The following definitions describe the standard design features of screw threads in general. Refer to Figure 1 for identification of the screw thread terms.
Screw Thread- A ridge of uniform section formed as a helix on the external or internal surface of a cylinder or cone.
External Thread- Threads produced on the outside surface of lengths of round stock.
Internal Thread- Threads produced inside bores stock such as those found in tapped holes and nuts.
Thread Profile- The configuration of the thread in an axial plane with its three essential parts being flanks, roots, and crest.
Figure 1: Screw Threads
Flanks- The sloped surfaces joining crests and roots.
Root- The bottom of the thread section.
Crest- The top area of the thread ridge.
Complete Thread (Full Form)- A thread having a full form at both its crest and root.
Incomplete Thread- Threads having crests or roots not fully formed. Incomplete threads occur at the end of pointed externally threaded products, at countersinks in the faces of threaded holes or nuts, and at thread run outs where the threaded section blends into the unthreaded shank.
Major Diameter- On an internal thread, the diameter at the root; on an external thread, the diameter of the thread at the crest.
Minor Diameter- On an internal thread, the diameter at the crests; on an external thread, the diameter at the root.
Height of Thread- The height (or depth) of a thread is the distance between the major and minor diameters.
Thread Pitch- The distance measured parallel to the thread axis between corresponding points on adjacent threads. Pitch is the reciprocal of the number of threads per inch (TPI). Unified threads are designated in threads per inch. Metric threads are designated by their actual pitch. Figure 2 shows two methods of determining threads per inch (TPI), or pitch.
Figure 2: Determining Pitch or TPI
Threads Per Inch (TPI)- The number of threads per inch (TPI) is the reciprocal of the pitch in inches.
Thread Series- Thread series are groups of diameter/pitch combinations distinguished from each other by the number of threads per inch applied to specific diameters. Examples include Unified Coarse, Unified Fine, and Metric Coarse.
Left-Hand Thread- A thread is left-handed if, when viewed end-on, it winds in a counter-clockwise and receding direction. All left-hand threads are designated "LH".
Right-Hand Thread- A thread is right-handed if, when viewed end-on, it winds in a clockwise and receding direction. A thread is considered right-handed unless specifically indicated.
Rule of Thumb- When viewed horizontally, the thread will slope down toward the hand of the thread. For example: down to the right for a right-handed thread.
Single and Multiple Threads- Most screws, bolts, studs, tapped holes, and nuts have single threads; the single thread has a single ridge in the form of a helix, as shown in Figure 3. One usually assumes, unless the thread is designated otherwise, it is a single thread.
Figure 3: Single Thread
The lead of a thread is the distance traveled parallel to the axis in one rotation of a part in relation to a fixed mating part. In single threads, the lead is equal to the pitch. A double thread, as shown in Figure 4, has two ridges starting 180 apart. The lead is twice the pitch. A triple thread, as shown in Figure 5, has three ridges starting 120apart. The lead is three times the pitch.
Figure 4: Double Thread
Figure 5: Triple Thread
Thread Fit- The combination of allowable allowances and tolerances in mating threads is called thread fit and is a measure of the tightness or looseness between them.
Allowance- An allowance is an intentional clearance between mating threads. When both external and internal threads are manufactured to their permitted maximum material condition, there will be a finite space between them. Both the Unified and ISO systems of thread standards apply allowances to the external thread.
Tolerances- Tolerances are amounts by which dimensions are permitted to vary to convenience manufacturing. The tolerance is the difference between the maximum and minimum permitted size. For an external thread, its maximum material condition minus its tolerances defines its minimum material condition. For an internal thread, its maximum material condition plus its tolerances defines its minimum material condition.
Multiple Threads are used where fast movement is desired with a minimum number of rotations, such as in aworm geararrangement. The mechanical advantage of multiple threads is less than the mechanical advantage generated with single threads.
Figure 6: Uniform Standard Thread Form
In practice, the amount of rounding varies as a result of tool wear. Unified standards are established by various thread series, which are groups of diameter-pitch combinations distinguished by the number of threads per inch applied to a specific diameter.
Fit is defined as "the measure of looseness or tightness between mating parts." The fit of a screw thread is the amount of clearance between the screw and the nut when they are assembled together. Classes of fit are specific combinations of allowances and tolerances applied to external and internal threads. For Unified inch series, there are three classes of fit for external threads and three for internal threads.
The three classes of fit for external and internal threads are allclearancefits. This means that they assemble without interference. The higher the classes number, the tighter the fit.
Internal and external members of corresponding classes generally will be assembled together, but parts made to different classes will be interchangeable and may be interchanged to obtain intermediate grades of fit. For example, a class 2A external thread could be used with a 1B, 2B, or 3B internal thread.
