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At first look, the practice of insulating flanges seems to have several clear advantages.  Insulating the flanges would:

• cut back on radiated heat, thus providing more thermal efficiency (and hence cost reductions) to most thermally intensive industrial processes.
• lower the risk of a worker accidentally coming into contact with the hot, exposed flange.
• provide an initial barrier of protection in the event of a high-temperature steam leak – or other hazardous substance.

 

While these reasons seem compelling, they miss the central point.  The most important function of a gasketed joint is to prevent leakage. When viewed in terms of this primary function, it becomes clear that the negative effects of insulating flanges outweigh the apparent benefits. 

1. Insulating the flanges boost the temperature of the studs. In the hottest of flange applications, this increases the chance that the studs could thermally yield.  Recognizing this fact, Chevron forbids insulating any flanges used in processes above 850-degrees F.  They further state that if any flanges at temperatures between 700-degrees and 850-degress are determined to need insulation (for whatever reason), that ASTM A 193 Grade B16 studs must be used, with ASTM A 194 Grade 7 nuts.
2. Insulating the flanges can increase the chance for leakage when a flange is cooled rapidly from the inside (such as when a wash-oil is run through a system). Under these conditions, the bolts cool slower than the flanges, resulting in a measurable (6,000 to 10,000-psi) drop in stud load.  This precipitous drop in stud stress is responsible for a high percentage of the leaks that occur with thermal upsets.  Even though this is a transient effect, it is only made worse by the use of insulation, which slows the rate at which the bolt can radiate its heat. (Chart on next page, courtesy of Chevron.)

Clearly, the practice of insulating flanges tips the scale in favor of developing more leaks.  So as a general rule, flanges should not be insulated. 

 

But what about worker safety?

 

If the primary purpose of the proposed insulation is personnel protection, that goal can be achieved through other methods that do not impact the thermal balance of the flange.  For example, Chevron uses expanded metal guards that prevent accidental contact with the flange, while still allowing free airflow.

Insulating flanges is not only going to make it more likely that you will suffer leaks, but it will also prevent you from noticing those leaks when they first occur – when they can most easily be abated.  

 

Conclusion: It is our position at ERIKS that as a general rule, flanges should not be insulated. 

One of the most frequent mistakes made when talking about forces is to use the words “load” and “stress” interchangeably, as if they mean the same thing.  They don’t. 

 

Simply put:

LOAD is in pounds; STRESS is in pounds per square inch.

 

So, there are two elephants – one in sneakers, one in stiletto high heels.  Both elephants weigh the same. However, as they approach the watering hole, the elephant in stilettos starts having problems, as her heels keep sinking into the ground.  Why?

 

Well, since they weigh the same, we know that the total amount of load (pounds) they put on the ground is exactly the same.  However, the better-looking elephant (the one in the heels) is concentrating that load onto a much smaller area (pounds per square inch), so the stress on the ground is much higher where she steps. 

 

In the world of equipment and gaskets, Load and Stress come into play with both the studs and the gaskets.  

 

STUDS

The limit on studs is always given in terms of stress.  If you look at the “Stud Data” tab on the Exchanger Gasket Workbook, you’ll find a table listing the “Yield Strength” of various stud materials. This value is a stress value, and is given in pounds per square inch, or “psi”.  You’ll see that all B7 studs up to 2-1/2” diameter have a minimum yield strength (We’ll get into what that means later) of 105,000-psi.

 

Let’s take two studs – one 5/8”, and one 1-3/4” – and look at how they relate in terms of stress and load.  Since they are both B7 material, and both are less than 2-1/2” diameter, they can both be stressed to 105,000 pounds per square inch.  Imagine these studs screwed into a heavy I-beam on one end, with a special nut with a clevis on the other end, as shown in this sketch.

 

How much weight (Load) would we have to hang from each of the clevises in order to reach 50,000-psi stress in the stud?  Clearly, the larger stud will take a lot more weight in order to get to the same stress.

 

 

The relationship between stress and load in a stud is determined by the “Tensile Stress Area” of the stud.  This is the effective cross-sectional area of the stud in tension.  The stress within the stud is equal to the total load on the stud divided by the tensile stress area:

 

Stress = Load / Tensile Stress Area

and:

Load = Stress * Tensile Stress Area

 

Since “load” is given in pounds, and “tensile stress area” is in square inches, the units of stress are pounds per square inch.

