Maintenance Engineering -- Gears and Gearboxes

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A gear is a form of disc, or wheel, that has teeth around its periphery for the purpose of providing a positive drive by meshing the teeth with similar teeth on another gear or rack.


The spur gear might be called the basic gear since all other types have been developed from it. Its teeth are straight and parallel to the center bore line, as shown in Fgr. 1. Spur gears may run together with other spur gears or parallel shafts, with internal gears on parallel shafts, and with a rack. A rack such as the one illustrated in Fgr. 2 is in effect a straight-line gear. The smallest of a pair of gears (Fgr. 3) is often called a pinion.

The involute profile or form is the one most commonly used for gear teeth. It is a curve that's traced by a point on the end of a taut line unwinding from a circle.

The larger the circle, the straighter the curvature; for a rack, which is essentially a section of an infinitely large gear, the form is straight or flat. The generation of an involute curve is illustrated in Fgr. 4.

Fgr. 1 Example of a spur gear.

Fgr. 2 Rack or straight-line gear.

Fgr. 3 Typical spur gears.

Fgr. 4 Involute curve.

Fgr. 5 Pressure angle.

Fgr. 6 Different pressure angles on gear teeth.

Fgr. 7 Relationship of the pressure angle to the line of action.

The involute system of spur gearing is based on a rack having straight, or flat, sides. All gears made to run correctly with this rack will run with each other. The sides of each tooth incline toward the center top at an angle called the pressure angle, shown in Fgr. 5.

The 14.5-degree pressure angle was standard for many years. In recent years, however, the use of the 20-degree pressure angle has been growing, and today, 14.5-degree gearing is generally limited to replacement work. The principal reasons are that a 20-degree pressure angle results in a gear tooth with greater strength and wear resistance and permits the use of pinions with a few fewer teeth.

The effect of the pressure angle on the tooth of a rack is shown in Fgr. 6.

It is extremely important that the pressure angle be known when gears are mated, as all gears that run together must have the same pressure angle. The pressure angle of a gear is the angle between the line of action and the line tangent to the pitch circles of mating gears. Fgr. 7 illustrates the relationship of the pressure angle to the line of action and the line tangent to the pitch circles.


Pitch circles have been defined as the imaginary circles that are in contact when two standard gears are in correct mesh. The diameters of these circles are the pitch diameters of the gears. The center distance of the two gears, therefore, when correctly meshed, is equal to one half of the sum of the two pitch diameters, as shown in Fgr. 8.

Fgr. 8 Pitch diameter and center distance.

Fgr. 9 Determining center distance.

C = Center distance D1 = First pitch diameter D2 = Second pitch diameter

This relationship may also be stated in an equation and may be simplified by using letters to indicate the various values, as follows:

Example: The center distance can be found if the pitch diameters are known (Fgr. 9).


A specific type of pitch designates the size and proportion of gear teeth. In gearing terms, there are two specific types of pitch: circular pitch and diametrical pitch. Circular pitch is simply the distance from a point on one tooth to a corresponding point on the next tooth, measured along the pitch line or circle, as illustrated in Fgr. 10. Large-diameter gears are frequently made to circular pitch dimensions.


The diametrical pitch system is the most widely used, as practically all common sized gears are made to diametrical pitch dimensions. It designates the size and proportions of gear teeth by specifying the number of teeth in the gear for each inch of the gear's pitch diameter. For each inch of pitch diameter, there are pi (p) inches, or 3.1416 in., of pitch-circle circumference. The diametric pitch number also designates the number of teeth for each 3.1416 in. of pitch-circle circumference. Stated in another way, the diametrical pitch number specifies the number of teeth in 3.1416 in. along the pitch line of a gear.

Fgr. 11 Pitch diameter and diametrical pitch.

For simplicity of illustration, a whole-number pitch-diameter gear (4 in.), is shown in Fgr. 11.

Fgr. 11 illustrates that the diametrical pitch number specifying the number of teeth per inch of pitch diameter must also specify the number of teeth per 3.1416 in. of pitch-line distance. This may be more easily visualized and specifically dimensioned when applied to the rack in Fgr. 12.

Because the pitch line of a rack is a straight line, a measurement can be easily made along it. In Fgr. 12, it's clearly shown that there are 10 teeth in 3.1416 in.; therefore the rack illustrated is a 10 diametrical pitch rack.