Class 3A and 3B threads have no specific allowance and are manufactured to restrictive tolerances. These classes of threads are intended for exceptionally high-grade commercial products, such as socket cap screws, set screws, aerospace bolts and nuts, and connecting rod bolts, where close or snug fit for precision is essential, as well as in applications where safety is a critical design feature.
Tolerances in the thread classes are a function of pitch, fastener diameter, and length of engagement for Unified threads with some overlap between the three combinations of classes. This should be considered when planning the manufacture of threads.
It is also of interest to note that regardless of thread class, working tolerances are extremely small. For example, the tolerances for 2A and 3A classes on a -inch fastener are 0.0065 inch and 0.0048 inch respectively. Values of tolerances usually are provided in detailed fastener design manuals. Also of importance is how maximum material condition is defined as the largest bolt or smallest permissible bolt thread diameter and the largest tapped hole in the nut.
Information regarding thread specifications, whether external or internal, is expressed in this order:
Figure 7 shows an example of thread specifications for the manufacture of threads on a fastener. The drawing provides all of the necessary thread specifications required to produce a thread. For external threads, the length of thread is given as a dimension on the drawing. The length given is to be the minimum length of full thread.
Figure 7: Inch Thread Specification
For threaded holes that go all of the way through the part, as in Figure 8, the term "thru" is sometimes added to the note. If no depth is given, the hole is assumed to go all the way through the part.
Figure 8: Inch Thread Specification for a Through Hole
For threaded holes that do not go all the way through the part, as in Figure 9, the depth is provided in the note; for example 1/2 - 13 UNC - 2B X 3/4 deep. The depth given is the minimum depth of full thread.
Figure 9: Inch Thread Specification for a Blind Hole
Figure 10: Basic M-Profile Metric Thread
The designations, terms, etc., used in the new metric standard differ in many respects from those used in the familiar Unified standard. Because the Unified standard incorporates terms and practices of long standing with which mechanics have knowledge and experience, an explanation of the M Profile metric system may best be accomplished by a comparison of metric M Profile standards to Unified standards.
A major difference between the standards is in the designation used to indicate the thread form. The Unified standard uses a series of capital letters, which not only indicate that the thread is of the Unified form, but also indicate if it is a coarse (UNC), a fine (UNF), an extra-fine (UNEF), or a constant pitch (UN), which is a series of pitches used on a variety of diameters. In the metric M Profile standard, the capital letter M is used to indicate the thread is metric M Profile with no reference to pitch classification.
A second major difference is that the Unified standard designates the thread pitch in terms of the number of threads in one inch of thread length. The metric M Profile standard states the specific pitch measurement, i.e., the dimension in millimeters from the centerline of one thread to the centerline of the next thread.
A third major difference is that the Unified system establishes limits of tolerance called classes. Classes 1A, 2A, and 3A apply to external threads only; Classes 1B, 2B, and 3B apply to internal threads only. Classes 3A and 3B provide a minimum, and Classes 1A and 1B a maximum. In the metric M Profile standard, the tolerance designation uses a number and a letter to indicate the pitch diameter tolerance, as well as a number and a letter to designate crest diameter tolerance. This results in four symbols for tolerance designation rather than the two used in the Unified system. In the metric M Profile standard, the lowercase letter "g" is used for external threads, and the capital letter "H" is used for internal threads. The numbers 4, 5, 6, 7, 8 are used to indicate the degree of internal thread tolerance; the numbers 3, 4, 5, 6, 7, 8, indicate the degree of external thread tolerance. Thus, two nominally similar threads are designated as follows:
3/8 - 16 - UNC - 2AM10 X 1.5 - 6g6g
Common shop practice in designating Unified screw threads is to state only the nominal diameter and the threads per inch. For example, 3/8-16 for a coarse thread, and 3/8-24 for a fine thread.
Most mechanics recognize the thread series by the threads per inch number. As to the thread class symbol, it usually is assumed that if no class number is stated, the thread is a Class 2A or 2B. This is a safe assumption because Classes 2A and 2B are commonly used tolerances for general applications, including production of bolts, screws, nuts, and similar fasteners. When required to produce screw threads, most mechanics use commercial threading tools (i.e., taps and dies) or, if chasing threads, commercial gauging tools for size checking. These commonly used tools are made to Class 2A or 2B tolerances; therefore, the mechanic depends on tool manufacturers to maintain screw thread accuracy. Because manufacturersproducts have extremely close tolerances, this practice is highly satisfactory and relieves the mechanic of practically all concern for maintaining sizes and fits within stated tolerances.