 

 

The Tensile Stress Area of the stud can be found on the “Stud Data” page of the Exchanger Gasket Workbook.  Looking up the studs found in our example, we find that the tensile stress area of the 5/8” stud is 0.226 square inches, while the tensile stress area for the 1-3/4” stud is 2.081 square inches.  From the equations above we can compute:

 

For the 5/8” stud:            50,000 * 0.226 = 11,300 pounds

For the 1-3/4” stud:        50,000 * 2.081 = 104,050 pounds

 

So, we would have to hang almost 10 times as much weight from the larger stud just to have the same stress as the smaller stud.  Stated differently, the stress is the same, but the load is higher because the cross-sectional area of the stud is so much greater.

 

Essential Thought Sequence:

In considering gasket applications, we almost always need to be thinking in terms of gasket stress – the compression in pounds per square inch on the gasket.  However, the gasket stress always comes from the stud stress.  To get from the one to the other, we need to go through the following steps:

 

1. Convert stud stress to stud load by using the tensile stress area.
2. Multiply the individual stud load by the number of studs to get the total load being transferred to the gasket.
3. Divide this total load on the gasket by the number of square inches (area) of the gasket being squeezed to determine the stress (pounds per square inch) of the gasket.
   
   

This simple conversion from stud stress to stud load, and back from stud load to gasket stress becomes second nature as we work with gasket sealing issues.

Background:  This hot feed / effluent heat exchanger has a history of leaking that extends back 10 years and includes clamping and leak-sealing by outside contractors, as well as changing the gasket styles on at least two occasions to “upgraded” styles suggested by manufacturers.  The most recent change in 2010 was to a kamprofile gasket with Mica/Graphite/Mica sealing elements.  At that time, the flange surfaces were inspected, the new gaskets were installed, and the flange was tightened per refinery procedures.  However the Tubesheet-to-Shell joint developed a small leak shortly after start up, and a steam ring was installed.

This is a TEMA type BEU exchanger with only one joint (channel – tubesheet – shell), with two gaskets and one common set of studs.  The stud material was originally B7, but has been upgraded to B16.  The joint utilizes 64, 1-1/8” studs that were torqued to a final value of 700ft.lb.

Analysis:  The information provided by the client was loaded into the Exchanger Gasket Workbook to analyze the gasket seating stress under operational conditions.  Barring defective components – whether flanges or gaskets – heat exchangers leak due to low gasket seating stress.  For this reason we always look first at the gasket seating stress at full operating pressure.  Regardless of the “Y” value of the gasket, we have found that reliable sealing of graphite-faced gaskets requires a gasket stress between 10,000 and 40,000-psi.  If possible, we work to achieve a gasket stress in the 20,000 to 25,000-psi range. Not only is that tight enough to provide a great seal at start up, but it is also tight enough to offset the inevitable relaxation found in heat exchanger joints.

Given the leakage history on this unit, we anticipated a low gasket stress.  However, the gaskets on this exchanger were loaded to nearly ideal levels.  (The final page of this analysis is a snapshot of the Exchanger Gasket Workbook.)  The question that must be resolved is how an exchanger that is designed to deliver 25,000-psi gasket stress can begin leaking shortly after start up.  What factors might have contributed (historically) to the low-stress conditions that allowed leakage?

A full review of the equipment drawings and the related documents led us to two significant factors related to the equipment and gasket
designs, and two additional factors that were assembly related.  Chief among these contributing factors was the threaded tubesheet.

The equipment drawings show (below) that the 64 holes in the tubesheet are threaded for the 1-1/8” studs.

As currently configured, the portion of the stud on the channel side acts independently of the portion of the stud on the shell side.  The total amount of stretch in the stud is 0.0245”.  However, if the tubesheet were drilled out, the full length of the stud would stretch under tension, providing 0.0321” of stretch – a 31% increase!  For a flange that leaks habitually, this is a huge improvement of stud response.

It is interesting to note that the gasket stress drops more than the stud stress, because of the effect of the hydrostatic end-load on the joint.  The table below summarizes the resultant gasket stress if there were a relaxation of 0.006” on each side of the tubesheet.  (This is not an excessive estimate.) Each side is considered independently, because they are isolated from each other by the threaded tube sheet.  Notice that the stud stretch and
stud stress are directly related, because the stud is a linear spring.  But the gasket stress drops far more, because of the stud stress that is diverted to offset the hydrostatic end force in the joint.

The table below models the same scenario – a relaxation of 0.006” on each side of the tubesheet, for a total of 0.012” relaxation.  However, in this case the tubesheet has been drilled out, allowing the full stretch of the stud to interact with both gaskets.  Compared to the above chart, the percentage of stud stretch and stud stress drops far less, and the resultant gasket stresses are improved 30 – 50%!

