Author Archives: Christy Behnke

Hydrogen Embrittlement Susceptibility of Case Hardened Steel Fasteners

Hydrogen embrittlement is a potentially catastrophic failure mode of high strength steel fasteners. One of the leading indicators of susceptibility to this phenomenon is hardness. While the susceptibility of case hardened fasteners is undisputed, minimal data exist to explain how the variables of case hardness, case depth, and core hardness work together to influence overall fastener susceptibility to failure. Peak Innovations Engineering principal engineer John Medcalf investigated this topic, and along with co-authors Brian Thomas and Salim Brahimi, presents his findings in the following paper. The research indicates that hardness levels found in current fastener standards may be susceptible to hydrogen embrittlement in corrosive environments. Future research is planned to address susceptibility to internal hydrogen embrittlement.

Find a link to the abstract here.

You can purchase the full article directly from SAE International.

Medcalf, J.S., Thomas, B.G., and Brahimi, S.V., “Hydrogen Embrittlement Susceptibility of Case
Hardened Steel Fasteners,” SAE Technical Paper 2018-01-1240, 2018, doi:10.4271/2018-01-1240

John Medcalf presents at ASTM Hydrogen Embrittlement Workshop

Peak Innovations Engineering’s Principal Engineer, John Medcalf, is an industry leader in his work with hydrogen embrittlement susceptibility of case hardened screws. On November 8th, 2018 John will present on the results of his studies at the 14th Annual ASTM F07 Hydrogen Embrittlement Workshop. You can find the abstract for this work here.

This ASTM event is the only of its kind and devoted to hydrogen embrittlement testing technology for the aerospace industry.

ASTM HE Event Information

Peak’s Newest Team Member: the MC950

Peak is excited to announce its most recent addition to the team with the purchase of a MC950 from Micro Control, Inc.

The purchase of the MC950 unit displays Peak’s commitment to evolving with the industry and remaining on the cutting edge of technology within ultrasonic joint testing and validation.  The newest model, MC950, provides users with faster readings working in combination with the newest program update which has also been recently released.

The purchase of the new MC950 unit increases Peak’s capacity and capabilities, and has already proven to be a great addition to Peak’s existing two MC900 units as well as a wealth of other tools and equipment to service a wide array of customer needs.

Contact Peak today to discuss our equipment lineup and how we can be a resource for your joint testing or validation needs.

Peak Princ. Eng. John Medcalf featured in Assembly Mag

Peak’s Principal Engineer, John Medcalf, was featured in a recent article posted to the April 6, 2018 issue of Assembly Magazine, specifically discussing “Why Threaded Fasteners Fail.”

John contributed expert testimony on the subject as quoted:

“The geometry, material, heat treatment, finish and other factors all play into how well a fastener performs from assembly through the life of the product it is installed in,” says John Medcalf, principal engineer and business lead for Peak Innovations Engineering. “So always treat the fastener like the engineered component that it is.”

You can read the full article and all of John’s input on the topic here:

https://www.assemblymag.com/articles/94240-why-threaded-fasteners-fail

 

It’s 10 PM, Do You Know Where Your Fasteners Are?

While some fastener manufacturers and distributors get involved with customer product design, more often than not, they are simply asked to quote and provide a fastener to a print or industry standard.  A reputable supplier can ensure the specification requirements are met, but where it goes from there is all up to the customer or end user.  Any fastener industry veteran can tell you that nine times out of ten it is the joint design and installation technique, not fastener quality, that is at fault when a fastener fails.  So knowing how and where that Grade 8 bolt you sold is being used may not keep you up at night, but the ensuing argument when parts and assemblies start failing sure can.

Just as we know it’s a good idea to be aware of who and what our kids associate themselves with, keeping an eye on what joints your fasteners hang out in can help prevent them from getting involved with the wrong crowd.  For example, take the troublesome twins Over and Under Tightened and their common outcomes:

  • Over-Tightening
    • Fasteners more prone to failure under service loads.
    • Thread strip may occur undetected at assembly leading to a compromised joint.
    • Mating components may be damaged.
  • Under-Tightening
    • Insufficient clamp load cannot resist service loads, leading to joint failure.
    • Fasteners are more susceptible to fatigue failures.
    • Joints will be less resistant to vibration and other loosening effects.

Regardless of whether the fastener was at fault or not, you can be assured that any one of these situations will lead to a call from the customer.