A similar measurement is illustrated in Fgr. 13, along the pitch line of a gear. The diametrical pitch being the number of teeth in 3.1416 in. of pitch line, the gear in this illustration is also a 10 diametrical pitch gear.

In many cases, particularly in machine repair work, it may be desirable for the mechanic to determine the diametrical pitch of a gear. This may be done very easily without the use of precision measuring tools, templates, or gauges. Measurements need not be exact because diametrical pitch numbers are usually whole numbers. Therefore, if an approximate calculation results in a value close to a whole number, that whole number is the diametrical pitch number of the gear.

The following three methods may be used to determine the approximate diametrical pitch of a gear. A common steel rule, preferably flexible, is adequate to make the required measurements.

Fgr. 12 Number of teeth in 3.1416 in.

Fgr. 13 Number of teeth in 3.1416 in. on the pitch circle.


Count the number of teeth in the gear, add 2 to this number, and divide by the outside diameter of the gear. Scale measurement of the gear to the closest fractional size is adequate accuracy.

Fgr. 14 illustrates a gear with 56 teeth and an outside measurement of 5/13 16 in. Adding 2 to 56 gives 58; dividing 58 by 5 13/16 gives an answer of 9 31/32.

Since this is approximately 10, it can be safely stated that the gear is a 10 decimal pitch gear.

METHOD 2 Count the number of teeth in the gear and divide this number by the measured pitch diameter. The pitch diameter of the gear is measured from the root or bottom of a tooth space to the top of a tooth on the opposite side of the gear.

Fgr. 15 illustrates a gear with 56 teeth. The pitch diameter measured from the bottom of the tooth space to the top of the opposite tooth is 5-5/8 in. Dividing 56 by 5-5/8 gives an answer of 9-15/16 in. or approximately 10. This method also indicates that the gear is a 10 decimal pitch gear.

Fgr. 14 Use of Method 1 to approximate the diametrical pitch. In this method the outside diameter of the gear is measured.

Fgr. 15 Use of Method 2 to approximate the diametrical pitch. This method uses the pitch diameter of the gear.

PITCH CALCULATIONS Diametrical pitch, usually a whole number, denotes the ratio of the number of teeth to a gear's pitch diameter. Stated another way, it specifies the number of teeth in a gear for each inch of pitch diameter. The relationship of pitch diameter, diametrical pitch, and number of teeth can be stated mathematically as follows.

P = N/D so N =D x P where, D == Pitch diameter P = Diametrical pitch N = Number of teeth If any two values are known, the third may be found by substituting the known values in the appropriate equation.

Example 1: What is the diametrical pitch of a 40-tooth gear with a 5-in. pitch diameter? P = N/D so N =D x P and P= 8 diametrical pitch Example 2: What is the pitch diameter of a 12 diametrical pitch gear with 36 teeth? D = 36/12 so, D = 3-in: pitch diameter Example 3: How many teeth are there in a 16 diametrical pitch gear with a pitch diameter of 3 3/4 in.? N = D x P or N = 3 _ 3=4 _ 16 or N = 60 teeth Circular pitch is the distance from a point on a gear tooth to the corresponding point on the next gear tooth measured along the pitch line. Its value is equal to the circumference of the pitch circle divided by the number of teeth in the gear. The relationship of the circular pitch to the pitch-circle circumference, number of teeth, and the pitch diameter may also be stated mathematically as follows:

Circumference of pitch circle P = pi x D P = pi x D/N where, D = Pitch diameter N = Number of teeth P = Circular pitch pi = 3:1416

If any two values are known, the third may be found by substituting the known values in the appropriate equation.

Example 1: What is the circular pitch of a gear with 48 teeth and a pitch diameter of 6 in.? P = 0.3927 inches

Example 2: What is the pitch diameter of a 0.500-in. circular-pitch gear with 128 teeth? D=20.371 inches Fgr. 16 Names of gear parts.

Fgr. 17 Names of rack parts.

The list that follows contains just a few names of the various parts given to gears.

These parts are shown in Fgrs. 16 and 17.

_ Addendum: Distance the tooth projects above, or outside, the pitch line or circle.

_ Dedendum: Depth of a tooth space below, or inside, the pitch line or circle.

_ Clearance: Amount by which the dedendum of a gear tooth exceeds the addendum of a matching gear tooth.

_ Whole Depth: The total height of a tooth or the total depth of a tooth space.

