PIPE DESIGN FOR ROBUST SYSTEMS

PIPE DESIGN FOR ROBUST SYSTEMS

Stress analysis makes for a solid pipe system design. Software can help

After the piping and instrumentation diagram (P&ID) of a process design has been completed, the next step is stress analysis of the piping network. In past years, engineering or operating companies had staffs of layout and stress designers. Today, both these functions might be in the hands of the individual engineer. Either way, a solid understanding of pipe stress, and how to handle it, will lead to better plant design and more reliable operation.

Engineers may perform pipe stress analyses on a daily or occasional basis, or review the pipe stress analysis of others. Whatever their role, most of these individuals have only a basic understanding of the topic. And many may be unaware of the software tools that can automate many of the analytical steps (Box, p. 92).

When to analyze

One of the most difficult decisions plant personnel face is whether or not to analyze an existing or new piping system. It is often hard to assess the point at which a piping system should be field-routed or when a full analytical solution is required. Although there are no simple answers to these questions, here are some conditions under which piping analysis is advisable:

Piping is attached to load-sensitive equipment or is carrying Category M fluids (hazardous chemicals, as defined by ASME B31.3 rules), and the design temperature is greater than 250 oF.

The pressure exceeds the maximum pressure for an ANSI Class 2500 B16.5 fitting

The system temperature is greater than 400 degrees F

The system carries gas that has cooled to a liquid state

The product of the pipe outside diameter (in.) times the pressure (psi) is greater than or equal to 1,157

The system pressure is greater than 3,000 psi

The system uses Glass Reinforced Epoxy (GRE) pipe

The piping connects to rotating equipment

The system uses one or more expansion joints

Most piping codes, and perhaps 90% of all pipe stress analyses, involve three principal loading types:

(1) Sustained loads, such as pressure and weight;

(2) Expansion loads (i.e., from thermal conditions); and

(3) Occasional loads, such as from wind and earthquakes.

Other types of loads include those caused by transient fluids, ice and snow, ship or platform pitch and roll, explosion loadings, pressure loads, frost heave, fault movement, fluid sloshing and through-wall thermal effects. These can all be analyzed, but are typically reserved for more experienced pipe-stress analysts.

In general, an analysis involves satisfying code requirements for stress in the pipe, and manufacturers' requirements for loads on equipment, flanges, vessel nozzles and the like. Basic code compliance involves running a computer program to compute a stress, and making sure that it is less than a code calculated allowable, i.e.,

-- Calculated Stress less than Allowable Stress

Stresses in a piping system are calculated by the following general-purpose equation:

Stress = PD/4t + i (M/Z)

Where:

P = design pressure, psig

D = pipe diameter, in.

t = pipe thickness, in.

i = Stress Intensification Factor (taken from the piping code for the applicable geometry, i.e., bends or tees)

M = square root of the sum of the squares of the three moments acting at any point in the piping system. Sustained expansion or occasional loads can cause the moment

Z = section modulus of the pipe [approximated by (r2)t; (r) is the mid-surface radius of the pipe]

1. If pressure and weight cause the calculated bending moments (M), the allowable stress is Sh, where Sh is the hot allowable stress from the applicable piping code.

2. If the calculated bending moments are due to thermal expansion, the allowable stress SA is:

SA = f(1.25(Sc + Sh) - Sl)

Where:

f = cyclic reduction factor (approximated by 6N-0.2) and N = the number of full-range thermal loading cycles

Sc = the cold allowable stress from the piping code

Sh = the hot allowable stress from the piping code

Sl = the weight and pressure stress in the piping system at the point under study

3. If the calculated bending moments are due to pressure, weight, wind or earthquake loads, the allowable stress is equal to k(Sh), where k is a value given by the piping code. For example, in ASME B31.3, a common piping code reference, k = 1.33.

The above equations are simplifications of the actual rules given in the piping codes, and serve to illustrate the intent of the stress-evaluation procedure. When performing or reviewing a stress analysis, an engineer should check each of the following allowables to be sure of compliance:

Sustained (weight and pressure) loads -- to prevent the system from collapsing or from experiencing excessive distortion (stress less than Sh)

Expansion (thermal) load -- To avoid fatigue failure, when cracks form, grow and then fail catastrophically (stress less than f[1.25{Sc + Sh} - Sl)

Occasional (wind and earthquake) loads -- to prevent the system from collapsing or from experiencing excessive distortion (stress less than 1.33Sh)

Markl's landmark paper, ``Piping Flexibility Analysis,'' [1] establishes most of these rules. It covers a number of the assumptions used in the formulation of the code rules that are not found in any of the code documents themselves, and is an excellent basic reference on the subject.

