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Piping designers specify supports for piping systems based on the design conditions specified prior to construction. During construction and years of operation, many of the initial design specs may change, such as operating temperatures, pressures and flow rates. Periodic inspection is necessary to evaluate the status of the individual pipe supports and overall support of the piping system. Safety of the system is an important consideration when considering replacement of individual pipe supports or recalibration of the entire support system.

It is important to understand the original design of the piping system and the function of each individual support prior to inspection. A study of the data sheets, drawings and specifications should establish a picture of how the designer expected the system to operate. Once this is done, then the inspector can focus on what is different at the time of inspection.

Much can be learned through the analysis of data collected by careful inspection of all the supports in the system. Two questions need to be considered:

• Do the individual supports still function as designed?
• Does the system of supports still function as designed?

In addition to the physical integrity of the support assembly, the inspector must check the position (amount of compression) of the spring coil. The original designer specified the size for the variable spring support at a certain point based on the load and travel of the pipe expected during start-up, operation, and shutdown cycles of the system. The resistance to movement that a spring coil provides depends on how compressed the coil is. If the indicator shows that the coil is not compressed to the proper position, this means something is different from the what the original piping designer had intended.

When the positions of all the spring coils have been observed, an analysis of the current status of the system can begin. One of the problems which can occur is a bottom out spring. This means the coil is completely compressed and can only move in one direction. This is a major concern for the safety of current operations. An analysis of the total system is necessary to determine if factory or replacement is required or if field adjustments can restore the support system to safely function at current conditions. There are consulting companies who specialize in providing this inspection and analysis service. They also provide training which can be integrated with your safety program.

Checklist for Inspection of Hanger Assembly

• Check the beam attachment (1), pin (2) and the spring hanger attachment (3) for any cracks, fractures, or signs of corrosion.
• Check the pipe clamp attachment (4), the weldless eye (5) and threaded rod (6) for integrity.
• Test the turnbuckle (7), locknuts (8), and other threaded items to be sure they will turn.
• Check the spring coil (9) for any sign of corrosion.
• Mark the position of the coil (9) for any sign of corrosion.
• The position of the coil (10) and compare it to design position and the operating range of the spring.

Checklist for a Base Spring Support

1. Same as 4, and 5 above
2. Check the load flange to be sure it can move.

Spring Hanger Replacement

1. Have the replacement spring and associated hardware, along with the necessary tools to complete the removal of the existing spring support and the installation of the new spring support
2. Place the top and bottom travel stops in the existing spring to prevent movement
3. Attach a temporary support to the piping system during removal of the existing spring. Examples of temporary supports: chains and pulleys, rods assemblies, wire, …
4. Remove the existing spring support and related hardware.
5. Install the replacement spring support and related hardware.
6. Remove the lower travel stop and adjust the spring down until the pressure is off the upper stops. Remove the upper stop by hand.
7. Remove the temporary support.
8. Make the final adjustments to hot position on spring scale indicator by adjusting the nuts.
9. Tighten all lock nuts and discard existing spring and hardware.

By Amit Patil

January 28, 2011

It is important to make sure that all pipe supports are properly installed and functioning as per design. The improper installation or malfunctioning of even a single support may upset the whole piping system and create unbalanced forces throughout. Therefore, it becomes critical to do a walk down of the whole piping system and inspect each and every pipe support on the line, even though the overall design may incorporate several different types of pipe supports

Pipe hanger systems consist of several components to keep the piping properly suspended during the thermal operations which would be seen during the normal operating and shutdown cycles. All this deflection and movement is compensated by different types of pipe supports, which keep the piping system balanced. In the majority of the cases, the pipeline runs between two pieces of fixed equipment. The fixed equipment components act as the anchor points of the pipeline. Pipe supports are located between these anchor points and provide mobility throughout the piping system.

Because most manufacturers of the fixed equipment place limitations on the allowable loads, which can be seen at their terminal points, one should take extreme care to ensure that all pipe support components are properly installed and functioning. This methodology would ensure that the pipe loads imparted to those fixed equipment components will produce the desired reaction forces as specified by the fixed equipment manufacturers.

With the availability of modern design tools, it is possible to calculate the supporting force (and associated deflection) at each suspension point. These design methods and tools will help obtain a clear and concise understanding of what type of support is best suited for each individual application.

Understanding the critical nature of properly functioning pipe supports and how they contribute to the overall system, one should thoroughly inspect each component for proper operation during different phases of the life of the unit. This inspection survey could be done in following phases.

New Unit: – While performing the walk down survey of a new unit, one should pay particular attention to the following:

• The supports are correctly oriented for proper functioning.
• All stops, locks, bands, shipping bars, etc. have been removed
• The correct installed load and/or position of the supports has been established (including all required offsets)

During the initial startup phase, perform a second walk down survey to ensure that the piping system is performing as anticipated and thereby allowing the pipe supports to execute their specified applications.