At this point the reader is likely thinking, “These are the manufacturer’s responsibilities and within their control.  I can’t afford the resources to do my job the way I’d like to, never mind help with theirs.”   True as those points may be, trends show increasingly competitive markets with direct online purchases gaining traction and setting a low floor for prices.  Sitting pat isn’t a safe bet for a sustainable future.  Rather, providing supporting services your customer needs and recognizes will be an increasingly important differentiator.

While reacting to problems effectively when they arise can build credibility, the best way to add value and build customer loyalty is to proactively work to prevent them.  The problem is the vast majority of joints don’t cause problems that put your fasteners in a bad light, and joints that will become problems are rarely obvious until it’s too late. So short of hiring a team of experienced application engineers or enlisting a fastener consultant, what steps can be taken?

In short, get to know your customers and their products. It can be surprising how often a customer is just a name to a supplier, without much time invested into what they make and how and where their products are used.  And while spending some time on their website or talking with their fastener buyer is a good start, there is no substitute for time on the manufacturing floor.  Chances are, sales and vendor managed inventory reps are already visiting the facility, so make the most of their time there.  Some areas to investigate may include:

  • What is the nature of their assembly equipment?
  • Is there a dedicated rework area?
  • Are there scrap bins along the assembly line?

The supplier’s team in the office can participate as well:

  • Is there a new package of parts being quoted, possibly indicating a new product launch?
  • Is there a heavily revised fastener print, possibly indicating joining issues the customer is chasing?

It doesn’t take a lot of expertise to ask the questions, but having the capability to address any red flags that might be raised is another matter. Because fastening is an assembly process, understanding it requires applicable manufacturing experience combined with in depth knowledge of exactly how bolted joints work. This is an increasingly rare set of skills and experiences.  Whether to invest in resources capable of preventing or troubleshooting problems, or to contract that capability to others is a complex decision beyond the scope of this article.  It comes down to assessing whether in-house capability will provide sufficient competitive advantage relative to contracting the need to those few people that specialize in it.

 

Case Study:

Peak Innovations Engineering was contacted by a fastener manufacturer whose customer was reporting that a nut and bolt combination was stripping threads at the specified assembly torque.  Through the use of the latest ultrasonic measurement techniques, Peak performed an in-joint torque-angle to failure test while continuously monitoring bolt tension.  It was found that while thread strip was the failure mode, it occurred above the bolts’ yield point and well above the stated assembly torque.  With reliable data in hand, the fastener manufacturer was able to work with their customer on diagnosing the torque control problems at assembly.  The cost to the fastener supplier for this definitive test that provided quantitative answers to what was a stalemated disagreement was $3,475.  Both the supplier and their customer benefitted.

 

Figure 1 – Stripped Bolt & Nut Connection

 

Figure 2 – Ultrasonic Torque-Tension Test

 

In closing, companies like Amazon are changing the way products are bought and sold, and it is time for industrial distribution to take note.  To stay relevant, fastener suppliers will either have to find a way to be cost competitive with internet sales models, or offer services that truly add value to their customers.  With dedicated fastener engineers going the way of the dodo bird, finding a way to provide joint design, validation, and troubleshooting should be of particular value.  So ask yourself, do you know where your fasteners are?

Author: John Medcalf

**This article originally appeared in the Sept/Oct 2017 issue of the American Fastener Journal

It’s 10 PM PDF DOWNLOAD

Bolted Joints from End to End

Like most subjects, bolted joint development contains a number of important considerations, some of which are often overlooked. While estimating the bolt tension achieved for a given tightening strategy is certainly a common focus, the effect that load has on the joint components is less fully discussed and understood. One of the reasons threaded fasteners are so widely used is that they can generate a tremendous amount of clamp load in a small area. For example a ½ – 20 Grade 8 bolt can supply nearly 20,000 lb of clamp load in little more than a one inch diameter. Clamp load is central because it is the mechanism by which joint components can be held together without moving relative to one another – arguably the primary requirement of a structural joint. However, placing a large load on a small area creates a high level of stress, which in turn can lead to problems. A common analogy is the effect spike heels can have on wood floors. While technically accurate, experience shows that use of an alternative comparison is better for one’s health.

This article will examine the impact of bolt tension on the two opposing areas over which that load is reacted – under the head or nut and in the mating threads. For example, Figure 1 shows the FEA analysis of an automotive differential housing. In what is a common occurrence, the highest stresses act on the threaded holes.

Being focused exclusively on assisting clients with the development, testing and validation of bolted joints, Peak Innovations Engineering has found that an effective means of explaining joint behavior is to give examples of the effect of common design decisions. Test experience also provides insight into where actual results tend to differ significantly from calculations.