_ Working Depth: The depth of tooth engagement of two matching gears. It is the sum of their addendums.

_ Tooth Thickness: The distance along the pitch line or circle from one side of a gear tooth to the other.


The full-depth involute system is the gear system in most common use. The formulas (with symbols) shown below are used for calculating tooth proportions of full-depth involute gears. Diametrical pitch is given the symbol P as before.


Addendum, a = 1/P

Whole Depth, W_d Dedendum


Fgr. 18 Backlash.

Backlash in gears is the play between teeth that prevents binding. In terms of tooth dimensions, it's the amount by which the width of tooth spaces exceeds the thickness of the mating gear teeth. Backlash may also be described as the distance, measured along the pitch line, that a gear will move when engaged with another gear that's fixed or immovable, as illustrated in Fgr. 18.

Normally there must be some backlash present in gear drives to provide running clearance. This is necessary because binding of mating gears can result in heat generation, noise, abnormal wear, possible overload, and /or failure of the drive.

A small amount of backlash is also desirable because of the dimensional variations involved in practical manufacturing tolerances.

Backlash is built into standard gears during manufacture by cutting the gear teeth thinner than normal by an amount equal to one half the required figure. When two gears made in this manner are run together, at standard center distance, their allowances combine, provided the full amount of backlash is required.

On non-reversing drives or drives with continuous load in one direction, the increase in backlash that results from tooth wear does not adversely affect operation. However, on reversing drive and drives where timing is critical, excessive backlash usually can't be tolerated.


Many styles and designs of gears have been developed from the spur gear. While they are all commonly used in industry, many are complex in design and manufacture. Only a general description and explanation of principles will be given, as the field of specialized gearing is beyond the scope of this book.

Commonly used styles will be discussed sufficiently to provide the millwright or mechanic with the basic information necessary to perform installation and maintenance work.


Two major differences between bevel gears and spur gears are their shape and the relation of the shafts on which they are mounted. The shape of a spur gear is essentially a cylinder, while the shape of a bevel gear is a cone. Spur gears are used to transmit motion between parallel shafts, while bevel gears transmit motion between angular or intersecting shafts. The diagram in Fgr. 19 illustrates the bevel gear's basic cone shape. Fgr. 20 shows a typical pair of bevel gears.

Special bevel gears can be manufactured to operate at any desired shaft angle, as shown in Fgr. 21.Miter gears are bevel gears with the same number of teeth in both gears operating on shafts at right angles or at 90 degrees, as shown in Fgr. 22.

A typical pair of straight miter gears is shown in Fgr. 23. Another style of miter gears having spiral rather than straight teeth is shown in Fgr. 24. The spiral-tooth style will be discussed later.

The diametrical pitch number as is done with spur gears establishes the tooth size of bevel gears. Because the tooth size varies along its length, it must be measured at a given point. This point is the outside part of the gear where the tooth is the largest. Because each gear in a set of bevel gears must have the same angles and tooth lengths, as well as the same diametrical pitch, they are manufactured and distributed only in mating pairs. Bevel gears, like spur gears, are manufactured in both the 14.5-degree and 20-degree pressure-angle designs.

Fgr. 19 Basic shape of bevel gears.

Fgr. 20 Typical set of bevel gears Fgr. 21 Shaft angle, which can be at any degree.

Fgr. 22 Miter gears, which are shown at 90 degrees.


Helical gears are designed for parallel-shaft operation like the pair in Fgr. 25.

They are similar to spur gears except that the teeth are cut at an angle to the centerline. The principal advantage of this design is the quiet, smooth action that results from the sliding contact of the meshing teeth. A disadvantage, however, is the higher friction and wear that accompanies this sliding action. The angle at which the gear teeth are cut is called the helix angle and is illustrated in Fgr. 26.

It is very important to note that the helix angle may be on either side of the gear's centerline. Or if compared with the helix angle of a thread, it may be either a ''right-hand'' or a ''left-hand'' helix. The hand of the helix is the same regardless of how viewed. Fgr. 27 illustrates a helical gear as viewed from opposite sides; changing the position of the gear can't change the hand of the tooth's helix angle. A pair of helical gears, as illustrated in Fgr. 25, must have the same pitch and helix angle but must be of opposite hands (one right hand and one left hand).

