole in

the run pipe is any excess pipe wall present in the reinforcement zone. This is

designated A1 for the run pipe (also known as the header) and A2 for the branch pipe.

Another source of excess material is the area of the fillet weld, designated A3.

Sometimes a reinforcing pad (or re-pad) is placed around the branch connection toadd strength to the joint. The ratio of the width of the re-pad to its height should be as

close as possible to 4:1 (within the limits of the reinforcing zone). It should never be less

than 1:1. This materials area is designated A4.

Still another method of reinforcing a branch connection is to weld on a saddle. These

are limited to use on 90 branches. Their use in general industry is not as common as

other methods of preparing branch connections, but they remain a viable alternative.

The metal contained in the saddle along the run pipe in the reinforcement zone

constitutes the additional metal that may be used to offset the material lost in cutting the

hole in the run pipe. The area of the metal in the saddle along the run pipe is designated

A5.


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So there are a total of five areas that can be added together to offset the loss of material

created by the hole in the run pipe, which is designated as A7. Since the pipe is

expected to retain its integrity throughout its design life, the wall thickness expected at

the end of the pipes design life is the thickness that must be used in the calculations.

The newer versions of the code call this pressure design area at the end of the service

life A6.

Where

tmh= The required minimum wall thickness in the header pipe

A = the additional thickness added to account for corrosion, erosion, grooving or

threading

Where

= angle between the axes of the run and branch pipes

Another way of stating this is that the required reinforcement area must be less than any

combination of:

1- A1 = Area of any excess pipe wall contained in the run

= 2- A2 = Area of an excess pipe wall contained in the branch.

= 3- A3 = Area of any welds beyond the outside diameters of either pipes or of weld

attachments of pads, rings, or saddles.

4- A4 = Area provided by any rings, pads, or integral reinforcement.


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5- A5 = Area provided by a saddle on a right angle connection.

= In practice, if you were using a saddle, you would not also have a re-pad, and vice-

versa. Therefore, A4 and A5 can be considered mutually exclusive.

The code lists specific requirements for closely spaced branch connections, branch

connections subject to external forces and moments.

The Pipe Fabrication Institute publishes worksheets (designated ES36) that aid in these

calculations.

Miters

Miters are perfectly acceptable fabricated fittings for pressure piping, if constructed in

accordance with the requirements of Paragraph 104.3.3. Note however that they are

usually only used in large bore piping where manufactured elbows are either

unavailable or very expensive. Miters require much fit-up and welding. It is easier to

simply purchase an elbow that conforms to one of the standards listed in Table 126.1 if

these are available.

137 Pressure Tests

After a pipe system is installed in the field, it is usually pressure tested to ensure that

there are no leaks. Once a system is in operation, it is difficult, if not impossible, to

repair leaks.

ASME B31.1 has established procedures for applying pressure tests to piping systems.

There are generally two types of pressure tests applied to a piping system. One is a

hydrotest and the other is a pneumatic test. The hydrotest is greatly preferred for the

following reasons:

Leaks are easier to locate.

A hydrotest will lose pressure more quickly than a pneumatic test if leaks are

present.


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Pneumatic tests are more dangerous, due to the stored pressure energy and

possibility of rapid expansion should a failure occur.

On the other hand, if a piping system cannot tolerate trace levels of the testing medium

(for instance, a medical oxygen system) then a pneumatic test is preferred.

137.4 Hydrostatic Testing

It is important to provide high point vents and low point drains in all piping systems to be

hydrotested. The high point vents are to permit the venting of air, which if trapped during

the hydrotest may result in fluctuating pressure levels during the test period. The drains

are to allow the piping to be emptied of the test medium prior to filling with the operating

fluid. (Low point drains are always a good idea though since they facilitate cleaning and

maintenance.)

A hydrotest is to be held at a test pressure not less than 1.5 times the design pressure.

The system should be able to hold the test pressure for at least 10 minutes, after which

the pressure may be reduced to the design pressure while the system is examined for

leaks. A test gauge should be sensitive enough to measure any loss of pressure due to

leaks, especially if portions of the system are not visible for inspection.

The test medium for a hydrotest is usually clean water, unless another fluid is specified

by the Owner. Care must be taken to select a medium that minimizes corrosion.

137.5 Pneumatic Testing

The test medium must be nonflammable and nontoxic. It is most often compressed air,

but may also be nitrogen, especially for fuel gases or oxygen service. Note that

compressed air often contains both oil and water, so care must be exercised in

specifying an appropriate test medium.