Under Head/Nut Area

With the exception of not testing to determine bolt tension and the resulting clamp load, exceeding the compressive yield strength of the material under the head or nut is arguably the most common deviation from recommended design practice. The most common target for bolt tightening is to achieve 75 % of the bolt’s proof load. Using a ½-13 hex head cap screw for illustration, Figure 2 summarizes the pressure that would be generated by that bolt tension on the mating material assuming a standard 9/16” clearance hole. It then compares that pressure to the estimated compressive yield strength of common steel clamp members with a range of hardness. Note that compressive yield strength is rarely specified, but tensile yield strength is a commonly accepted estimate for some common materials such as steel, though not others (notably all types of cast iron). To show how relatively small dimensional changes can have great impact on area, the same calculation is performed for a flange head screw. Use of flat washers is not included because the calculation would be dependent on the relative hardness of the washer, clamped member and fastener, as well as the ID/OD of the washer chosen. The take-away from Figure 2 appears to be that as long as Grade 8 fasteners aren’t used on materials softer than the Rockwall C scale, the area under hex head fasteners is sufficient. Unfortunately the actual pressure is highly uneven, and contact area is often less than predicted due to uneven surfaces. Therefore the maximum pressure actually present is much higher than the theoretical average value. Figure 3 shows scanned images of pressure sensitive film placed directly under the head of a hex cap screw and under a combination hex cap screw head and flat washer. The standard washer’s thickness is not nearly great enough to react loads evenly across its diameter. Note that a hardened washer of the same dimensions would behave the same way. Of course there are other reasons for the use of flat washers other than to spread bolt loads, such as providing a constant friction coefficient against varying clamps materials and finishes, elimination of galling or stick-slip, and to protect the integrity of the underlying finish.

While flange heads are stiffer than standard flat washers, a different variable is to be considered. The contact face is not perfectly flat, but instead generally slightly conical. This creates a different type of pressure gradient across the diameter, and one that is more difficult to predict. This is illustrated in Figure 4 showing photos of pressure sensitive film studies of flange nuts and screws. This review of ½” and M12 flange hardware was undertaken when a customer test revealed a flange head screw with a convex surface (contact at the edge of the clearance hole). An example is shown in the bottom image of Figure 3. In addition to increasing surface area, increasing the contact diameter of the fastener on which torque is applied increases the average radius on which resistance due to frictions acts. This reduces the bolt tension generated for a given torque. In this case the tension reduction of substituting flange head for hex head fasteners is approximately 12%.

Figure 5 extends this example of uneven clamp load by showing the pressure distribution across two 2.25” dia. x 2 “ long cylinders clamped by a ½”-13 screw and nut. Even at this thickness the peak pressure is still 40% greater than would be predicted by dividing the bolt tension by the contact area at the joint faces.

This discussion is not intended to imply that compressive yield (commonly referred to as embedment) is to be avoided at all times. In many instances the reduction in clamp load required to prevent embedment would be more detrimental to reliability than the relaxation caused by yield of high stress areas. It is suggested that this is an area that deserves more attention, through preliminary calculation and subsequent testing. Some situations that require particular attention are:

* Joints with significant embedment that are regularly serviced. Subsequent bolt installation may have greatly reduced and uneven contact areas, leading to increased potential for fatigue failure due to bolt bending and clamp load loss

* Joints that combine high axial loads and bolts with small length-to-diameter ratios. In joints where compressive yield is imminent, the additional bolt load can be great enough to create much greater clamp load loss than the relaxation which regularly occurs due to localized yield immediately after tightening.

* Joints operating at temperatures significantly different that those at which tightening occurred that also have clamped members with different rates of thermal expansion than the bolts. A common example is engine bolts in aluminum castings. At operating temperature the greater expansion of the aluminum castings can results in clamp load loss through either embedment, bolt yield or thread yield.

Thread Area

The fundamental issue in reacting bolt tension within the mating threads is the same as those discussed in the area under the head or nut face; that a large load must be dissipated over a small area. Two factors make the threaded area potentially more problematic. First is the fact that the internal threads are sometimes provided as separate standardized elements (nuts) and sometimes by the manufacturer’s design (tapped holes). Secondly, the forces in the threads don’t act normal to the mating surfaces, as under the head and nut face. The triangular thread form results in forces that both compress and expand the internal threads. This effect can be seen in Fig 6, an FEA screenshot of stresses on a typical nut and mating threads. As with the load under the bearing surfaces, load along the length of threads is not uniform. Studies show the first engaged thread absorbing about one third of the bolt tension, decreasing until effectively all the load is absorbed by the 6th thread. This should raise the questions as to the reason for some tapped holes having significantly more than six threads. Because the material into which holes are tapped is often lower strength than the mating screw, that material can yield before the screw does. Looking at Figure 6, one can imagine if the surface of the most highly stressed thread crushed even minutely, load would then be transferred down the line of threads, causing partial