To reach this same condition was one of the objectives during efforts to establish metric M Profile standards. Because metric threads are in wide use throughout the world, a highly desirable condition would be one in which the standard was compatible with those in world use. As the International Organization for Standardization (ISO) metric standard was, for all practical purposes, recognized as the worldwide metric standard, the metric M Profile standard is in large measure patterned after it. It has a profile in basic agreement with the ISO profile and features detailed information for diameter pitch combinations selected as preferred standard sizes.
Metric M Profile screw threads are identified by a letter (M) for thread form, followed by the nominal diameter (major diameter) size and pitch expressed in millimeters, separated by the sign (x), and followed by the tolerance class separated by a dash (-) from the pitch. For example, a coarse pitch metric M Profile thread for a common fastener would be designated as shown in Figure 11.
Figure 11: External Thread M Profile, Right HandThe simplified international practice for designating coarse-pitch ISO screw threads is to leave off the pitch. Thus, a M14 x 2 thread is designated just M14. In the ANSI standard, to prevent misunderstanding, it is mandatory to use the value for pitch in all designations. Thus, a 10mm coarse thread is designated M10 x 1.5. When no tolerance classification is stated, it is assumed to be classification 6g6g (usually stated as simply 6g), which is equivalent to the Unified classification 2A, the commonly used classification for general applications.
The standard metric screw thread series for general-purpose equipments threaded components and mechanical fasteners is a coarse thread series. The diameter pitch combinations selected as preferred standard sizes in the coarse pitch series are listed in Table 1. These combinations are in basic agreement with ISO standards.
The metric M Profile designation does not specify series of diameter/pitch combinations as does the Unified system (i.e., coarse, fine, etc.). Although no indication of such grouping is given in the designation of a metric thread, series groupings are recommended. The coarse pitch series of diameter/pitch combinations shown in Table 1 is described as standard metric screw thread series for general purpose equipments threaded components and mechanical fasteners.
In preparation for cutting internal threads, a hole must be drilled slightly larger than the threads minor diameter. The reason for the oversized hole is to provide clearance between the wall of the hole and the roots of the tap threads. This gives chip space and allows free turning of the tap, reducing the tendency for the threads to tear. A practice of long standing when tapping Unified screw threads has been to make the hole 25 percent larger than the minor diameter. This results in an internal thread that is 75 percent of standard. The 25 percent that are missing from the crest do not appreciably reduce its strength. Unified thread tap drill charts listing tap drill sizes are a common shop convenience. Most carry this notation: based on approximately 75% of full thread.
The metric tap drill chart (see Table 3) lists both coarse and fine diameter/pitch combinations, as well as the recommended drill size in millimeters, to produce approximately 75 percent internal metric threads. The values in the table result from the following formula: Tap Drill Size = Major Diameter Pitch.
The raw material establishes the base property for the fastener, while the final combination of mechanical properties of the fastener is obtained during the manufacturing and post-treatment process.
Tensile strength is best defined as a measure of the ability of a material to withstand a longitudinal stress, expressed as the greatest stress that the material can stand without breaking.
Tensile strengths normally are expressed in terms of stress-pounds per square inch (psi) for inch fasteners and megapascals (MPa) for metric fasteners.
To convert stress to a load value expressed in pounds, simply multiply the stress (psi) by the stress area of the thread. Some data for stress areas is provided as an example in Table 4 and Table 5 below.
In wedge tensile testing, a hardened washer with a beveled surface is placed directly under the head of the fastener to be tested.
As the load is applied, the resulting wedging action induces a severe bending stress that is concentrated at the junction of the shank to the head. To be acceptable, the fastener must support, without fracture, its specified minimum tensile strength. The fracture must occur at a location other than the junction of the shank to the head. The test is intended to demonstrate the fasteners ductility quality and structural rigidity at the junction. Figure 12 shows an example of a wedge tensile strength test.
Figure 12: Wedge Test
It is difficult to test full-size fasteners for yield strength because of the different strain rates in areas such as: the fully threaded portion; the thread runout; and the unthreaded shank, which comprises the stressed length. Because of this, the "Proof Load" system was introduced as an approved technique for testing fastenersdeformation characteristics.
Proof load represents the maximum useable load limit of the fastener for many design/service applications. Proof load is commonly defined as the tension-applied load that the fastener must support without evidence of any deformation. Often, proof load and yield strength are interpreted as being the same.
Proof load is a force measurement; the units are pounds or Newton's. Yield strength is a stress measurement; the units are PSI or MPa. The stress at the proof load is 90-93% of the yield strength.
The proof load testing procedure involves measuring the overall length of the fastener, applying and releasing the specified proof load, and remeasuring the overall length. To be acceptable, the after-loading length must be the same as the original length within a small tolerance permitted for errors in measurement. An absence of permanent deformation demonstrates that the yield point of the fastener has not been exceeded.