A preliminary pneumatic test is often applied, holding the test pressure at 25 psig to

locate leaks prior to testing at the test pressure. The test pressure for pneumatic tests is

to be at least 1.2 but not more than 1.5 times the design pressure. The pneumatic test

must be held at least 10 minutes, after which time it must be reduced to the lower of the

design pressure or 100 psig (700 kPa gage) until an inspection for leaks is conducted.


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If a high degree of sensitivity is required, other tests are available such as mass-

spectrometer or halide tests.

7.3 ASME B31.3 Process Piping

The term process piping may be thought of as any piping that does not fall under the

other B31 codes. It is generally considered to be the piping that one may find in

chemical plants, refineries, paper mills, and other manufacturing plants.

This code is structured similar to B31.1 in that it is organized into chapters, parts, and

paragraphs. Note that while the paragraphs of B31.1 are numbered in the 100s, those in

B31.3 are numbered in the 300s. This convention follows throughout the B31 codes.

There are several very important concepts in this code that should be identified before

we delve too far into the particulars. Because we have entered into the realm of process

piping, it is necessary to recognize some of the inherent hazards associated with

handling dangerous chemicals.

Scope

The scope of this code includes all fluids. This scope specifically excludes the following:

1. Piping with an internal design pressure between 0 and 15 psi (105 kPa)

2. Power boilers and BEP which is required to be in accordance with B31.1

3. Tubes inside fired heaters

4. Pressure vessels, heat exchangers, pumps, or compressors.

Definitions

There are several very important definitions included in this paragraph under the term

fluid service:

a) Category D fluids are those in which all of the following apply:

1. The fluid is nonflammable, nontoxic, and not damaging to human tissue.


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2. The design pressure does not exceed 150 psig (1035 kPa).

3. The design temperature is between -20F and 366F (-29C and 186C).

b) Category M fluids are those in which a single exposure to a very small quantity could

lead to serious irreversible harm, even if prompt restorative measures are taken.

c) High pressure fluids are those in which the Owner has specified that the pressures

will be in excess of that allowed by the ASME B16.5 PN420 (Class 2500) rating for the

specified design temperature and material group.

d) Normal fluids are everything else that does not fit into the above categories. These

are the fluids most often used with this code.

Design Conditions

This section requires the designer to consider the various temperatures, pressures, and

loads that the piping system may be subject to. While it is a good checklist, most of the

items contained are common sense.

Uninsulated Components

This paragraph describes how to determine the design temperature of uninsulated

piping and components. Of particular interest is the description of how to determine

component temperatures using the fluid temperature? The instructions indicate that for

fluid temperatures above 150F (65C) the temperature for uninsulated components

shall be no less than a certain percentage of the fluid temperature. For example, the

temperature used for lap joint flanges shall be 85 percent of the fluid temperature.

Note that unless you use the absolute temperature in degrees Rankine or Kelvin, such a

calculation has no meaning, since a percentage cannot be applied to the Fahrenheit or

Celsius scales.

Design Criteria

Note that B31.3 also has a Table 326.1 that corresponds to B31.1s Table 126.1. A

comparison between the two tables shows that Table 126.1 is focused more on steel


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pipe and fittings, while Table 326.1 pertains more to nonmetallic pipe and fittings. The

obvious reason is that process piping deals with more fluids that are corrosive to steel.

In many cases, thermoplastics, thermosetting plastics, and resins will be more

appropriate materials for the fluids handled in the purview of the process piping code.

This set of paragraphs states that if the components listed in Table 326.1 are rated for a

specific temperature/pressure condition, then they are suitable for the design pressures

and temperatures allowed by this code. If they have no specific temperature/pressure

rating, but are instead based on the ratings of straight seamless pipe, then the

component must be de-rated by 12.5 percent, less any mechanical and corrosion

allowances.

In other words, you have to determine the minimum wall thickness of the straight pipe

based on the design temperature and pressure, as well as the mechanical and

corrosion allowances. Once you apply the mill tolerance of 12.5 percent, you will be

safe in selecting a fitting that satisfies the same requirements as the straight pipe to

which it is connected.

Allowances for Pressure and Temperature Variations

There are paragraphs in both B31.1 and B31.3 that describe allowable deviations from

operating conditions. These are called allowances for pressure and temperature

variations. The rules for such operating excursions are not complicated, but inpractice

industrial users do not chart how often the operating pressures exceed the allowable

pressures.