leveling of individual thread stress. This is how additional threads are engaged. To a lesser degree this effect is present in standard nuts properly matched with the mating screw. Nuts are specified to have slightly lower yield strength that the matched screw to take advantage of this effect. It is important to note that this doesn’t mean the nut is the weaker of the two elements, as its height is established so that there is enough thread area that the bolt will fracture before the nut threads suffer noticeable damage. The width or diameter of the nut also plays a part in its load capacity. Figure 6 shows the nut threads expanding away from the screw as mating threads slide relative to one another radially. This reduction in thread engagement, and therefore load capacity, is a function of the nut member’s stiffness in the radial direction. Width across the flats of standard nuts is about 1.5 to 1.6 to times the nominal thread diameter. This is actually a compromise between strength and size as a ratio of nearly 2:1 is required to eliminate dilation. Nuts are available in larger widths and lengths when required. A much more isolated, but related effect are specialty nuts, such as those used in aerospace, which have a thick flange to minimize dilation, but have a thin wall above the flange to save weight. This often results in axial compression of the nut body prior to bolt yield.

The most common design challenge on the threaded end of a bolted joint is determining the required thread engagement when using tapped holes. As with nuts, the objective is to ensure that the failure mode in the event of over-tightening is by bolt fracture rather than thread strip. This mode is preferable as it is more obvious (stripped threads don’t generate a loose bolt) and repair is generally less expensive and more dependable. Figure 7 summarizes estimated thread engagement relative to the nominal fastener diameter for a number of common nut member materials. Ultimate shear strength, the material property required to estimate required thread engagement, is generally only available for common materials. It is often estimated as a percentage of ultimate tensile strength, though. While Figure 7 is based on calculations for a single thread size, the L/D ratio remains within about 5% across the range of fastener diameters except at the small end (under ¼” or M6). This calculation doesn’t account for common features, such as:

* The incomplete threads at the screw tip in blind holes. This can reduce capacity by 20% in hard materials that require only short thread engagement.

* The chamfer added to the hole entry after tapping. This is often incorrectly included in thread engagement. Because it is often not seen as an important dimension, and is difficult to measure accurately, this feature is often not well controlled. As with incomplete threads, the impact is greater the smaller the L/D ratio.

* Because the radial dimensions of thread engagement are quite small, small dimensional changes in either internal or external threads can have a measurable impact on load capacity. For example, if all the dimensional and material tolerances on standard threads are at a worst-case condition, failure mode can change bolt fracture to thread strip. Dimensional variation is more common in tapped holes, particularly by those produced by lower volume production methods, than in standard fasteners.

In summary, the high stresses produced in bolted joints magnify the effect of “theory vs. reality”. While increasingly powerful analytical tools reduce development time and cost, physical testing of bolted joints is essential in avoiding costly failures.

Author: Dave Archer

 

***This article also appears in Assembly Magazine under the title “Analyzing Bolted Joints for Clamp Load and Joint Stress”  You can find that article here: http://www.assemblymag.com/articles/93778-analyzing-bolted-joints-for-clamp-load-and-joint-stress

 

In Response to the 2015 Bolt Failure on the Lemnhult Wind Farm

On Christmas Eve 2015 a 390 ft. tall wind turbine located in the Lemnhult wind farm near Vetlanda Sweden collapsed when the bolts in the tower’s lowermost flange joint, containing one hundred M64 bolts, failed. The incident came into the news recently with the release of the Swedish Accident Investigation Authority’s final report.

Based on a short summary that was made available in English, the failure investigation yielded the following conclusions

  • The bolts failed in fatigue, due primarily to wind forces. The root cause of the bolt fatigue was insufficient clamp load (bolt tension).
  • There was no check that the desired clamp load had been achieved.
  • The torque tool was not adequately maintained
  • The operator was inexperienced and not adequately trained
  • Rain during installation altered the friction at the bolt’s mating surfaces, in turn decreasing the clamp load in the joint
  • The municipality and the control manager limited inspections to the foundations
  • The wind farm had previous problems with loose or broken bolts that were not reported to authorities
  • The recommendations resulting from the investigation focused primarily on documentation, inspection and compliance with pertinent regulation.

From the experience of conducting hundreds of tests on bolted joints over the last fifteen years, I recognized that many important elements of this case are common to bolted joint problems across all industries.