Helical gears may also be used to connect nonparallel shafts. When used for this purpose, they are often called ''spiral'' gears or crossed-axis helical gears. This style of helical gearing is shown in Fgr. 28.

Fgr. 23 Typical set of miter gears.

Fgr. 24 Miter gears with spiral teeth.

Fgr. 25 Typical set of helical gears.

Fgr. 26 The angle at which teeth are cut.

Fgr. 27 Helix angle of teeth: the same no matter from which side the gear is viewed.

Fgr. 28 Typical set of spiral gears.

Fgr. 29 Typical set of worm gears.

Fgr. 30 Herringbone gear.


The worm and worm gear, illustrated in Fgr. 29, are used to transmit motion and power when a high-ratio speed reduction is required. They provide a steady quiet transmission of power between shafts at right angles. The worm is always the driver and the worm gear the driven member. Like helical gears, worms and worm gears have ''hand.'' The hand is determined by the direction of the angle of the teeth. Thus, for a worm and worm gear to mesh correctly, they must be the same hand.

The most commonly used worms have either one, two, three, or four separate threads and are called single, double, triple, and quadruple thread worms. The number of threads in a worm is determined by counting the number of starts or entrances at the end of the worm. The thread of the worm is an important feature in worm design, as it's a major factor in worm ratios. The ratio of a mating worm and worm gear is found by dividing the number of teeth in the worm gear by the number of threads in the worm.


To overcome the disadvantage of the high end thrust present in helical gears, the herringbone gear, illustrated in Fgr. 30, was developed. It consists simply of two sets of gear teeth, one right hand and one left hand, on the same gear. The gear teeth of both hands cause the thrust of one set to cancel out the thrust of the other. Thus the advantage of helical gears is obtained, and quiet, smooth operation at higher speeds is possible. Obviously they can only be used for transmitting power between parallel shafts.


Many machine-trains utilize gear drive assemblies to connect the driver to the primary machine. Gears and gearboxes typically have several vibration spectra associated with normal operation. Characterization of a gearbox's vibration signature box is difficult to acquire but is an invaluable tool for diagnosing machine-train problems. The difficulty is that (1) it's often difficult to mount the transducer close to the individual gears, and (2) the number of vibration sources in a multi-gear drive results in a complex assortment of gear mesh, modulation, and running frequencies. Severe drive-train vibrations (gearbox) are usually due to resonance between a system's natural frequency and the speed of some shaft. The resonant excitation arises from, and is proportional to, gear inaccuracies that cause small periodic fluctuations in pitch-line velocity.

Complex machines usually have many resonance zones within their operating speed range because each shaft can excite a system resonance. At resonance these cyclic excitations may cause large vibration amplitudes and stresses.

Basically, forcing torque arising from gear inaccuracies is small. However, under resonant conditions torsional amplitude growth is restrained only by damping in that mode of vibration. In typical gearboxes this damping is often small and permits the gear-excited torque to generate large vibration amplitudes under resonant conditions.

One other important fact about gear sets is that all gear sets have a designed preload and create an induced load (thrust) in normal operation. The direction, radial or axial, of the thrust load of typical gear sets will provide some insight into the normal preload and induced loads associated with each type of gear.

To implement a predictive maintenance program, a great deal of time should be spent understanding the dynamics of gear/gearbox operation and the frequencies typically associated with the gearbox. As a minimum, the following should be identified.

Gears generate a unique dynamic profile that can be used to evaluate gear condition. In addition, this profile can be used as a tool to evaluate the operating dynamics of the gearbox and its related process system.

Gear Damage

All gear sets create a frequency component, called gear mesh. The fundamental gear mesh frequency is equal to the number of gear teeth times the running speed of the shaft. In addition, all gear sets will create a series of side bands or modulations that will be visible on both sides of the primary gear mesh frequency. In a normal gear set, each of the side bands will be spaced at exactly the 1X or running speed of the shaft and the profile of the entire gear mesh will be symmetrical.

Fgr. 31 Normal profile is symmetrical.

Normal Profile

In a normal gear set, each of the side bands will be spaced at exactly the 1X running speed of the input shaft, and the entire gear mesh will be symmetrical. In addition, the side bands will always occur in pairs, one below and one above the gear mesh frequency. The amplitude of each of these pairs will be identical. For example, the side band pair indicated as _1 and รพ1 in Fgr. 31 will be spaced at exactly input speed and have the same amplitude.