Routine Maintenance: – Walk down surveys performed as part of routine maintenance, should focus on the following:

• The actual operating conditions do not exceed anticipated operating conditions including loading and/or movement of the pipe supports.
• Pipe supports, or any component thereof, have not been damaged during operation.
• The lifespan of the pipe supports have not been exceeded.
• Pipe supports have been properly maintained including lubrication, cleaning and adjustments.

At the Event of Failure or Shut down: – There could be several reasons a system failure or accident happens. When a process upset happens, some imbalanced forces and/or reactions occur throughout the system, and as a result the pipe supports may become imbalanced and fail to function properly. Thus, it becomes important to do a system walk down in order to inspect the functionality of the pipe supports and perform all necessary repairs, adjustments and/or replacements.

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Slide bearing plates are a very cost-effective way of providing for movement of mechanical systems.

Piping Technology & Products, Inc. supplies slide bearing plates for a variety of applications including support of piping, heavy equipment such as pressure vessels, and structural steel members. The plates provide a low coefficient of friction which can be attached to a supporting structure. This combination provides support while simultaneously allowing an object to move (slide) along the supporting surface.

Slide Plates: The “Sandwich” Concept

Like other leading designers in our field, Piping technology & Products, Inc. uses the “sandwich concept when applying slide plates to our customers’ system. Figure 1 shows a “sandwich” composed of two identical slide plates, one on top and another on the bottom. Each slide plate is composed of two components: a metal backing plate (which is the bun of the “sandwich”) and a low coefficient of friction material which is bonded to the metal backing plate.

Figure 1

In a typical application a slide bearing plate is welded to a structural steel member which is strong enough to provide the required support, but whose coefficient of friction is too high. Figure 2 shows an application in which a slide bearing plate is welded on top of one steel beam supporting another beam. When the top beam moves (due to thermal expansion, for example) it slide across the surface of the bearing plate without contacting the supporting beam. To return to our “sandwich” metaphor, the top half of the “sandwich” is bonded to the sliding beam, and the bottom half to the supporting beam.

Figure 2

Materials

One combination of materials that we often use is that of PTFE, 25% Glass Filled bonded to stainless steel. Both materials resist oxidation and have long lives even in stressful environments. For large slide plates, galvanized steel can be used in place of stainless to reduce the cost. PTFE, 25% Glass Filled provides a low coefficient of friction for most combinations of temperature and load. Figure 3 shows the recommended conditions for 3/32” PTFE, 25% Glass Filled. Note that a 500 psi load would be at the limit at 400 degrees F.

When the slide bearing plate must function at higher temperatures, graphite can be used instead of PTFE, 25% Glass Filled. The ideal temperature range of graphite is around 1,100 degrees F. For combinations of temperature and load beyond the capabilities of graphite, special designs must be considered.

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Difference – The major differences between standard springs and furnace springs are in their construction and their intended use. Furnace springs operate in much the same way as other variable spring supports. They are designed to support the pipe or tubes that are subjected to vertical thermal movement. The design changes are necessary because furnace springs are exposed to extreme temperatures.

Coatings – Furnace springs are almost always coated with red oxide primer in order to avoid any potential hazards when exposed to high temperatures. Standard springs are usually hot dipped galvanized, which, when exposed to high temperatures, tends to melt the zinc coating. The molten zinc can then cause damage to the surrounding pipe and equipment.

Construction – Whereas many of the standard variable spring components are interchangeable, the furnace spring assemblies are constructed from components unique to its application and intended use. The furnace spring incorporates a welded design for the housing assembly unlike standard spring housings that use a bolted configuration. The internal components of the furnace spring are designed to center the spring coils within the spring’s housing to prevent misalignment. In addition, the spring housing is modified to accommodate lug attachments on existing furnace tubing and equipment. Special fabricated casings, spring coils, and nameplates may also be used to accommodate increased travel.

Design – Some furnace springs are designed in order for the spring to be used as a means to determine the loading of the catalyst tubes. In these applications, an exact spring rate is determined for each assembly. The exact spring rate (which may differ slightly from published spring rate values) can then be used to determine the weight of the tubing system in order to balance the unit.