Fatigue fracture is the most common cause of sudden catastrophic failure of bolted joints. Fortunately the appearance of the fracture surfaces is distinctive, and metallurgical tests to confirm the visual diagnosis are widely available. But what initiated the cracks that ultimately led to failure? It is common to conclude that the bolts were not to specification, but that is rarely the case.  Much more common, and as was determined in this case, is that the force with which the joint is clamped together is not enough to handle the external loads acting on it.  In this case the primary external force is created by the wind causing the 400 ton wind turbine nearly 400 ft. in the air to sway and put periodic bending loads on the bolts.

So if the bolts are as strong as expected, were the forces acting on the joint greater than projected?  Manufacturers of highly engineered products that carry high costs associated with failure or poor performance, like these turbines, generally do a good job of estimating these loads and apply a factor of safety to allow for uncertainty.

If the manufacturer does a good job estimating the external loads why would they design a joint that isn’t strong enough to handle them?  Whenever we are speaking of a bolted joint, the answer is usually that their calculations are based on a critical assumption that is difficult to confirm in practice.   That assumption is how much force, or clamp load, did the bolts generate to hold the joint together.  It is important to understand the difference between how strong the bolt is and how much clamp load it creates.  Bolts have specified and easily verified requirements for strength; that is how much force it takes to pull it apart. However by definition fatigue failure occurs by repetitively loading and unloading a component (the bolt) to failure at forces below the bolt’s ultimate strength.  Think of bending a paper clip back and forth until it breaks.  You aren’t nearly strong enough to pull it apart, yet you were easily able to break it.  This wind turbine collapsed because the tower bolts failed like the paper clip did – just with much less bending over a much greater number of cycles.

But what does clamp load have to do with the paper clip example, and what exactly is clamp load?

Clamp load is created when a bolt is stretched.  You can rotate a bolt into a nut or tapped hole by hand until the plate between the head and threads stop it.  Then, with the assistance of a wrench you can continue to the rotate the bolt.  Because the threaded end advances into the mating threads, as long as the plates in between aren’t being crushed at the same rate that the end of the bolt is advancing, then the bolt has to be stretched.  Like a rubber band restricts a deck of cards from sliding back and forth, the tension in the stretched bolt creates a clamp load in the plates between the bolt head and the threads.  A sufficient level of clamp load prevents fatigue failure by preventing two related events. First, clamp load prevents the clamped components from sliding relative to one another. Relative movement of structural joint components will eventually lead to failure in the vast majority of cases.  Most of us have seen, and an unfortunate few have experienced, the results of insufficient clamp load applied to wheels on a car, truck or trailer.  If the wheel is not clamped securely to the hub, the resulting movement causes the studs to bend back and forth by a small amount as the wheel rotates, eventually fracturing them and allowing the wheel to separate from the vehicle.  In many cases of wheel separation some studs are still intact. Relative movement not only causes failure of the studs but also loosening of the lug nuts.  The nuts often back off completely, so in some cases before all the studs fracture the wheel slips over the end of the remaining studs.

The second and more difficult to imagine risk of low clamp load that the increase in stress acting on the bolts from external loads is dependent on the clamp load in the joint. For example, if instead of wind loads acting parallel to the ground, imagine King Kong grabbing the turbine and pulling straight up on it again and again in an attempt to pull it out of the ground.  With each pull a portion of that force is transmitted to the bolts.  In what is a counter-intuitive phenomena, the more the bolt is stretched at assembly (producing more clamp load) the smaller the portion of external pulling load is transmitted to the bolt.  Reducing the min-max range of loads acting on components is key in reducing the likelihood of fatigue fracture.

So why is estimating clamp load generated at assembly so difficult? In most cases it is simply equal to
the tension created when the fastener is elongated during tightening. Therefore it is a common misconception that an accurate torque wrench will produce an accurate estimate of clamp load.  This is far from a true statement because the torque applied to a bolt is actually just a measure of the resistance to turning it.  If all that torque went into tensioning the bolt the relationship between torque and the initial bolt tension could be fairly accurately calculated.  However all but about 10-15% of the input torque is used to overcome friction under the head or nut and in the mating threads.  The tightening process is very inefficient.  As friction is the source of this inefficiency, many assume that an accurate measure of the coefficient of friction (COF) would solve the problem. Because one can readily find COF tables on-line this would seem a straight forward proposition.  However the various geometries, surface finishes, coatings, and lubricants in bolted joints represent a far more complex system than the two flat surfaces of raw material sliding against one another upon which COF calculations are based.  In fact the torque tables that would pop up in a Google search of “how much torque do I need for a ¼-20 bolt” are based on modified friction factors rather than COF, and are still inadequate for critical joints such as the one in question. Why? Because saying that the torque for all zinc-plated bolts can be defined by using a friction factor K of 0.20 would be the equivalent of instructing a house painter to use the color blue. There are too many variables in play to make a decision without testing the specific application to determine the most successful result.