Now, address pressure factors

Most piping system designs start well before the plant or unit is laid out. The operator and the engineering design firm agree on a piping specification to govern the basic pressure and material design of the piping system. Piping specifications are primarily designed to ensure that the piping system will withstand the expected pressure and is suited for the intended process. Piping specifications are typically written around an applicable piping code. The most common piping codes, arranged by nation, are:

United States (ASME):

B31.1 - Power

B31.3 - Petrochemical

B31.4 - Oil and Slurry Pipelines

B31.5 - Refrigeration

B31.8 - Gas Pipelines

B31.9 - Building Services Piping

Section III -Nuclear Piping: NB (Class 1); NC (Class 2); ND (Class 3)

Canada:

Z183 - Canadian Oil Pipelines

Z184 - Canadian Gas Pipelines

Z662 Oil and Gas Pipeline Systems

Europe:

BS806 - British Power Piping

FDBR - German Power Piping

Other relevant standards include the National Electrical Mfrs. Assn. SM 23; Standards of the Tubular Exchanger Mfrs. Assn.; and Bulletins of the American Petroleum Institute (API). The most widely used piping code is ASME (or American National Standards Institute, ANSI) B31.3, Chemical Plant and Petroleum Refinery Piping.

Chapter 2, Part 2 of B31.3 gives the minimum requirements for pressure in piping components. Piping specifications are designed around these Chapter 2, Part 2 pressure requirements. Typically, the guide's requirements are satisfied well before any stress analysis (Chapter 2, Part 5) is performed. However, analysts should not assume that a safe pipe-flexibility analysis means a safe pressure analysis.

Piping flexibility (or stress) analysis tests pressure to a small extent but is primarily concerned with weight, thermal and occasional loads. Most pipe stress programs do not satisfy these Part 2 requirements, since the programs are primarily based on a different part: Chapter 2, Part 5.

A novice analyst might incorrectly assume that a piping system satisfies B31.3 allowables for thermal weight and pressure because a pipe stress program indicates that B31.3 is satisfied. For shop- or field-fabricated intersections, engineers must design these for pressure in addition to performing the piping flexibility analyses.

This situation arises most often when a plant makes minor modifications to existing pipe work. The plant engineer responsible for the modification accepts the design of the tie-ins based only on a flexibility analysis, without checking Part 2 pressure requirements. This is a significant mistake. The most dangerous problems occur when the pipes being connected have large diameters with small wall thickness, or where the branch pipe centerline is not 90 deg to the run pipe. The engineer should perform simple hand calculations to ensure that the Part 2 requirements are met.

Software speeds analysis

Piping analysis software is available in a range of capabilities, from completing simple drawings to performing complex flow and stress studies (Table 2). Techniques such as finite element analysis have proven useful. The software can save engineer's time and effort on the system's analysis.

Among the software's analytic capabilities are: static and dynamic analysis to calculate displacement or support reactions; loads caused by time-dependent phenomena such as water hammer; code compliance; nonlinear reactions; and the integration of structural components in the piping design.

Once engineers decide to perform formal stress calculations, they should realize that they will spend much of their time inputting data or adjusting computer models of the piping system. Engineers often rush to meet a deadline or bring a quick end to the tedious task of inputting data. Errors made during input account for a majority of analysis problems. Often the software will produce results even if the engineer has input inaccurate data.

Although software usually has interactive graphical capabilities so that users can verify that the physical representation matches the input data, it will not identify all user errors. The most common types of errors made are:

Omitting data about weight of insulation, valves and operators

Incorrectly entering data about how pipes intersect

Oversimplifying or incorrectly entering piping boundary conditions

Ignoring or oversimplifying pipe shoes, guides, anchors and trunnion supports

Using program algorithms without questioning underlying assumptions, such as spring hanger design

To reduce errors, the pipe stress program user should follow these guidelines:

List and check the input. Don't assume the input is correct just because the program does not report fatal errors. The first time that you list and check an input, and don't find an error is when you can stop checking inputs for errors

Don't just review code stresses in the output report. Look at printed displacements (one of the factors leading to the stress determination), making sure that they are reasonable, and that the system is deflecting in an expected manner. Frictionless supports will occasionally allow for ridiculous horizontal movements