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Piping systems are used to transport gases, liquids and suspended solids between units of a plant. The temperatures of the contents and the pipe when the plant is in operation are often quite different from the ambient temperature of the surroundings. The heat transferred during changes in temperature causes metal pipe to expand and contract. Piping designers must provide for movement due to temperature changes and forces caused by pumps, gravity and other elements of the piping system. Proven engineering practices are included in various codes for piping services developed by professional societies such as the American Society of Mechanical Engineers. Most companies use these codes to develop specific standards for their products. The Manufacturers’ Standardization Society has developed standards for pipe supports. Designs which satisfy codes are important to satisfy laws and to obtain financing and insurance for projects. Table 1 provides a list of some important standards. PT&P’s Variable Spring Assembly Line Variable Spring Supports Ready for Shipment

Variable springs are one of the most important devices available to help support pipe. They are used to balance the concentrated gravitational load of vertical sections of pipe while allowing for thermal movement. Control of the direction and amount of movement of various sections of pipe is a major function of piping design. Computer models are used during design to predict the forces and movement at the point where a spring support is attached to the pipe. This is the information required to select the least expensive variable spring to use. Variable springs are relatively inexpensive and very reliable. When replacement is necessary, the engineer needs to determine if the original size is still appropriate or if the system has changed and different forces and movement must be considered. Pipe support manufacturers product catalogs which describe the items they routinely supply. The engineered spring support section of our catalog contains detailed information on variable springs. The load and travel tables contain the information required to size variable springs. An abbreviated version of this table is given in Figure 6.

Physics of Spring Coils Variable spring supports are attached to piping systems so the coil is in a vertical position as shown in Figure 1. Forces which compress the coil act in the same direction as the force of gravity. Movement due to temperature changes may be either up or down. Movement is a vector which has both direction (up or down) and magnitude. Travel is a scaler which describes the length of movement or compression of the spring coil. A spring can only be compressed so much before the individual coils come together and it “goes solid”. The usual method used to provide for larger travel is to provide a longer coil or stack standard coils together as shown in Figure 2.

Newton’s Third Law of Motion states that every force is accompanied by an equal and opposite force. A spring coil resists the force which compresses it. The more the coil is compressed, the greater resistance it has to further compression. This is why pipe supports which apply forces directly to the spring coil are called variables. The magnitude of the force required to compress a spring coil depends on the size of the wire used to form the coil and on the diameter of the coil. A “stronger” coil can be produced by using a larger wire or by winding the wire “tighter”. The two industry standards are the PTP-1 (short spring) and the PTP-2 standard spring. Twenty-three different wire sizes from .23 inches to 2.5 inches are used to increase the size of the force required to compress the coil. At PT&P, we use 00, 10,…,200, 210, 220 to designate the wire size (resistance to force) of the coil as shown in the top of Figure 6. Figure 3 shows a graph of pounds-force required to compress a PTP-2, size 90 coil as a function of the amount of compression. The graph is a straight line through the origin with a slope of 200 pounds-force per inch of compression. The slope of this line is the spring rate of this variable support. We compress the coil three inches during assembly (using 600 pound-force). The variable produced can be further compressed an additional 3.5 inches (range of deflection) by applying 1300 pounds-force. Each additional 200 pounds-force will compress the coil one inch.

Manufacturers recommend that only 80% of the range of deflection (called the operating range) be used for variables. The PT&P load and travel table uses red lines to enclose the loads in the operating range. Figure 3 marks both the minimum and maximum loads (700 and 1200 pounds-foce) for the operating range with red lines. The load at the center of the operating range (950 pounds-force) is marked with a blue line. The total compression of the coil at 700 pounds-force is 3.5 inches and the compression at 1200 pounds-force is 6.0 inches so the operating range is 2.5 inches. Figure 6 show how to obtain these values from the load and travel table of our catalog.

Selecting the Proper Size for a Variable The coil size (force required for compression) should be chosen to balance the expected force on the support when the system is operating. To do this, choose the size (00 to 220) which has its center compressive force (blue line) in Figure 6 as close as possible to the Operating Load (“hot load”). Choose the shortest (least expensive) coil and operating range which will satisfy the expected movement the support will experience during the temperature changes expected going from cold load (non-operating force applied) to operating load. Figure 5 illustrates the relation between the three quantities, operating load, cold load and movement. The direction of movement is defined by the change of load (force on the support) when the system changes from cold to operating. If the operating load is greater, the increase in the force will further compress the coil so the direction of the movement is down. If the operating load is smaller, the decrease in the force on the support will allow the coil to expand so the movement is up.

Variability Constraint The change in the force on the suport when changing from cold to operating is equal to the spring rate multiplied by the Travel (magnitude of the movement). Good practice as specified in MSS code requires that this change experienced by the support should be nom ore than 25% of the operating load, thus the inequality in Equation (a) must always be satisfied. Note that the spring rate is a function of both the coil size (Z) and the coil length (X). Coil size Z is chosen to match the operating load, so X is chosen as small as possible to satisfy Equation (a). The possible values of X are 1,2,4,6, and 8 which correspond to PTP-X and define the operating range of the support. (a) Spring Rate (Z,X) * Travel < 0.25 * Operating Load If you are interested in a computer program that will size variable spring supports shown in the PT&P catalog, you can download here.