Shear strengths of fasteners generally are 60% of their specified minimum tensile strength. For example, an SAE Grade 5 hex cap screw has a specified minimum tensile strength of 120,000 psi. Its shear strength is approximately 60 percent of this value, or 70,000 psi.
Tapping screws and stainless steel metric machine screws (ASTM F738) are the only industrial fasteners for which torsional strength is a specified requirement.
Hardness is a measure of a materials ability to resist abrasion and indentation. Hardness testing for fasteners is an easy, quick and non-destructive process. For steel fasteners, there is a close relationship between hardness and tensile strength. ASTMA370 and SAEJ417 identify hardness conversion tables with corresponding tensile strengths. Hardness levels of fasteners usually are expressed in terms of:
In most fastener strength grade standards, a hardness range is specified for the base material. Minimum hardness corresponds to the specified minimum tensile strength, and the maximum hardness corresponds to the maximum tensile strength, beyond which the fastener may be too brittle.
For quenched and tempered steel fasteners, hardness is measured at the surface in the threaded section and is compared against the hardness of the fasteners core. Lower surface hardness indicates that "decarburization" may be present. This is a soft surface layer caused by carbon loss during the heat-treating process. If the surface hardness is greater than the cores hardness, the surface may be carbonized. A hard skin surface could be more brittle than the underlying metal. Either of these conditions would result in poor fastener performance.
In general, carbon steel fastener strength grades can be placed into three broad groupings involving low-carbon, medium-carbon, and alloy steel. The most widely referenced strength grade for carbon steel external threaded fasteners is detailed in the SAE J429 standard. The system is comprised of bolt grades made from low-carbon steel through to alloy steels.
The common grades of the SAE system are repeated and expanded upon in separate ASTM standards, notably A307, A449, A325, and A490. The strength grade system for steel metric fasteners is provided in ISO 898/1. Table 6 indicates the SAE and ASTM strength grades for inch series threads.
The term "kip" means kilo pounds (kilo = 1,000). For example, multiply the KSI value by 1,000 to give a psi value for tensile strength.
Table 7 provides the specific strength and mechanical properties for metric fasteners.
Table 8 is used as a reference and indicates ASTM and SAE markings for steel bolts and screws in the inch series. This table identifies bolt and screw markings for low-carbon, medium-carbon, and alloy steel.
Fasteners made from low-carbon steel have insufficient carbon to permit a strengthening heat treatment. However, these steels have excellent workability and good strength properties, which can be further improved through cold working procedures. They can also be welded and hardened.
(Piping Bolt) the low carbon steel fastener ASTMA307 is a special bolt used in piping and flange work. It has properties similar to other low-carbon steel bolts except that it has the added requirement of a specified maximum tensile strength. The reason for this is to ensure that the bolt will fracture before breaking a cast iron flange on a pump or valve, if the bolt is inadvertently over tightened.
These steels have excellent workability, and as the carbon content increases so do the toughness and tensile strength, but the life of the manufacturing tools and dies decreases. To counter this problem, the raw material can be softened in a pre-treatment process such as normalizing.
On a strength cost comparison, medium-carbon steel heat-treated fasteners provide more load carrying capability per unit of cost than any other known fastener metal. In addition, the yield ultimate strength ratio is the lowest of all heat-treated steels. This assures superior ductility qualities. These bolts do absorb punishing treatment and service abuse. This is why SAE Grade 5, ASTM A449, and ASTM A325 are the most popular strength grades for externally threaded carbon steel fasteners in use.
Many different alloy steels are used in the manufacture of fasteners. SAE Grade 8, ASTM A354, and ASTM A490 are commonly used alloy steel fasteners for mechanical and structural fastening.
Property class designations, as found on the head of a metric bolt, are numerals indicating the following information:
Not all metric designations give exact tensile and yields values, as discussed earlier. Each gives reasonable approximates.
In Table 7, where only the class number is provided, those property classes are recognized in both ISO and ASTM standards. The others are classes recognized in ASTM, but not yet within ISO.
It is a mandatory regulation in SAE and ASTM standards that inch series fasteners of the medium-carbon and alloy steel strength grades and metric fasteners of all property classes be marked for grade identification. The only exceptions are slotted and recessed head screws and bolts smaller than 5 mm. Also of major importance is that these same standards require that all steel fasteners be marked to identify the manufacturer.
Identification markings are the purchaser's/ user's best guarantee of product effectiveness. By indicating the strength properties and the manufacturer, there is provision made for tractability and accountability when suspect fastener problems arise. Fasteners without markings should be viewed with a high degree of suspicion.
Table 9 identifies the ISO property class markings for steel metric bolts and screws. An equivalent SAE grade is provided to indicate comparable bolts or screws in that system.