Check restraint reports closely. Make sure that excessive loads do not exist at steel support locations. Where large loads exist, decide if friction should be considered in the analysis, or if an error in the input is causing the loads. A close review of displacement and restraint load reports often reveals areas of the model that are incorrect

Check loads on all rotating equipment nozzles. To verify these, use the manufacturers' guidelines, or industry accepted guidelines

Check displacements and rotations at expansion joint ends. Make sure these displacements do not exceed manufacturers' limits for the number of design cycles

Review the graphic results closely. Look at displaced shapes to be sure that they make sense. Make sure that high stresses occur in the correct places.

Most stress-analysis programs follow an easily manipulated, trial-and-error procedure. Usually, engineers need only make a few iterations to adjust supports or design springs so that stresses are within allowables.Linking to rotating equipment

Usually the most difficult work for the stress analyst is meeting the requirements of rotating equipment and end connections. Pumps, turbines and compressors are very sensitive to loads from attached pipe, and a single application of an overload on a pump or turbine can result in leaking seals, failed shafts, excessive wear and vibration. A pipe having only a 4,000-psi stress (bending stress, not fluid pressure) can easily overload a pump nozzle.

To ensure compliance with rotating-equipment or end-connection codes, refer to NEMA SM23 (for steam turbines), API 610 (for pumps), and API 617 (for compressors). Each of these documents gives allowable loads for different sizes of piping inlets and outlets. Most computer programs include these equipment specifications as part of their standard evaluation procedures. In addition, some manufacturers of rotating equipment provide their own guidelines or use a multiplier of NEMA or API values. Use the manufacturer's information whenever it is available.

It is more difficult to find a line routing or support configuration that will satisfy rotating-equipment allowables as the temperature increases. Most rotating equipment codes have two types of allowables:

Individual nozzle limits provide standards for each individual product nozzle on the rotating equipment. For example, the inlet to a pump may have an axial allowable of 2,000 lb, while the outlet has an axial allowable load of 1,750 lb. Another approach to this analysis is to evaluate the suction side of the pump independently from the discharge side.

Resolved load limits ensure that forces and moments on the product nozzles do not act in combination to create excessive cumulative loads or moments on the equipment casing or base. The analyst must resolve all loads acting on all product nozzles attached to the equipment about a single point on the equipment body.

There are limits to the resolution of these loads. Since pipe analysts do not typically study inlet and outlet piping systems together, they must remember that loads from both lines must satisfy most equipment requirements. For example, use suction loads and discharge loads to evaluate resolved load limits for pumps.

When adjusting supports in the vicinity of rotating equipment it is important to remember that:

1. Supports should not require construction or installation tolerances beyond the capability of the shop or field. Do not add supports where steel is not readily accessible. Avoid moment restraints, and translational restraints requiring precisely set gaps

2. Properly designed, horizontal loads should not cause deflection in the attached steel, which negates the support's usefulness. Steel constructions designed to keep piping loads away from rotating equipment will transfer the load back to the rotating equipment if excessive deflection occurs. The pipe stress analyst must determine a tolerable amount of deflection of the steel construction

3. Evaluate the effect of friction at supports on large lines, or on highly loaded supports. Where friction is undesirable, the designer might use polymeric friction plates or metal sliding plates to reduce friction. Some analysts always use friction in their analyses; others never use it

4. Take the millwright's practices into account when designing springs, stanchions, and other supports near rotating equipment. Some spring designs produce unbalanced loads on equipment when the system is in a cold condition. Millwrights may improperly adjust or locate these springs in the cold condition so that the system is perfectly balanced. Typically, the designer did not intend for the balance to occur in the cold condition

Meeting at the end points

If a piping system does not attach to rotating equipment, it probably connects to a pressure vessel, a heat exchanger, or another piping system. In these cases, a stress increase will occur at the end or terminal point in the system. There, a geometric discontinuity exists due to the intersection of two cylindrical shapes meeting to form an opening.

While some analytical software takes these terminal points into account; others do not, so the designer should exercise care. Stresses at terminal points are often the governing forces in a design. Some software programs use standards from the Welding Research Council (WRC) Bulletins 107 or 297 to determine the terminal point stresses. Analysts whose software doesn't perform either of these calculations can use the WRC standards to manually compute stress at terminal points.