Recent advances in ultrasonic test equipment has made the previously unattainable goal of accurately measuring bolt tension without changing the joint or fastener or adding other components a reality.  The ultimate goal of tightening standard bolts directly to tension in a production environment is not yet a reality, however the relationship between torque, angle of rotation and bolt tension can be determined through testing unmodified production hardware.  From this information various tightening strategies can be compared to achieve the one that will provide the desired clamp load with minimal scatter.  The technique is also capable of monitoring changes due to time and temperature. This is important because being able to accurately determine clamp load at assembly doesn’t mean one is home free. The clamp load that existed in the tower joint when it failed was not the same as when the bolts were originally tightened.  All bolted joints relax due to local yielding and/or material creep.

While certainly easier to list than to implement, it is always worth reviewing the three fundamental requirements of assembling structural bolted joints.

  1. How much clamp load is needed?
  2. What process of bolt tensioning (tightening) will consistently achieve it?
  3. How can we verify we’ve achieved the desired clamp load, either by direct measurement or by process control?

If your answer to the first question is “I don’t have a number, but what we’re doing now is working”, I suggest performing a test to measure what that number is and how much it varies. Then you will have a basis for comparison when what you’re doing suddenly stops working.

You can’t develop the most effective tightening process unless you are aware of what tools and techniques are available. There tends to be too much focus on the tool and not how it’s utilized and the tightening strategy applied. After some fairly rudimentary process development, a target clamp load can be much more accurately controlled with a $10 crescent wrench than a $20,000 programmable DC tool calibrated to ensure that the stated 0.25% accuracy is applied to an installation torque that is taken from a reference table.

The method for verifying the clamp load is obviously very much dependent on the tool and tightening process used. However, one widely applicable observation is that unless it is obviously damaged or malfunctioning, a tool is more likely to be a major contributor to not achieving desired clamp load as a result of it being setup or programed incorrectly than it being out of calibration.  For example, in our investigation of a wind turbine gear joint problem where bolts were occasionally being tensioned to failure at assembly, ultrasonic bolt tension measurement showed the presumed torque vs. bolt tension  relationship needed adjustment.  However, another major contributor to the bolts failing was that the hydraulic torque wrench used for tightening had the wrong factor converting hydraulic pressure to the torque selected.  The tool had previously been shipped back to the manufacturer who correctly found it to be in specification.  This isn’t suggesting that calibration is a waste of time but that, as in this case, valuable time is often expended on an area that is generally not a root case.  This example also illustrates a strategy for improving the effectiveness of calibration.  The more closely calibration can be performed in a condition that reflects actual use, the more accurate and valuable it will be.  Calibrating the system rather than the sensor is usually more challenging but always beneficial.

In summary, a review of the Authority’s findings shows that the first two bullet points encapsulate what happened and why. If the clamp load could have been reliably determined, the failure would not have occurred and the report would not have been requested.  This direct relationship between cause and effect cannot be achieved by other corrective action, such as improved inspection, training and documentation.

By: David Archer

 

To view the accident report from the Swedish Accident Investigation Authority click here:  http://www.havkom.se/en/investigations/vaegtrafik-oevrigt/olycka-med-vindkraftverk-i-lemnhult

*** Pictures are from the report

 

 

Part 2: The Three Least Understood Truths in Bolted Joint Design

By: David Archer

In Part 1 of this article, we discussed the tension of structural joints and external tensile loads applied to joints and the effect on bolt tension. The third least understood truth, sizing fasteners, will be revealed in this final installment. What follows is more of what you always wanted to know about bolted joints but were afraid to ask!

3. When sizing a fastener for a given application, the saying, “When in doubt, make it stout,” is a guideline you should throw out.

Typically, unless one is designing products where the cost of failure is high, the detailed analysis and testing required to determine joint loading is not performed, and therefore the quantitative basis of how to size fasteners is limited. Even when the best analytical and test techniques are used, there is always uncertainty as to how well the analytical assumptions or the test protocol replicated actual use conditions. It is therefore common to apply some type of safety factor to the assumptions that go into threaded fastener sizing. While they often produce the same result, in some cases the engineer applies a factor to the assumed loads from which bolt sizes will be calculated, while in others the engineer will simply use the next biggest fastener. Often the latter approach is used when there is no load data available and the joint is being designed based on a product currently in the field. In either case, on the surface it seems a reasonable assumption that if big is good, bigger is better and safer. Actually the opposite is often true.