In contrast, the strength of a nut, while equally reliant upon thread series, material, and size, is closely connected to its external geometry. Nut height and wall thickness relate directly to load carrying capability. The height of a nut establishes the length of thread engagement, and wall thickness establishes resistance to dilation. Dilation refers to outward radial spreading at the nuts bearing surface when it is axially stressed. A thick-walled nut will support more load than one with a flanged bearing surface. Nut strength grade systems must not only define mechanical properties, but they must also identify the dimensional design of the nut.
Torque is defined as the turning or twisting force exerted on an object, such as the head of the bolt in Figure 13. Tension is a pulling force. The force is applied parallel to the axis of the part or bolt, as in Figure 14.
Figure 13: Bolt Torque
Figure 14: Bolt TensionThe incline plane of a standard thread form is applied in the torque/tension relationship. By applying a turning motion (torque) to a nut, as in Figure 15, a pull (tension) is created on the bolt, thus producing a torque/tension relationship. By turning the nut clockwise with the open-wrench, the nut will clamp the two plates together, as well as provide bolt tension.
Figure 15: Bolt Preload (Tension)
Torque is the product of two measurements: force and distance. See Figure 16.
Most torque wrenches read out in foot-pound, inch-pound, or inch-ounce units.
Figure 16: Torque = Force x Distance
Elastic Limit- The amount or distance an object can be distorted (compressed, bent, stretched) and still return to the original shape when the force is removed. Refer to Figure 17 for an example of elastic limit.
Figure 17: Elastic Limit
The steel bars at 1A and 1B are at rest. Each bar is aligned with the black dot. At 2A, the steel bar is bent within elastic limit and, when pressure is removed, it springs back to its normal position at 3A. The steel bar in 2B is bent beyond its elastic limit and, when pressure is removed, the bar only springs partway back, as in 3B. '
Distortion- The change in shape or configuration of an object being altered due to application of some force is commonly referred to as distortion.
Tensile Strength- The amount of pull an object will withstand before breaking is tensile strength. This is demonstrated in Figure 18. A bar of steel has been set up in a tensile strength test machine. Heavy tension is applied, exceeding the elastic limit, causing the bar to stretch. The increased pull finally snaps the bar as tension exceeds tensile strength.
Figure 18: Tensile Strength
Residual Tension- The stress remaining in an elastic object that has been distorted and not allowed to return to its original dimension is referred to as residual tension.
Elasticity- This property relates to the ability of an object to return, after distortion, to its original shape and dimension once the distortion force has been eliminated. Figure 19 shows a steel bar in a straight original position. The bar is then deflected by pressure; if the pressure is removed, the bar returns to its original position.
Figure 19: Elasticity
Compression- A force that attempts to squeeze or compress an object is commonly referred to as compression. Figure 20 shows an object at rest; the object then is subjected to a compression load as the ram builds up pressure.
Figure 20: Compression
Cold Flow- This property refers to the tendency of an object under compression to expand outward, thus reducing its thickness in the direction of compression. Figure 21 shows a nut not tightened, therefore there is no compressive force on the gasket. As the nut is tightened, the gasket is compressed, causing it to flow outward as its thickness decreases.
Figure 21: Cold Flow
For most industrial applications, a fastener should be tightened until it has built up tension within itself that is at a value of approximately 50 to 60% of its elastic limit. Table 10 provides a reference for bolt preload as a percentage of yield strength.
Table 10Table 10 indicates that between 40 and 60% yield is where the fasteners safety factor exists. Minimum preload on the fastener is 50% yield, while the maximum preload is at 60% yield.
Table 11 is an elongation chart for common bolting materials. The indicated elongation figures are for various percentages of yield strengths of different bolts using a wrench on a bolt having a 1-inch (25.4 mm) grip length (i.e., two 1/2-inch plates = 1-inch grip). To obtain the desired elongation for a particular metal, read the elongation figure under the appropriate percentage of yield.
Next, multiply by the handle length of the wrench in inches. For example, to obtain the expected elongation for a SAE Grade 2 bolt stretched to 80% of yield with an 8-inch (203 mm) wrench handle, select the appropriate elongation figure from Table 11, which in this example is 1.5, and multiply by 8. The elongation for this Grade 2 bolt is 0.012 inch (.305 mm). Note: Divide 12 by 1000 to get .012 inch.
Always select the proper tool to accurately and safely secure a fastener. The correct tool usually is the "safest tool."
Figure 22: Proper Wrench Fit
Figure 23: Proper Wrench Technique
Figure 24 demonstrates one type of hydraulic tensioner. Hydraulic fluid from a pump is supplied to the ram in the head of the tensioner. This action creates an axial force, which is transmitted to the bolt by the puller.
Figure 24: Hydraulic TensionerThis force stretches the bolt. The bolts new length or extension is retained by simply turning the special nut socket with a small-diameter bar.