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Vibrations produced by rotating machinery are very common for industrial piping systems. Bracing devices can be used to cope with vibration. Hold downs are one of the most widely used pipe support devices to restrain or dampen pipe vibration. A common location for applying a hold down exists near a compressor. The objective is to limit vibration caused by rotating equipment in order to avoid damage to the surrounding piping system. The hold down restrains the vibration introduced by the compressor as it compresses the gases in the pipe line. Hold downs are effective for the control of vibration and/or movement.

Vibration Control: PT&P’s Capabilities

Pipe Technology & Products, Inc. can custom design and fabricate hold downs for any pipe size application. In additional to hold downs, PT&P also offers other types of vibration control devices to accommodate different pipe orientations and locations:

Types:

Anchor Type (PTP HD-1)
Generally used on pipe lines where little or no insulation is required
Temperature: 750°F

Anchor Type (PTP HD-1)

APPLICATION: The PTP Fig. HD-1 is our anchor type hold down. This hold down is generally used in pipe lines where little or no insulation is required.
Material: Carbon Steel
Finish: Black or galvanized
Temperature: 750°F Black Steel
Temperature: 200°C Galvanized Steel

Anchor Type (PTP HD-2)
*generally used on compressor lines or similar applications with high temperature and massive thermal movement
*the clamp radius and the base are line with PTFE slide plate, allowing axial movement while restraining lateral movement.
*Temperature: 400°F

Anchor Type (PTP HD-3)
*generally used on compressor lines or similar applications with low thermal movement
*the clamp radius is lined with a belting material designed to absorb shock from vibration
*Temperature: 200°F
*belting material remains impervious to heat up to 200°F

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Mechanical Snubbers provided by Piping Technology and Products, Inc. have two modes of operation. In passive mode, i.e., motion caused by thermal loads, the resisting mechanism is disengaged and the snubber “free wheels” with very low resistance. In active mode the mechanism is engaged, and the snubber limits the acceleration to a low threshold value. There are other types of mechanical snubbers, but these are the most common ones.

The Mechanical Snubber Assembly

The snubber operates on the principle of limiting the acceleration of any pipe movement to a threshold level of .02 g’s. This is the maximum acceleration that the snubber will permit the piping to see. Should a disturbance attempt to accelerate the pipe in either direction, a braking force will be applied within the snubber of whatever magnitude required to limit the acceleration to a value less than .02 g’s. At the same time, thermal expansion, being a gradual movement, is not restricted. A particular feature of the snubber is that at no time does it lock and thereby become a rigid strut. Should a sudden acceleration occur and sustain continuously in one direction, the snubber will apply whatever force is necessary to limit the pipe movement to its present threshold value. The snubber’s performance is independent of the amount of force being applied.

Cross Section View of a Mechanical Snubber

The principle of operation can best be seen in the cross section view above. Two structural telescoping members, colored red and yellow, are connected between the pipe and fixed structure. Within these telescoping tubes is a ball screw and nut which serve to convert the linear telescoping motion, which would occur during a seismic disturbance or thermal changes, to rotational motion of the blue ball screw and drum assembly. This rotational motion is coupled to an inertia mass colored red. The coupling consists of a resilient capstan spring colored orange.

When a disturbance occurs that exceeds the threshold “g” level (.02 g’s), the ball screw and drum attempts to angularly accelerate the inertia mass. The inertial resistance of the mass causes the resilient capstan spring to tighten around a hardened mandrel which is part of the red structural tube. In this manner, a restraining force is applied against rotation of the ball screw, and, in turn, linear telescoping of the silver and red members.

The design of the unit is completely symmetrical, and the same capstan spring will apply this braking action in both the tension and compression loading, which in turn means clockwise or counterclockwise angular acceleration. Therefore, the braking characteristics of the unit in tension and compression are identical.

It can also be noted that when the sudden force is applied, the resisting force is applied by the inertia mass. The inertia mass is mounted to turn freely, and therefore the moment the acceleration drops below the threshold value, it no longer applies a braking force. In additional, the capstan spring is always urging the inertia mass back to an un-braked condition. The net effect is a design which continuously throttles or brakes to limit and control the acceleration. During thermal compensation, the gradual movement, normally associated therewith, is far below the threshold acceleration setting; and therefore, the inertia mass will gradually move without tightening the capstan brake. Should this thermal movement be uneven or jerky as might occur because of a hanger or skid sticking, the unit might momentarily brake, while permitting the pipe movement.

Standard Sizes, Load Rating and Maximum Stroke

PT&P has two test machines specifically designed for testing snubbers for piping applications, one for routine testing of hydraulic snubbers, and a more sophisticated machine that can perform a wide range of tests on both hydraulic and mechanical snubbers.

PT&P’s STADAS Snubber Test Machine

The STADAS is a trailer mounted snubber test machine manufactured by Paul-Munroe Inc., California. This computer controlled, hydraulic test machine is used to test the performance of mechanical and hydraulic snubbers ranging in size from 150 lb. to 135,000 lb. The STADAS may also be used to verify snubber conditions at the point of manufacture, or during an outage (refueling) at the plant site.