Most often, any pressure excursions are prevented through the use of pressure relief

devices, such as pressure relief valves, pressure safety valves, or rupture disks.

Also, it is important to note that the allowable stresses are temperature dependent. So if

there are temperature excursions (as allowed for in both B31.1 and B31.3) the allowable

stress may vary. Unless someone has taken the trouble to build a database of the

relationships between operating temperature and pressure, and allowable temperature

and pressure, then the designer will be well-advised to base the design pressure on the


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MAXIMUM temperature that the system will ever see, and not to rely on the allowance

for temperature variations.

Therefore, from a practical standpoint, it is best to not rely upon any allowances for

temperature or pressure excursions above the design conditions. Choose your designconditions so that the temperature and pressure will not be exceeded.

Pressure Design of Straight Pipe

As we noted above, we are required to calculate the required wall thickness to satisfy

the design temperature and pressure conditions. We did the same thing for B31.1. But

B31.3 handles things a little differently.

where

tm = Minimum required wall thickness [in or mm].

This minimum wall thickness includes any mechanical, corrosion, or erosion

allowances.

If the piping system contains bends (not elbows), then you also must compensate

for thinning of the bends, as in ASME B31.1. Because this is not common, the

interested reader is referred to ASME B31.3 Paragraph 304.2.1 for the formulas

required to determine after-bend thicknesses

t = Pressure design thickness, as determined by any of the Formulas (3a)

through (3b) [in or mm].


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c = Mechanical, corrosion, or erosion allowances [in or mm]. Note that for

unspecified tolerances on thread or groove depth, the code specifies that an

additional 0.02 in (0.5 mm) shall be added to the depth of the cut to take the

unspecified tolerance into account.

T = Pipe wall thickness, either measured or minimum per purchase specification

[in or mm].

Unless specially ordered (which is usually prohibitively expensive) piping is

generally purchased based on commercially available schedules (or wall

thicknesses). These thicknesses must take into account the mill tolerance which

may be as much as 12.5 percent less than the nominal thickness.

Therefore, under ordinary circumstances, the pipe wall thickness (T) will be 87.5

percent of the thickness of the listed schedule.

d = Inside diameter of pipe [in or mm].

P = Internal design gage pressure [psig or kPa (gage)]

The pressure is either given, or solved for in the equations.

D = Outside diameter of pipe [in or mm]

The outside diameter will be the OD of a commercially available pipe. Carbon

steel pipe dimensions are shown in Appendix 1 of this text.

E = Quality Factor from ASME Table A-1A or A-1B. Table A-1A relates

exclusively to castings. Table A-1B relates to longitudinal weld joints. The quality

factor is a means of de-rating the pressure based on the material and method of

manufacture. Thus, for A106 seamless pipe, the quality factor E = 1.00.

Casting quality factors may be increased if the procedures and inspections listed

in ASME B31.3 Table 302.3.3C are utilized.

The quality factors are in place to account for imperfections in castings, such as

inclusions and voids. Machining all of the surfaces of a casting to a finish of 250


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micro inches (6.3 m) improves the effectiveness of surface examinations such

as magnetic particle, liquid penetrant, or ultrasonic examinations.

The Quality Factor E is analogous to the Weld Joint Efficiency E or Casting

Factor F in B31.1. But note that the Quality Factor E in B31.3 is NOT included inthe stress values provided in B31.3 Tables A-1 and A-2. See Paragraph

302.3.1(a).

S = Stress in material at the design temperature [psi or kPa].

These values are tabulated in ASME B31.3, Appendix A. Note that they are dependent

on the temperature to which the material will be exposed. This temperature is the metal

temperature. This would normally be the temperature of the fluid in the pipe, but if a

pipe was to be exposed to a high temperature externally, it would be the fluid

temperature outside the pipe. See also Paragrah 301.

Once again, note that the values tabulated in ASME B31.3 Appendix A DO NOT include

the Quality Factors. Therefore, the tabulated values are only the values of S.

W = Weld Joint Strength Factor. This factor accounts for the long term strength of weld

joints at elevated temperatures. In the absence of specific data such as creep testing, W

is taken as 1.0 at temperatures of 950F (510C) and below. W falls linearly to 0.5 at

1500F (815C).