Accurate and dependable axial loading is difficult to achieve through the use of traditional methods. Torque measurement is not a reliable method of determining bolt load, as up to 80% of the energy required to tighten a threaded fastener is wasted in overcoming friction.
Hydraulic tensioners offer the following advantages:
The heating method for preloading bolts is commonly used on applications such as steam turbines and large compressors. The stud used to clamp the upper and lower casing halves together is pre-drilled through its center. The drilled hole accommodates an electrode-type rod, which when energized transmits heat throughout the length of the stud. By using an outside micrometer or special caliper, the studs elongation can be measured. When the desired elongation is reached, the nut is run down to the casing face and correct tension to the fastener is achieved as the stud cools.
There are three basic methods used to control the amount of preload developed in a bolt during tightening. These are:
Measuring the tightening torque means measuring an indirect relationship, while measuring bolt elongation and load involves direct measurement of changes in bolt properties. Within each of the three basic methods, there are special tools and techniques that further control, within varying amounts, the amount of preload "scatter."
That person makes every effort to ensure the 100 installations are as identical as possible.
Research indicates that if the resulting preloads could be measured accurately, they would be distributed through a range of up to plus or minus 10% from the average. The variations in preload are caused by:
As each of these variables is brought further under control, the "scatter range" lessens. However, reducing the variations below plus or minus 5%, even under the most favorable conditions, is extremely difficult. It is difficult to control the variables.
Unacceptable scatter could be caused by the use of fasteners from different manufacturers, poorly handled or infrequently calibrated tightening equipment, and fatigue of those performing the task.
When bolts are tightened, they become stressed in tension. Approximately 50% of the tightening torque applied is needed to overcome friction between the nut or bolt head turning against the joint surface. Another 40%, on average, is consumed overcoming friction between the mating threads.
The remaining 10% develops useful tension in the bolt or stud. The mathematical formula for describing the relationship between torque applied and tension developed is:
Special coatings and plating's commercially available can lower "K" to below 0.10.
Taking all fastener surface conditions into account, the value of "K" ranges from approximately 0.06 to 0.35.
Four measuring practices based on this concept will be discussed:
The turn-of-nut method also offers very low preload scatter. When the preload is less than yield, any scatter in the bolt elongation values magnifies the resulting preload scatter. When yield is exceeded, because the bolt has entered its nonlinear plastic zone, variations in the elongation range are possible without overly broadening the range of preload values.
The turn-of-nut method is primarily for rigid joints connected with ductile bolts that can safely be preloaded beyond their yield strength. This practice is not recommended for gasket joints, frequently disassembled joints, or joints with fasteners that must be pre-loaded to values less than yield.
Tension control systems have built-in automatic inspection capabilities to sense and signal missing fasteners as in multiple-bolt joints, cross threading, and defective threads, and to determine whether the fastener has actually yielded but is not over tightened.
Tension control systems are primarily suited to assembly line production because of their size and complexity. This equipment is expensive, and smaller handheld analyzers are being developed for practical field use. When preload accuracy and uniformity is an absolute requirement, this sophisticated equipment should be considered.
When a bolt is tightened, it stretches and changes the length of the transit time or the wavelength of the frequency required for resonance. These changes are linear functions that are proportional to bolt elongation. Ultrasonic extensometers are available with readouts translating the ultrasonic change in bolt elongation or bolt load.
The accuracy of ultrasonic equipment depends on careful extensometer calibration, which can be done by testing the load elongation of sample fasteners.
Use of ultrasonic principles is rapidly developing. Very accurate preload measurement is possible, and improvements will continue in the development of this equipment.
There are many types of torque wrenches. Figure 25 identifies a deflecting beam torque wrench. The pointer on the scale indicates when the desired torque has been reached.
Figure 25: Deflecting Beam Torque WrenchFigure 26 shows a type of torque wrench where a dial on a scale indicates when the desired torque has been reached.
Figure 26: Dial Indicating Torque Wrench
On other types of torque wrenches, the desired torque can be preset. A sign device tells the operator, by means of a beep or flashing light, when the preset torque has been applied.
If the lever length of a torque wrench is measured in inches and the force in pounds, as in Figure 27, the torque developed is measured in inch-pounds.
Figure 27: Inch-Lb Torque Wrench
If the lever length is measured in feet and the force in pounds, as in Figure 28, the torque developed is measured in foot-pounds.
Figure 28: Foot-Lb Torque Wrench
If the lever length is measured in meters and the force in Newton's, as in Figure 29, the torque developed is measured in Newton-meters.
Figure 29: Newton-Meter Torque Wrench
Torque indicators can range from inch-grams, inch-ounces, inch-pounds, to foot-pounds (Newton-meters). Inch-pounds, foot-pounds and Newton-meter torque wrenches are the most common scales.