The STADAS machine is capable of controlling a test, following any load sequence needed. It can then plot the test results immediately, and save the results to a disk for future reference.

The main frame of the STADAS machine is comprised of two drive cylinders, which supports the snubber being tested. The smaller drive cylinder is used for snubbers rated to 6000 lb., and the large drive cylinder is used for snubbers rated up to 150,000 lb. Each drive cylinder is equipped with an electro-hydraulic servo-valve load cell and connections to the hydraulic system.

PT&P has two test machines specifically designed for testing snubbers for piping applications, one for routine testing of hydraulic snubbers, and a more sophisticated machine that can perform a wide range of tests on both hydraulic and mechanical snubbers.

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Most hydraulic snubbers have a piston which is relatively unconstrained in motion at low displacement rates. At high displacement rates the piston “locks up”, that is, the force required to move the piston increases substantially, usually as a result of the closing of a valve.

Some of the features include:

• External pressurized hydraulic reservoir for positioning flexibility in any spatial orientation.
• Allows free thermal movement of piping under normal operations.
• Restrains shock loading, both in tension and compression.
• System’s movement is controlled by a flow control device.

Standard Sizes of PT&P Hydraulic Snubbers

PT&P offers seven sizes with cylinder bores of 1 1/2 to 8 inches. These units have a normal load range of 3,000 lbs to 130,000 lbs. All are made to include reservoirs in 6, 12, or 18 inch strokes. Snubbers, as they are sometime referred to, are available with remote reservoirs.

Application

For use on piping systems or equipment when restrained thermal movement must be allowed, but which must be restrained during impulsive or cyclic disturbance. The unit is not effective against low amplitude, high frequency movement. Preferred usage, with standard settings, to prevent destructive results due to earthquakes, flow transients, or wind load. Special settings are available to absorb the continuous thrust resulting from safety valve blow-off or pipe rupture. For the most effective operation of the unit, please specify the mounting position, vertical or horizontal.

Ordering

Please specify figure no., cylinder size, stroke, load, cold, and hot pistons settings, and piston end option. If clamp is required, please specify nominal pipe size, or special O.D. and clamp material. Specification or description of any additional optional features or special settings is required.

Dimensions

*Loads must not be applied outside a 10° included angle cone of action to the pipe clamp axis w/o special authorization

E-takeout for Fig. 2100 Hydraulic Snubbers and Sway Strut Assemblies

Schematic of Hydraulic Snubber under normal conditions	Schematic of Hydraulic Snubber under locked conditions

Hydraulic Snubber Test Machine

The Hydraulic Snubber test machine is mainly used to perform testing for hydraulic snubbers, it has a similar test bed to the STADAS machine, however, it is not computer controlled. It is adequate to perform all the tests normally required for hydraulic snubbers, such as quality control and periodic tests that are sometimes required as part of plant maintenance programs. Mechanical snubbers can also be tested on this machine, but due to the complication of the process, mechanical snubbers testing is usually performed on the STADAS machine.

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Objective

To show the results of temperature and stress analysis in three pipe support shoes intended for cryogenic operation

It is necessary to know the temperature distribution for cryogenic conditions because most steels lose ductility as the temperature decreases from normal operating conditions. That is, the steel becomes more brittle.

Any steel structure will have local areas of high stress. This may be caused by, for example, sharp corners in the design, or by inclusions in the material. At normal temperatures the material in the region of these concentrations will yield, redistributing the load locally. Typically these have no significant effect on the integrity of the structure as a whole. In brittle materials, the material cannot yield, so the only alternative is for it to form a crack. Frequently, once the crack has been formed, it will propagate at the approximate speed of sound until it reaches the end of the part, thus causing a catastrophic failure.

Further complicating the picture is the fact that the tendency for brittle fracture increases with part thickness, and this increase is temperature dependent. Clearly, temperature is a critical variable in selecting steels for cryogenic applications.

The three shoes considered are of the same basic design, but are for three different pipe sizes: 6, 20, and 42 inch. The basic design is shown in Figure 1.

Figure 1: Complete pipe shoe assembly

The shoes have three major components. They are:

1. An outer cover and base of carbon steel, shown in Figure 2A. This has ring stops at either end to engage the insulation to keep it from moving. A point of particular concern was the temperature and stress of the steel on the inner surface of this ring.

2. A pipe which carries the cryogenic fluid. This has an attached ring of stainless steel, again for engaging the insulation. This is shown in Figure 2B.

3. The polyurethane foam insulation (PUF) between the first two parts, shown in Figure 2C. As illustrated, the insulation has outer grooves for the shell ring stops, and an inner groove for the pipe ring.