Y = A coefficient used to account for material creep, as in B31.1. The table of Y

coefficients in B31.3 is virtually identical to the table given in B31.1. As previously

noted, the variation of Y with temperature allows the wall thickness equation to behave

in accordance to the Modified Lam Equation at low temperatures (with Y = 0.4), and

in accordance with a creep-rupture equation at high temperatures (with Y = 0.7).

The values are taken from Table 304.1.1 for t < D/6. For ductile metals (including steel),

the value is 0.4 across the range of temperatures. For t >= D/6,


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Note that the difference between Formulas (3a) and (3b) is that (3a) begins with the OD

of the pipe, and (3b) begins with the ID of the pipe.

B31.3 Chapter VII deals with nonmetallic piping and piping lined with nonmetals.

Paragraph A304 explains how to calculate minimum wall thicknesses in such cases.The method and equations closely parallel Paragraph 304.1.1. The Allowable Stresses

are replaced with Hydrostatic Design Stresses for nonmetals in Table B-1.

Pressure design of high-pressure piping (pressures in excess of the Class 2500 rating

for the design temperature and material group), is covered in Paragraph K304. The

equations given use Table K-1 for the basic allowable stresses. These stresses are

higher than those listed in Table A-1 for the same materials.

Limits for sustained stresses

where,

FAX = Axial force due to sustained ( primary ) loading

Mi = In-plane loading moment due to sustained ( primary )

Mo = Out-plane loading moment due to sustained ( primary ) loading.

ii , io = in-plane and outplane stress intensification factors.

Sh = Basic allowable stress at operating temperature.

Limits for occasional stresses

The code states that in order to calculate the stresses due to sustained and occasional

loads independently as per the above equation and then add them absolutely. The sum

should not exceed 1.33 Sh.


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Limits for Expansion

where,

SE = Expansion stress range

Mi = Range of in plane bending moment due to expansion (secondary) load

Mo= range of out plane bending moment due to expansion (secondary) load

MT = Range or torsional bending moment due to expansion load

SA = Allowable stress range.

304.3 Branch Connections

Similar to B31.1, B31.3 permits fabrication of branch connections. We are faced with the

following prerequisites:

1. The run pipe diameter-to-thickness ratio (Dh/Th) < 100 and the branch-to-run

diameter ratio (Db/Dh) is not greater than 1.0. If we examine the tables of commercially-

available pipe data, we see that the first condition in which we might see a diameter-to-

thickness ratio in excess of 100 would be 24 in diameter Schedule 5S, which has a wall

thickness of 0.218 in. For thicknesses above the standard wall thickness, there is little

chance that the ratio will exceed 100. So it is clear that while this is an important

consideration for large bore, thin wall pipes, it is a situation that most of us are not likely

to encounter. Looking next at the requirement that the branch-to-run ratio is not greater

than 1.0, we see that this means only that it is impossible to stub a larger branch onto a

smaller run. And if we tried to do that, we might be inclined to reverse the names of thebranch and run. In other words, the branch pipe is always the smaller diameter, unless

they are both the same diameter. Whichever pipe is designated as the run pipe, must

still satisfy (Dh/Th) < 100.


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2. If Dh/Th >= 100, the branch diameter Db has to be less than one-half the run

diameter Dh.

3. The angle between the branch and run is at least 45. See Figure 4.8. This is

analogous to B31.1s angle . We will examine the similarities and differences betweenB31.1 and B31.3 regarding the branch connection calculations.4. The axes of the branch and run pipe must intersect each other. This was also a

requirement of B31.1.

Once the prerequisites are established, we can examine any requirements for

reinforcing the branch connection.

Here is an example of how the codes complicate matters. The names assigned to the

various reinforcing areas in B31.3 are different than those we examined in B31.1. See

Table 4.1 for a comparison of the terminology as they pertain to branch connections.

For B31.3:


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The subscripts b and h refer to the branch and run pipe (or header pipe) respectively. In

order to satisfy the reinforcement requirements of B31.3,

A2 + A3 + A4 >= A1 Formula (6a)

As in B31.1, we have a reinforcement zone bounded by the parallelogram shown in

Figure 4.9. Also, as in B31.1, there are also specific requirements for closely spaced

nozzles (overlapping reinforcement zones) and branch connections subject to external

pressure, forces, or moments.

305.2 Specific Requirements

This section describes what piping may be used for certain services. A review of the

four fluid services reveals that the most benign service

Pipelines stress analysis Report [Mechatronics]

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