Torque wrenches are made in different sizes or ranges, as well as in varying calibrations, such as the following:
A torque wrench will produce best results if used somewhere near the middle half of its range. For example, a 0-100 foot-pound torque wrench will provide the most accurate readings at 25 to 75 foot-pounds.
Adapters are available to help torque oddly shaped fasteners and those in hard to reach places. Usually an adapter does not increase the effective length of a torque wrench. Therefore, the reading on the wrench is the same as it would be without the adapter. Figure 30 shows various types of adapters used on torque wrenches.
Figure 30: Torque Wrench Adapters
If an adapter to a torque wrench increases the length of the wrench, it must be treated as an extension. Extensions are used to increase the torque wrench length. Examples of some typical extensions are shown in Figure 31.
When using a torque wrench with an extension, as in Figure 32, an adjustment in the reading must be made. Remember that force multiplied by length equals torque.
The extension has increased the effective length of the torque wrench. The effective length of the torque wrench is the length of the torque wrench itself added to the length of the extension.
Figure 32: Calculations with a Torque Wrench ExtensionThe formula for calculating the new torque value is as follows:
The dimension A (length of the extension), is not the actual length of the extension, but the effective length by which the torque wrench is extended, as shown in Figure 32.
If the torque value that is to be applied to the fastener is known, the desired torque wrench reading may be calculated using the following formula:
Example 1: A 3-foot long torque wrench with an extension having an effective length of 6-inch is being used to torque a metal fastener to 115 foot-pounds. What should the indicator read in order to achieve this value on the fastener? (Refer to Figure 33).
Figure 33: Torque Wrench with Extension
A=6 inches=0.5 ft
Ti=115 ft-lbs(3 ft)/(.5 ft+3 ft)
Example 2: If a 2.5-foot long torque wrench is used with an extension having an effective length of 1-foot to tighten a metal fastener, and the torque indicated by the wrench is 175 foot-pounds, what is the torque actually being applied to the fastener (Ta)? (Refer to Figure 34.)
Figure 34: Torque Wrench With Extension
Ta=175 ft-lbs(1 ft+2.5 ft)/(2.5 ft)
Example 3: It is necessary to torque a metal fastener to 255 foot-pounds using the torque wrench shown in Figure 35. The length of the wrench is 3.5 feet, and the effective length of the extension is 18 inches. To what indicated value should the torque wrench be taken in order to achieve 255 foot-pounds at the fastener?
Figure 35: Torque Wrench With ExtensionSolution:
Ti=255 ft-lbs(3.5 ft)/(1.5 ft+3.5 ft)
It may be necessary to use torque multipliers to achieve the required torque for large-diameter fasteners. When torque multipliers are used, the reading of the torque wrench is not the actual torque exerted on the fastener. The ratio between the actual torque on the fastener and the torque shown on the torque wrench dial is determined by the actual mechanical advantage of the multiplier or combination of multipliers. See Figure 36 for a multiplier example.
Figure 36: Torque Multiplier
Use a ratchet in the wrench arrangement when using more than one multiplier. Spring action in the multiplier's gear teeth can create a safety hazard when the operator releases the handle hold during tightening. A solid stop must be provided on the bottom multiplier so that the ratchet and multiplier can react against it as necessary.
A single torque multiplier can be driven by any standard ratchet or torque wrench, or it may be used in combination with several torque multipliers. Refer to Figure 36 for examples of torque multiplier combinations.
Example: If 360 foot-pounds is the required torque for a nut when one 4 to 1 torque multiplier is used, the result is (360 divided by 4) = 90 foot-pounds. This is the required torque wrench reading with the 4 to 1 multiplier.
It is recommended to use a tightening sequence on multi-bolted joints. Figure 37 demonstrates the torque sequence on an engine cylinder head. It is best to start from the center and work outward on alternating sides of the head in a circular pattern. Figure 37 also shows torquing bolts in a criss-cross sequence.
Figure 37: Torquing Sequence
Table 12 lists the torque recommendations for inch series SAE Grade 5 and Grade 8 carbon steel bolts either oiled or dry-assembled.
Table 13 through Table 18 are torque guides for Grade 2, 5, and 8 bolts, either plain or plated. These tables also indicate the clamp load that is attained with each coarse- and fine-thread fastener. Table 19 is for 4.8, 8.8, and 10.9 metric bolts.
In Table 12 through Table 18, the torque coefficient "K " = 0.20 for plain steel fasteners and 0.15 for plated fasteners.
When torque recommendations are not available from an equipment manufacturer, consult one of the torque charts identified in this section. By using the bolt's head marking, diameter and pitch or TPI, an approximate torque setting may be determined. If the fastener is threaded into aluminum, brass or thin metal, the torque figures will have to be reduced to prevent stripping. On some assembles, such as hot pipe flanges, cylinder heads, or exhaust manifolds the fasteners may have to be re-torqued after a certain period of operation.