Figure 2A: Shell	Figure 2B: Pipe and Ring	Figure 2C: Insulation

STRESS & THERMAL ANALYSIS

The stress and the thermal analyses are, for the most part, quite similar. The finite element program, Ansys was used for both. There are two main differences between the analyses. Primarily, the material properties are different. Young’s modulus vs. thermal conductivity, for example; the other main difference is in the choice of elements. In this case we used Ansys element 92 for the stress analysis, and thermal element 83. These are both 10 node tetrahedral solid elements. The primary difference is that the thermal element has only one degree of freedom at each node (for temperature), while the stress element has three, one for x, y, and z displacements.

STRESS ANALYSIS

The main concern in the 42 inch pipe shoe case was the inner surface of the outer retaining rings. If the temperature is too low for the stress, the specified steel may not be appropriate according to the code.

The other two cases are very similar to this one and those stress results are given in Appendix A.

Applied loads for all three cases are given in Table1. Note that while the loads were specified in kN, the actual computations were done in inches and pounds, and the results are reported in those units.

Table 1: Applied loads for the stress analysis of the three shoes

Figure 3: Stresses in the complete assembly for the 42″ pipe shoe
(It is more informative to view the stresses on the individual components)

STRESS ANALYSIS

Figure 4 shows an overview of the stresses on the shell and base, and Figure 5 shows a detail of the stresses in the area of maximum stress. The stresses in the shell are generally quite low. However, at the intersection of the base with the cylindrical shell there is a local stress concentration with somewhat higher stresses. This should not be a problem at this location. As we shall see, the temperature here is quite high, as compared to the rest of the shoe.

Figure 4: Stresses in shell and base	Figure 5: Detail of stresses in the location of maximum stress

The stresses in the pipe and its ring are shown in Figure 6. The stresses are largest where they would be expected, at the intersection of the two.

Figure 6: Stresses in the pipe and its ring

STRESS ANALYSIS

In Figures 7 and 8 we show the stresses in the insulation. Figure 7 is an overview. Figure 8 is a cut away view showing stresses on the inside. The fact that the stress is mostly in the bottom quarter is the result of the significant vertical load.

Figure 7: Stresses in the insulation	Figure 8: Section detail of insulation stresses

THERMAL RESULTS

For all three shoes, the thermal conditions were the same. The inner surface of the pipe was -270°F. The ambient air temperature was -20°F, with a wind velocity of 78.74 in/sec (2 m/sec). Based on this wind speed, and the diameter of the outside shield, we computed a film heat transfer coefficient. The values for the three cases are shown in Table 2.

Table 2: File heat transfer coefficients used in the study

The values of thermal conductivities used are in Table 3. Note that for the 32 lb PUF case the coefficient is linear with temperature. For the 20 Lb PUF we had only one data point, so we were unable to use a temperature dependent relation.

Table 3: Thermal conductivities for materials used in the study

TEMPERATURE DISTRIBUTION

In Figure 9 we show the temperature for the complete assembly. As expected, the temperature on the pipe is the specified temperature, and the temperature on the base is slightly lower than the ambient.

Figure 9: Complete assembly (°F)

Again, it is helpful to look at the individual parts. Figure 10 shows the temperature on the shell and base. The coldest temperature is greater than -40°F, and thus much higher than the temperature that would cause concern.

Figure 10: Shell-base assembly (°F)

TEMPERATURE DISTRIBUTION

In Figures 11 and 12 we show the temperature distributions on the other two parts. There is nothing particular to note concerning these plots. (Temperatures are °F)

Figure 11: Ring and Pipe	Figure 12: Insulation

Figure 13 shows the interface between the steel and the insulation. As expected, the temperature in both the inner and outer rings is little affected by the temperature in the surrounding insulation, and is nearly the same as if there were no insulation present.

Figure 13: Section view showing interface between the steel and insulation

The report includes both the temperature and the stress at each point of the shoe assembly. This should be adequate for a proper selection of steels for the design.

APPENDIX A

In this Appendix we show the results of the stress analyses of the 20 and 6 in. pipe cases. There is little to add to what has been said about the 42 in. case.

Figure A1: Stresses in 20 inch assembly

Figure A2: Stresses in shell-base part	Figure A3: Detail of maximum stress

Figure A4: Stresses in ring pipe	Figure A5: Stresses in insulation

Figure A6: Cut away view of insulation stresses

Figure A7: Overall stresses on 6 inch assembly

Figure A8: Shell-Base stresses, 6 inch case	Figure A9: Detail at maximum stress

Figure A10: Stresses in pipe and ring	Figure A11: Stresses in 6 inch insulation

Figure A13: Detail of ring	Figure A12: Stresses on insulation section

At the left, we show a expanded view of the ring – pipe intersection. The high stresses shown in Figure A10 are very local, and only near the pipe end. This is where the loads were applied. The total load is correct, and, as can be seen, stresses near the intersection are more moderate.