WRC 107 and 297 methods for stress and stiffness at nozzle connections are based on simplified deformation theories, adjusted by tests and experience. Because of their ready availability in many computer programs, these methods have been used for wide varieties of geometries, well outside of their original intended scope. Users should be cautious using them when the d/D ratio is greater than 0.5, or when the t/T ratio is one or less. In general, the WRC documents will be conservative for stress, although possibly by hundreds of percent.

WRC 297 is the only WRC document that predicts stiffnesses of nozzles in cylinders. WRC 297 can be nonconservative for stiffness. Some computer implementations of WRC 297 may predict stiffness values hundreds of percent too low. The worst errors occur for out-of-plane stiffnesses, and when the parameter (d/D)root(D/T) is greater than one. The WRC 297 user is encouraged to verify stiffnesses generated by comparing them to a finite element result, or to predictions given in ASME Sec. III NB3685.7.

Piping systems are attached to vessels either by welding or by flanges. Pipe designers use welding throughout an entire piping system when leaks cannot be tolerated under any circumstances. When small leakage can be tolerated, most pipe designers use flange joints because maintenance is easier.

Most flanged joints are as strong as the attached pipe. A good software program will enable the user to describe the properties of these attachments in cases where the joints are not as strong as the pipe. Some of these exceptions are systems that have:

Fiberglass-reinforced plastic flanges

Soft gaskets made with synthetic fiber

Ring-type joints

Flange components made of cast iron

Larger (greater than 12 in.) flanges (which are notoriously susceptible to leakage caused by external moments)

Flanges should be located in the piping system at points of small bending moments, when possible. ASME B16.5 gives pressure and temperature ratings for flanges of various materials. It is not uncommon to see bending moments and axial forces converted to equivalent pressure and a comparison to B16.5 made. This comparison provides a safety factor of approximately five to eight.

How plastic pipe differs

The use of plastic piping systems is increasing because of declining prices and improved corrosion resistance. Plastic piping systems differ from other systems when pressure or temperature is a concern. Plastic systems often expand as much or more due to pressure instead of temperature. Piping analysts should not assume that knowledge of steel piping systems transfers to a plastic system.

Plastic piping systems fail most often at flanges or at connections of pipe to thicker components, such as bends and tees. Many of these failures occur because of incorrect support or because of improper joint makeup. Others occur because engineers do not properly evaluate water-hammer events. Many manufacturers are developing their own solutions to these problems as more and more plastic systems are being used.

What fluids can do

Most pipe stress analysts are only concerned with a fluid's temperature and its specific gravity. These analysts usually leave issues like flow velocities, valve closure rates, pump trips, chemical decompositions or resonant acoustic vibrations to the specialist, who is called in only when an extraordinary problem is discovered.

As new software becomes available, the pipe stress analyst is becoming more involved in transient as well as steady-state fluid-flow problems. Newer fluid programs use the 3D structural mesh of the piping system to construct a 1D mesh of the fluid system.

Before the software's advent, the process engineer usually performed the fluid calculations. The mechanical engineer (pipe stress analyst) was rarely concerned with flowrates or pressure drops, since most piping systems are not sensitive to the slight pressure drops caused by the addition of an expansion loop or extra pipe and bends. Now, piping engineers can involve themselves in the varieties of fluid problems because they have tools that will minimize the time needed.

Pipe stress engineers should be most concerned when fluid pressure produces mechanical loads on the piping systems. This classical situation can happen when:

A valve opens or closes quickly

Two-phase flow occurs

Notable flow oscillations exist

There are long liquid lines

There is piping in and around compressors and pressure-relieving devices

Spring hangers -- hot and cold

Spring hangers are important in piping systems because they support the weight of hot piping as it thermally expands. If a rigid support is used in place of a spring hanger, it will exert a force back on the pipe, creating a stress on the pipe or a load on an equipment nozzle when the pipe expands.

Because the spring hanger is flexible, it moves with the pipe as it expands. The spring load will change slightly from this movement, but this load change will not be detrimental to the piping system, and creates significantly less load than a rigid support.

Although a design with spring hangers may be accurate on paper, improper installation can defeat its usefulness. It is not at all unusual to walk through a facility noticing springs that are not carrying loads or that are bottomed out. Piping engineers need to provide specific instructions with any field-adjusted equipment, such as a load flange or a turnbuckle on a spring support.