A torque nut is a specialized nut used on a mating bolt or stud to create enormous tightening pressures on bolted flanges for piping and mechanical equipment systems.
Figure 38 shows a front and side view of a torque nut. The torque nuts main thread is only used for positioning. After the torque nut is properly positioned, actual tensioning is accomplished by tightening the jacking-bolts positioned around the main thread. Ordinary hand wrenches can be used to create enormous tightening pressures.
Figure 38: Torque Nut
A torque nut is a simple device in comparison to hydraulic tensioning devices, which can achieve levels of tension up to 10,000 tons or more.
The relative ease of mounting and removal can reduce the application time, ease maintenance procedures, and eliminate shrink fits, hydraulic nuts, and mechanical take-up devices.
A torque nut will not come loose if properly installed, even on vibrating or pulsating equipment or on constantly reversing loads. No special tools are needed to generate high torque. An ordinary torque wrench or impact wrench is all that is needed. Torque nuts can be safely used on the following applications:
Figure 39 demonstrates how a large-diameter torque nut is used to tighten a high-pressure flange bolt. In this example, a 3-inch torque nut is being tightened by torquing eight jacking bolts to 183 foot-pounds each. The tension applied to the flange is 256,000 pounds (116,364 Kg).
Figure 39: Torque Nut OperationTable 20 indicates the allowable tension load that torque nuts provide and how much torque must be applied to each jacking-bolt.
Two major complications are introduced in multi-fastener bolted joints:
The magnitude and overall severity of these two stresses on the bolt depends on the bolts location within the bolt pattern on the assembly.
Figure 40: Multi-Bolt Shear ConnectionWhen the load is increased, bolts #1 and #5 will yield and plastically deform. Bolts #2 and #4 then will receive a greater portion of the total load. Just before failure, all of the bolts will support the load equally, in theory, and they will fail together. In an extremely long joint, as in bolted structural applications, where several rows of fasteners are used, it is possible that the outer bolts will actually fail before the centrally located bolts are fully stressed. Failure then occurs in an "unbuttoning" manner, moving progressively by rows.
In Figure 41, the connecting bolts in the rigid bolted joint are subjected to bending stresses due to the "prying" of the load on the bolted joint.
Figure 41: Load Prying Stress
The joint in this application acts as a lever, with point "A" acting as the fulcrum. Bolt #1 is more critically stressed than bolt #2 because of its closeness to the applied load. Prying action bends the bolts, and the bending stresses across the bolts are non-uniform. The bolt closest to the side of the joint where the load is applied suffers the highest stress. Bolt deformation, because of bending stresses, tends to be concentrated In the threaded section.
Design guidance is available from ASTM on how to accommodate differential loading, prying action, irregular bolt patterns, and hole misalignment. Emphasized features include the importance of joint-to-bolt stiffness ratios and high bolt preloads.
Fasteners in any statically loaded bolted joint usually do not loosen. If there is no movement in the structure where the bolted joint is located, there is no reason for fastener loosening to occur.
Fasteners used to assemble a bolted joint subjected to dynamic loading are introduced to possible vibration effects, which could lead to subsequent loosening.
In tightened bolt/nut assemblies, the axial tensile stress develops frictional resistance between mating threads and between bolt and nut surfaces bearing against the joined material. Through practice, it has been found that if this frictional resistance is sufficiently reduced, even for a portion of a second, the same tensile stress in the bolt will now encourage the mating threads to "walk." This occurs because of the downhill slope of the threads helix angle. Any motion (vibration-induced), even microscopic, reduces bolt preload. This action happens progressively over a number of pulsations. The motion continues until preload is completely lost and the bolt/nut assembly is shaken into full separation from the bolted joint.
Welding the nut and/or bolt to the joint material is not recommended because the heat may:
Relaxation in a bolted joint is caused by numerous factors. Designers of bolted joints must consider relaxation as an important element in bolted assemblies.
Experience has proven that relaxation continues for several days after tightening due to further "setting in", adjusting to the first application of the service loads, and, in plated bolted joints, the possible creep or flow of plating metal, adding another 2 to 5% preload loss. A total loss averaging 10% is a reasonable amount of loss of preload after approximately three weeks. Additional relaxation may also occur over a period of time, depending upon adverse conditions.
It is recommended to tighten, loosen, and then re-tighten the bolts. This practice may help to pre-flatten rough spots on flange faces. Re-tightening the bolts after a period of time is also a good practice.
In multi-bolt patterns, it is best to tighten and re-tighten in a sequence that minimizes the occurrence of different preloads.