The primary methods of varying bolt strength are its size and the material from which it is made. For example, a Grade 8 bolt obviously has greater capacity than a Grade 5 bolt of the same size (in the metric system, property classes 10.9 and 8.8 are the near equivalent of Grade 8 and Grade 5, respectively). But what if we don’t need greater bolt strength and simply want to optimize the joint at the existing strength. One option would be to use a higher grade bolt than planned, but reduce its size to yield the same capacity. This is the option we will explore in Table 1. The contents of Table 1 reinforce the axiom that one should strive to use the smallest diameter fastener with the largest L/D ratio possible.

 

 

Part 1: The Three Least Understood Truths in Bolted Joint Design

In fastening, as in most endeavors, the most common misconceptions relate to the most fundamental principles. These errors are commonly made simply because their universal nature provides the greatest opportunity for making them. For example, the reason that most auto accidents occur within a small radius of the driver’s home isn’t because those streets are the most dangerous, but because they are the ones that the driver travels on most often. Similarly, one of the main reasons these fastening misconceptions exist so widely is because they are so basic to the use of threaded fasteners. Therefore, those with less experience assume it is common knowledge. What follows is hopefully some of what you always wanted to know about bolted joints but were afraid to ask.

  1. Tension, rather that torque, is the quantity that should determine how “tight” a structural joint should be.

The tensile capacity, and generally the longevity of bolted joints in structural applications, are determined by the force with which the bolt “squeezes” the components being secured, referred to as the clamp load. The pressure exerted by the joint on the bolt, an equal and opposite reaction to the clamp load, is the bolt tension. Tension is generated in the bolt when one set of threads is turned relative to another set. This movement wants to shorten the distance between the bearing surfaces of the two parts (usually the face under the nut and bolt head), but the stack of components between those faces provides a great deal of resistance to allowing the bearing faces to move closer to one another. So the faces stay in about the same relative position, and the length of fastener that lies between them stretches instead – generating both tension in the bolt and the mating clamp load on the joint. How does this tension relate to the torque needed to rotate the fastener? Actually, it’s a direct relationship most commonly expressed as T = KDF, where T, K, D and F are the torque, nut factor, nominal bolt diameter and the force (or tension) respectively. So for a given bolt, torque is directly proportional to tension, with a factor K (often referred to as the nut factor or K factor)- the only variable that must be known to be able to calculate torque from tension, or visa-versa. Unfortunately, this K factor is a compilation of several friction components which vary with materials, finishes, pressure, geometry, and the speed of relative motion. It is therefore quite difficult to predict (even varying during the tightening), so assigning a value for a given joint can’t be done accurately without experimentation.

So why not measure tension directly during assembly as a way to determine how tight is right? This is in fact a possibility and is done in limited cases. More often, the torque vs. tension relationship is established in the lab with the results used to establish target torque. However, bolt tension is more difficult and expensive to measure than the torque applied to generate that tension. So the most common method of testing involves tightening the bolt just beyond its yield point in the joint. At yield, bolt tension can be estimated based on the fastener’s proof load (similar to yield strength). This provides a method of estimating the relationship between torque and tension without directly measuring tension.

The take-away from this discussion is that while tension is the quantity of interest, torque is an easier to measure alternative, directly related to tension through a difficult to measure “K” factor. Therefore, the accuracy with which one can estimate tension from tightening torque can vary widely, with errors often ranging from 10% to 50% or more. It is important to note that this error has nothing to do with the accuracy of the torque wrench itself. The tool error is simply the uncertainty of how close the wrench’s torque reading is to the actual torque.  The K factor uncertainty represents how well the input torque can determine the bolt tension. This error is in addition to the tool error and is generally much greater. For example, a torque wrench can have less than 1% error but not be able to tighten bolts within 50% of their target tension.

  1. When an external tensile load is applied to a joint, that increased load is not directly added to bolt tension.

In simple terms, the bolt does not “feel” the entire external load. In fact, in most cases the vast majority of this additional load is absorbed by the clamped joint members because the joint is usually much stiffer than the bolts. This is the reason bolts tensioned at or near yield can often resist significant external tensile loads, such as the head bolts in your car’s engine.  At some point, however, the external load can be great enough that the clamp load on the joint is unloaded completely. Any additional load from this point will be entirely additive to the existing bolt tension. The balance between the forces acting on the bolt and the joint are generally displayed on a graph called a “joint diagram.” However, these diagrams aren’t clearly understood without some explanation and study, so we put together what we thought was a more easily understood illustration.