Engineers often misapply constant-effort springs and place an excess number of these springs where they are not needed. They often make these mistakes when they are concerned about temperature, and thus support the piping system over a large part of its length using springs.

In these cases, engineers don't realize that once a line turns in the horizontal direction, considerable flexibility is gained by each pipe run between supports. There is a point in a horizontal run of pipe where entering horizontal runs will absorb any entry or exit vertical thermal expansion. Intermediate locations simply do not require springs.

Even though springs are designed to carry weight through a specified distance, the proper use of springs immediately adjacent to equipment nozzles solves many rotating equipment problems. In this case, the spring carries the specified weight exactly, even though there is little movement in the support. Rigid supports around hot rotating equipment -- if not properly installed -- can cause significant problems. The designer should exercise considerable caution when analyzing and designing these installations.

If a designer fails to incorporate supports with axial or horizontal stops for hot piping systems, the pipes may sometimes move slowly around on the pipe rack, contacting other pipes or steel. Between two fixed anchors, most piping systems have a location where zero thermal movements occur in one or more directions. Engineers should place supports at these thermal nodes in the computer model to provide fixed stability to the system and to prevent the piping from moving.

For a hot pipe that cycles, it is best to install guides and limit stops wherever possible. Use long pipe shoes or other supports wherever excessive axial displacement is expected. (It is very embarrassing to have a 12-in. pipe shoe slide off a steel support!)

Cold spring is the practice of purposely fabricating certain sections of piping to be either too short or too long. Cold spring will never improve a piping system's response to thermal stress (for B31 code applications), but it can reduce operating loads on rotating equipment, and can reduce creep problems caused by hot stresses.

Some operating companies do not permit cold spring to be used in a piping system design. Cold spring can overload equipment in the cold condition and engineers can easily forget about it years later when they make modifications. In general, it is best to leave cold spring out of a piping system calculation unless nozzle or support loads cannot be reduced any other way.

Nozzle stiffness is the degree of flexibility at terminal points in vessels and heat exchangers. It permits reduced piping loads at these locations because of an inherent flexibility in the shell of the vessel or heat exchanger. It can also significantly affect force and moment distributions in large-diameter, thin-walled systems or in short, stiff systems that have no inherent flexibility.

Users employing nozzle stiffnesses to reduce end-connection loadings should be aware that some implementations can produce inaccurate results, comparable to the situation described on p. 90. If engineers use nozzle stiffness to produce a significant change in final results, they should compare the stiffness values to an alternative source. The most readily available information on nozzle flexibilities comes from: WRC 297, NB3685.7, and some finite-element analysis programs.

Expansion joints are flexible portions of a piping system for reducing loads or absorbing thermal expansion. Engineers can design expansion joints in a variety of ways and can change their function to take lateral or axial loads. Expansion joints are often the most critical part of a piping system. Extra care should be taken when using expansion joints in important piping applications. Most expansion joint manufacturers are very helpful when it comes to providing design guidance and recommendations for use of their products.Hot `n' rusty

The effects of corrosion and high temperature on pipe stress analysis is among the least quantitatively understood of all stress related effects. Tests on piping fittings usually are done on essentially non-corroded fittings. The effect of corrosion on fatigue varies widely and may be significant. There is no quantitative way to evaluate these effects today.

When a system is subject to intensely corrosive conditions, and undergoes a significant number (several thousand) of thermal or pressure loading cycles, extreme caution should be used in the design. If a particular pipe material and chemical medium has not been previously used, metallurgical study of the interaction of corrosion and fatigue for the particular combination of fluid and medium may be warranted.

In most pipe stress programs today the corrosion allowance specified is removed uniformly from the wall thickness of the pipe before stress calculations are made. Depending on the piping code employed and the particular program options used, corrosion may not be removed, may be removed from sustained stress cases only, operating cases only, thermal cases only, or from any combination of the load cases. It is a wise user who questions the application of corrosion in a particular software program. It is the prophet who knows what to do with the answer.

High temperature in piping implies systems that operate in the creep range for the material. Material creep involves a complicated nonlinear interaction between temperature, stress and time. High-temperature rules for piping and pressure vessels are based on only the simplest of material laws.

Creep phenomena usually become significant at about 750 degrees F for steel and its alloys. High-temperature piping should be well supported and well controlled. Springs should be used carefully, and travel stops provided to limit excessive displacement.