Clearly, the short studs amplify the loss of gasket stress when relaxation occurs. Because of these dynamics, ERIKS recommends drilling out all threaded tubesheets.  The purpose of the stud is to stretch to maintain gasket stress, and threaded tubesheets limit the ability of the stud to do so.  We have clients who are now routinely drilling out tubesheets whenever they come across them.

If a refiner feels that they need to retain the ability to hold the Tubesheet-to-Shell joint tight while removing the Channel, they can do this with only one stud in four.  So even if it is felt that drilling all the holes is not prudent, drilling three out of four holes would still improve the stud load on the gasket, while preserving the ability to remove the Channel independently.

Summary:  While a simple numerical analysis seemed to indicate that there was adequate seating stress on the gaskets to prevent leakage, the fact that leakage was regularly occurring indicated underlying problems that were resulting in low gasket stress.

Beside issues related to the specific style of gasket utilized, the re-use of studs, and the lack of a post start-up retorque, the greatest contributor to the loss of gasket stress was the threaded tubesheet; which, in this case predated any changes in the gasket style.  This conclusion is in line with field observations of leaks that are related to threaded tubesheets.

Recommendation: Drill out all 64 holes in the threaded tubesheet to allow the stud the maximum stretch with the calculated stud stress.  As a bonus, this will make assembly easier, as the stud does not have to be tightened from both ends.

Pittsburgh, Pennsylvania – ERIKS North America (ERIKS NA), a leading distributor of industrial solutions, is pleased to announce the appointment of Rich Holderman as Key Industry Director.

In this role, Rich will drive growth and development in the Steel, Tire, Aggregates and Pulp & Paper industries, focusing on ERIKS NA’s conveyor belting business.

Rich brings 38 years of experience in the belting industry to ERIKS NA. He began his career as a belt splicer and worked his way up to a sales role and then into management. With extensive knowledge of the industry, both from a commercial and operational standpoint, Rich is a past President of the NIBA organization and served on distributor advisory boards of the largest industry suppliers in North America. Before joining ERIKS NA, Rich worked at Almex Group as their COO.

“We are thrilled to welcome Rich back to the ERIKS NA team,” said Jeff Crane, CEO of ERIKS North America. “Rich’s extensive experience and knowledge of the belting industry, coupled with his previous experience with our organization, will be instrumental in driving growth and development in our Key Industries.”

Commenting on his appointment, Rich said, “I am excited to rejoin the ERIKS NA team and take on this new role. I look forward to utilizing my experience and knowledge to reinvigorate our belting business and drive growth in our strategic markets.”

ERIKS NA is proud to welcome Rich Holderman back to the team and looks forward to his contributions in his new role as Key Industry Director.

About ERIKS North America:

ERIKS North America, a portfolio company of LKCM Headwater Investments, is a leading distributor of fluid and material conveyance solutions for industrial customers. Our technical solutions and services keep our customers running, reduce downtime and total cost of ownership.

Contact:

Lauren Shaffer, Digital Marketing Specialist

Lauren.shaffer@eriksna-dev.local

412-925-7390

February 6, 2023

PITTSBURGH-ERIKS North America, a leading distributor of fluid and material conveyance solutions for industrial customers, announced today that it has appointed Michael Pastore as its Chief Operating Officer (COO).

Michael brings over 20 years of industrial distribution experience to ERIKS North America and has held executive leadership positions across multiple functional areas. He most recently served as Vice President, Operations at Kaman Distribution Group and, throughout his career, Michael has championed excellence in Supply Chain, Branch Operations, and IT. He brings a proven track record of improving operating efficiencies and service levels. 

As COO, Michael will lead all manufacturing, quality, supply chain, EHS and customer service activities at ERIKS North America. He will also be responsible for setting broad operational goals for the organization, implementing best practices, optimizing inventory management and improving productivity. 

“We are thrilled to have Michael join our team as COO,” said Jeff Crane, CEO of ERIKS North America. “His track record of success driving operational excellence makes him a valuable asset to our organization. We look forward to the positive impact he will have on our company.” 

“I am excited to join ERIKS North America and to be a part of an organization that is committed to continuous improvement and outstanding customer service. I look forward to working with the team to solidify the company’s position as the preferred provider of fluid and material conveyance solutions,” said Michael Pastore, COO of ERIKS North America.

About ERIKS North America: 

ERIKS North America, a portfolio company of LKCM Headwater Investments, is a leading distributor of fluid and material conveyance solutions for industrial customers. Our technical solutions and services keep our customers running, reduce downtime and total cost of ownership.

Contacts

Lauren Shaffer
lauren.shaffer@eriksna-dev.local
412-925-7390

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