Using a component from an assembly on which we recently conducted some torque-tension testing, we performed a simple test showing the impact of external tensile loading on the bolt tension of a pre-loaded bolt. As shown in Figure 1, we modified the existing assembly by replacing the eight 5/16 – 18 socket head cap screws with four Grade 5 hex head cap screws. These fasteners were prepped with ultrasonic sensors to allow real-time measurement of blot tension in the joint. Adapters allow the test assembly to be mounted into a tensile tester.

The bolts were tightened to a target preload of 2,500 lbs per bolt. Load was then applied to the joint via the tensile tester, while tension in one of the four bolts was monitored as shown in Figure 2.  Immediately evident is the change in rate of increasing bolt tension. While the load is applied uniformly, the bolt tension initially rises very gradually, with the slope increasing a bit until an inflection point where it rises steadily in a linear fashion. That point, noted on the graph as Point A, is the load at which the joint is unloaded completely and 100% of the additional applied load goes directly into increasing bolt tension (sometimes called the “critical load”). That inflection point was estimated to occur at an applied load of 12,364 lb and a bolt tension of 2,801 lb.

So, up to Point A, an external load of 3,091 lbs per bolt (12,364/4) resulted in only a 261 lb (2,901-2,501) increase in bolt tension. After Point A, the additional bolt tension should equal the added external load on a per-bolt basis. The calculations are close to what was expected, as the additional 389 lb load per bolt {(13,918 – 12,364)/4} results in 421 lbs of increased bolt tension. Selecting the precise beginning of the linear potion of the tension plot is likely the primary source of error. A question concerning Point A that may arise is why the critical load does not equal the total clamp load that was initially applied (12,364 lbs vs. 10,410 lbs). That difference indicates that the joint is stiffer than the bolts, and actually allows us to calculate the relative stiffness. The 12,364 lb critical load is 18.8% greater than the 10,410 lb pre-load. Through analysis, it can be shown that the ratio of the bolt and joint stiffness equals the fractional increase in load required to unload the joint relative to the initial preload. Therefore, in this case the joint is estimated to be 5.3 times as stiff as the bolts (1/0.188=5.3).

Watch for Part 2 of this article, where we will discuss sizing fasteners and the impact of substitutions.

Did you know…?

  • The one-piece carbon fiber fuselage of the Boeing 787 Dreamliner eliminates the need for 40,000 to 50,000 fasteners that would have been required with a typical aluminum design.
  • Products regularly submitted for third-party testing and certification are subject to recall about one-third as often as those that are not.
  • There are approximately 3,500 spot welds in the typical automotive sedan.
  • An investigative panel determined that the root cause of the fire that doomed the maiden flight of the Falcon I rocker that was a fuel leak caused by a corroded aluminum nut. The Falcon I is a low-cost rocket system developed by SpaceX, a company privately funded by PayPal founder Elon Musk.

Peak innovations Engineering Receives its Renewal Accreditation ISO/IEC 17025: 2005 from A2LA

Peak innovations Engineering is pleased to announce that they have been granted ISO/IEC 17025: 2005 accreditation by the American Association for Laboratory Accreditation (A2LA), after a thorough assessment and review of its quality management system and competence to perform specific tests.

This achievement of A2LA accreditation demonstrates an organization’s competence to manage and perform the activities defined by its A2LA Scope of Accreditation 2511.01.  These activities include various methods of torque and torque – tension testing of fasteners.

“Peak innovations Engineering’s value proposition is ‘providing the ultimate confidence in fastening that lasts a lifetime.’  Ongoing assessments and accreditation from A2LA are central to providing that confidence in our methods and test results.  We are proud to demonstrate that the same level of service provided by Archetype Joint in Orion, Michigan will now be provided by our new location in Machesney Park, Illinois.”  -John Medcalf, Engineering and Business Lead

ABOUT A2LA:

A2LA is the largest U.S.-based, multi-discipline accreditation body with over 35 years of experience providing internationally recognized accreditation services and quality training.  A2LA’s world-class accreditation services encompass testing and calibration laboratories, medical testing laboratories, inspection bodies, proficiency testing providers, reference material producers and product certification bodies.  Organizations are accredited to international standards and field-specific requirements developed with government and industry collaboration.  A2LA also offers a wide variety of both public and private on-site training programs to complement the various accreditation programs.