APPENDIX B: Temperature Distribution

Note: Temperatures are degrees °F

Figure B1: Complete assembly for 20 inch pipe shoe

Figure B2: Shell-base assembly for 20 inch pipe shoe

Figure B3: Pipe-ring assembly for 20 inch pipe shoe

Figure B4: Insulation block for 20 inch pipe shoe

Figure B5: Cut away view of insulation block for 20 inch pipe shoe

Figure B6: Pipe shoe assembly for 6 inch pipe shoe

Figure B7: shell-base assembly for 6 inch pipe shoe

Figure B8: Pipe-ring assembly for 6 inch pipe shoe

Figure B9: Insulation for 6 inch pipe shoe

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Pipe Clamps

PT&P manufactures a range of standard clamps covering common pipe sizes. These are made of carbon steel and either galvanized or left black. In addition, we frequently fabricate custom clamps for special applications. Applications examples include high temperatures requiring alloy steel, or those requiring nonstandard dimensions.

In these analyses, we have assumed that the bending radius of forming the ears is 1.5 times the plate thickness. This is the value calculated and used as a practice at PT&P. The AISC recommends that the values are three times the thickness of the plates of 1 to 1-1/2” thickness and four times the thickness for greater thickness. The Finite Element Analysis demonstrates that larger radiuses would result in lower stresses. These cases prove that Finite Element Analysis at PT&P has led to significantly lower costs to our customers by justifying the use of less steel.

FEA & Specially Designed Clamps

Recently, a stress analysis was conducted of an 82 plastics pipe constrained in the axial direction by a special pipe clamp. PT&P’s final design assumes that three clamps in a series will resist at a total expected axial force of 8,000 lbs. Since the area of the pipe in contact with the support affects the overall stress distribution, it is necessary to determine how much contact there is at different pressures via FEA in two steps.

Step #1

The first step establishes whether or not the stresses in the clamp and pipe

were in acceptable limits and fixes a starting condition for the remainder of the analysis. Here, the clamping force (estimated to be 13,300 lbs.) alone acted upon the clamp and pipe. The initial design of the clamp indicated stresses exceeding permissible boundaries by being greater than 50,000 psi. The design was thus modified to alleviate the problem.

Step #2

This step involves determining the highest load achieved before slipping.

Figure 1 shows stresses on the complete assembly and points that the maximum stress on the clamp is 28,800 psi. The high stress is localized in very small areas at the tip of the gusset. In Figure 2, we have plotted stresses in the plastic pipe only. The maximum stress in the pipe is an acceptable 2,360 psi.

Summary

The maximum stress on the clamp is determined to be 28,800 psi. Stresses in Figure 1 would normally be of little concern in mild steel at typical operating temperatures. If it were a problem, extending the gussets upwards is recommended. In addition, it is recommended to increase pipe length if higher stress is desired. This is due to increased contact between the pipe and clamp, which spreads stress over more area allowing for greater pressures.

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Pipe Clamps

PT&P manufactures a range of standard clamps covering common pipe sizes. These are made of carbon steel and either galvanized or left black. In addition, we frequently manufacture custom clamps for special applications. Examples of such applications are high temperatures requiring alloy steel, or applications requiring nonstandard dimensions.

PT&P normally analyzes such special designs using Finite Element Analysis. One reason for this is to ensure that the design has sufficient strength to support expected loads. This is also done to see if the clamp can be made more economically and still function satisfactorily.

FEA & Specially Designed Clamps

Recent examples include two Figure 80 type clamps for a high temperature application. Figure 80 clamps are heavy-duty three-bolt pipe clamps. Due to the high temperature duty, they were made from alloy steel, A387 Grade 91. This steel contains 9% Chromium, 1% Molybdenum and small amounts of Niobium and Vanadium. Both clamps are for 24” pipes, one for a load of 68,200 lbs., and one for 33,500 lbs.

Results: Clamp Test #1

The initial FEA was conducted on the clamp designed to sustain a load of 68,200 lbs. For this load, the PT&P Clamp Sizing program called for a clamp made from two 2” thick x 12” wide strips of steel. This case required the FEA find a maximum stress of 6150 psi. This is considerably less than
the allowable stress for the steel at the operating temperature. As a result, the clamp thickness was reduced to 1-1/2“. It was confirmed that the stresses were again below the allowable.

12 x 2” Clamp for 24” pipe with load of 68,200 lbs

Results: Clamp Test #2

The second FEA was conducted on a clamp designed for a load of 33,500 lbs. According to PT&P’s standard sizing program, this clamp’s design was made from 1-1/2” x 10” steel. The FEA found a maximum stress of 4250 psi, well below the 10,300 psi allowable. Thus, PT&P engineers reduced the size of the clamp to 1-1/4” x 8”, with resulting stress of 7331 psi.