Many high-pressure pipe spools are delivered in thicknesses that vary considerably from the nominal specified. This altered thickness affects the total weight of the system which is not included properly in a spring hanger design. The spring then pushes too much or too little on the pipe resulting in creep concentrations at critical bending sections. Supports on high-temperatures piping systems should be designed very carefully.

Piping stress analysis can be a complicated, dangerous business, and should be approached with great caution. Like any other technology, piping stress analysis software is only as good as the person who uses it. While the programs perform calculations and give guidelines, engineers must apply common sense and judgement when using these tools.

TABLE 1. Adapted from ASME B16.5, this table provides guidance on the allowable pressure rating for flanges and fittings based on metal type (material group), temperature and size

TABLE 1.PRESSURE-TEMPERATURE RATINGS FOR CLASS 150 PIPING Material

Group No. 1.2 1.4 1.5 2.1 2.4 2.7

Carbon Alloy Austenitic

steels steels steels

C-1/2Mo Type304 Type321 Type310

Temp., F

pressure, psig

-20 to

100 290 235 265 275 275 260

200 260 215 260 235 235 230

500 170 170 170 170 170 170

750 95 95 95 95 95 95

1,000 20 20 20 20 20 20

TABLE 2. READING PIPE STRESS PROGRAM OUTPUT REPORTS

Output Discussion: The three most common types of reports printed from pipe stress program are shown above. The deflection report shows the displacements of the system in the selected right-handed cartesian coordinate system. Y or Z is typically vertical, and the other axes are selected by the analyst to ease the input of data. The forces and moments on restraints show the loads the piping exerts on supports. The stress report shows how close the system stresses are to the applicable Code allowable.Load :

Dead Weight + Pressure 1 + Thermal 1

System Deflections

Point Displacements/in. Rotations/degree

Name X Y Z X Y Z

5 0.000 0.000 0.000 0.000 0.000 0.000

8 0.000 0.001 0.013 -0.001 0.003 0.005

10a 0.000 0.001 0.013 -0.001 0.003 0.005

10b 0.039 -0.016 -0.001 0.152 0.210 0.118

15 0.039 -0.016 -0.001 0.152 0.210 0.118

Deflection:

1) Verify displacements in the vicinity of all supports, particularly around supports with friction and gaps, and directional characteristics, i.e., +Y supports.

2) Check large horizontal displacements and be sure that they are in the correct direction and of approximately the right magnitude.

Forces and Moments acting on Restraints

Point Forces/lb Moments/in.-lb

Name X Y Z X Y Z

5 -418 3233 -2,909 -21,459 19,609 32,937

40 0 -6,075 0 0 0 0

75 -222 435 2,736 -30,408 -26,186 4,824

Force on Restraints:

1) Make sure that loads on terminal points are reasonable:

a) Will the support carry this much load without failing?

b) Will the support deflect under load, distributing the load somewhere else?

c) Is the support ``rigid,'' or does it have a realistic stiffness?

2) Check large horizontal loads and be sure that steel supports will carry large loads.

3) At locations of large loads, make sure that the inclusion of friction will not significantly alter the solution.

System Stresses (ASME B31.1)

In- Out-

Point Plane Plane Section Stresses/

Name SIF SIF Modulus psi Longitu. Princ. Code Allow.

5 1.00 1.00 22.03 1,401 1,589 2,248 3,164 43,750

8 1.00 1.00 22.03 1,401 1,389 2,143 2,670 43,750

8 6.25 6.25 22.03 1,401 7,267 7,360 12,831 43,750

System Stresses:

1) Check the stresses at terminal points. There, the stress intensification factors (SIFs) are 1.0. If a terminal point is another pipe or a vessel the SIF will not be 1.0.

2) If stresses are close to allowables, be sure that other safety concerns are null, i.e.:

a) Does the system cycle more than 7,000 times in its life?

b) Is the system heavily corroded?

c) Is the system temperature greater than 700 degrees F.

d) Is the D/T for the system greater than 80?

e) Is there any possibility of dynamic loading due to fluids?

TABLE 3. NOZZLE LOADINGS FROM ROTATING EQUIPMENT

Pumps and other devices put a stress (moment) on the interface between the device and piping. The following table, based on API Standard 610 (Table 2), provides stress estimates based on nozzle flange diameters

Edited by Nicholas Basta - Chemical Engineering

By Michael Bussler, P.E., is the founder and president of Algor, Inc and Anthony W. Paulin is president of Paulin Research Group

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