1-1/2 x 10” Clamp for 24” pipe with a load of 33,500 lbs

Conclusions

In these analyses, we have assumed that the bending radius of forming the ears is 1.5 times the plate thickness. This is the value calculated and used as a practice at PT&P. The AISC recommends that the values are three times the thickness of the plates of 1 to 1-1/2” thickness and four times the thickness for greater thickness. The Finite Element Analysis demonstrates that larger radiuses would result in lower stresses. These cases prove that Finite Element Analysis at PT&P has led to significantly lower costs to our customers by justifying the use of less steel.

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FEA: Case Studies

Product: Transition Piece

Technical Information: 3D Modeling Software, Parametric Technologies Modeler, Finite Element Analysis Software, COSMOS

Case Study Information

This item was modeled to the specifications included in drawing 6819001-159020. It uses a standard flange at the top as specified. It has been modified to include the extra length of pipe and to eliminate two flanges. The rectangular flange is fabricated of 1.25 x 5 inch stock and is made of A36 steel. The whole transition and pipe is fabricated of 3/8” thick A516 Grade 70 steel and this was expressed in the model. The model has a temperature condition of 105° C/221°F.

Two cases were modeled in which all references to pressure, PSI, refers to PSIG (PSI gage). The first was at operating condition when the transition piece experiences 18.7 PSI internal pressure. The second was at full vacuum when the transition piece experiences 15 PSI external pressure. The stress and displacement results are expressed in the following plots.

*Note: All references to pressure, PSI, refers to PSIG (PSI gage)

Illustration of Pressures and Displacements – Transition Piece: Internal Pressure = 18.7 PSI

Stress Plot #1 – Transition Piece: Internal Pressure = 18.7 PSI

This model has an internal pressure of 18.7 PSI on every internal surface. The chart on the left of the model shows the stresses at every point in the model by a color coding system.

Stress Plot #2 – Transition Piece: Internal Pressure = 18.7 PSI

This model has an internal pressure of 18.7 PSI on every internal surface.

Displacement Plot #1 – Transition Piece: Internal Pressure = 18.7 PSI

This model has an internal pressure of 18.7 PSI on every internal surface. The chart shows the displacement at every point in the model by a color coding system which represents inches of displacement.

Illustration of Pressures and Displacements – Transition Piece: External Pressure = 15 PSI

Stress Plot #1 – Transition Piece: External Pressure = 15 PSI

This model has an external pressure of 15 PSI on every external surface. The chart above shows the stresses at every point in the model by a color coding system.

Stress Plot #2 – Transition Piece: External Pressure = 15 PSI

The chart above shows the stresses at every point in the model by a color coding system. This model has an external pressure of 15 PSI on every external surface.

Displacement Plot #1 – Transition Piece: External Pressure = 15 PSI

The chart above shows the displacement at every point in the model by a color coding system which represents inches of displacement.

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Objective

Stress analysis was performed on a dual hinged expansion joint to determine whether the plate thicknesses are sufficient to insure that the stresses are within code allowable.

Figure 1. Portion of a 36 inch Bellows assembly (EJ-004)

Figure 2. Portion of a 18 inch Bellows assembly (EJ-009)

Figures 1 and 2 show a portion of a similar dual hinged expansion joint that was built by U.S. Bellows, Inc. The figures show the connection of the tie rods to the bellows. The tie rods are connected to a plate-gusset assembly, which is welded to the bellows. This note describes a stress analysis to determine whether the plate thicknesses are sufficient to insure that the stresses are within code allowable.

Parameters

Ansys Finite Element Program

The analysis was done using Ansys Finite Element Program. The element used was the Ansys solid 187 element, a 10 node tetrahedral element.

The results for EJ-004 are shown in Figures 3, 4, and 5. Figure 3 shows the stresses in the complete assembly.

Figure 3. Von Mises Stresses for complete assembly.

Stresses in the Plate

Figure 4. Stresses in the plate-gusset assembly of EJ-004 (Front View)

Figure 5. Stresses in the plate-gusset assembly of EJ-004 (Back View)

The stresses on the assembly are all nearly below the allowable. There are two small areas where there are stress concentrations, but the areas are small, and the stresses drop rapidly away from the peak. These local areas should not be of concern.

Stresses in EJ-009 Assembly

Figure 6. Stresses in the complete assembly for EJ-009

Figure 6 shows the stresses in the complete portion of the assembly analyzed. There are some quite high stresses shown. These are in the pipe at the tip of the guests. They are clearly due to very large change in stiffness at that point, – from 3/8 in. in the pipe to 1 7/8 in. at the tip of the gusset.

Figure 7. Stresses on the tie rod plate for EJ-009

Figure 7 shows the tie rod support plate. The stresses are all clearly below the allowable. In conclusion, the thickness of both plates analyzed are adequate.