If the gear mesh profile were split by drawing a vertical line through the actual mesh (i.e., number of teeth times the input shaft speed), the two halves would be exactly identical. Any deviation from a symmetrical gear mesh profile is indicative of a gear problem. However, care must be exercised to ensure that the problem is internal to the gears and induced by outside influences. External misalignment, abnormal induced loads, and a variety of other outside influences will destroy the symmetry of the gear mesh profile. For example, the single reduction gearbox used to transmit power to the mold oscillator system on a continuous caster drives two eccentrics. The eccentric rotation of these two cams is transmitted directly into the gearbox and will create the appearance of eccentric meshing of the gears. The spacing and amplitude of the gear mesh profile will be destroyed by this abnormal induced load.

Excessive Wear

Fgr. 32 illustrates a typical gear profile with worn gears. Note that the spacing between the side bands becomes erratic and they are no longer spaced at the input shaft speed. The side bands will tend to vary between the input and output speeds but will not be evenly spaced.

Fgr. 32 Wear or excessive clearance changes side band spacing.

In addition to gear tooth wear, center-to-center distance between shafts will create an erratic spacing and amplitude. If the shafts are too close together, the spacing will tend to be at input shaft speed, but the amplitude will drop drastically. Because the gears are deeply meshed (i.e., below the normal pitch line), the teeth will maintain contact through the entire mesh. This loss of clearance will result in lower amplitudes but will exaggerate any tooth profile defect that may be present.

If the shafts are too far apart, the teeth will mesh above the pitch line. This type of meshing will increase the clearance between teeth and amplify the energy of the actual gear mesh frequency and all of its side bands. In addition, the load bearing characteristics of the gear teeth will be greatly reduced. Since the pres sure is focused on the tip of each tooth, there is less cross-section and strength in the teeth. The potential for tooth failure is increased in direct proportion the amount of excess clearance between shafts.

Cracked or Broken Tooth

Fgr. 33 A broken tooth will produce an asymmetrical side band profile.

Fgr. 33 illustrates the profile of a gear set with a broken tooth. As the gear rotates, the space left by the chipped or broken tooth will increase the mechanical clearance between the pinion and bull gear. The result will be a low amplitude side band that will occur to the left of the actual gear mesh frequency. When the next, undamaged teeth mesh, the added clearance will result in a higher energy impact.

The resultant side band, to the right of the mesh frequency, will have much higher amplitude. The paired side bands will have non-symmetrical amplitude that represents this disproportional clearance and impact energy.

If the gear set develops problems, the amplitude of the gear mesh frequency will increase and the symmetry of the side bands will change. The pattern illustrated In Fgr. 34 is typical of a defective gear set. Note the asymmetrical relation ship of the side bands.


You should have a clear understanding of the types of gears generally utilized in today's machinery, how they interact, and the forces they generate on a rotating shaft. There are two basic classifications of gear drives: (1) shaft centers parallel, and (2) shaft centers not parallel. Within these two classifications are several typical gear types.

Shaft Centers Parallel

There are four basic gear types that are typically used in this classification. All are mounted on parallel shafts and , unless an idler gear is also used, will have opposite rotation between the drive and driven gear (if the drive gear has a clockwise rotation, then the driven gear will have a counter-clockwise rotation).

The gear sets commonly used in machinery include the following.

Fgr. 34 Typical defective gear mesh signature.

Spur Gears

The shafts are in the same plane and parallel. The teeth are cut straight and parallel to the axis of the shaft rotation. No more than two sets of teeth are in mesh at one time, so the load is transferred from one tooth to the next tooth rapidly. Usually spur gears are used for moderate to low speed applications.

Rotation of spur gear sets is opposite unless one or more idler gears are included in the gearbox. Typically, spur gear sets will generate a radial load (preload) opposite the mesh on their shaft support bearings and little or no axial load.

Backlash is an important factor in proper spur gear installation. A certain amount of backlash must be built into the gear drive allowing for tolerances in concentricity and tooth form. Insufficient backlash will cause early failure be cause of overloading.

As indicated in Fgr. 11, spur gears by design have a preload opposite the mesh and generate an induced load, or tangential force (TF) in the direction of rotation. This force can be calculated as:

TF = (126, 000 + HP) / Dp + RPM

In addition, a spur gear will generate a separating force, S_TF , that can be calculated as:

S_TF = TF _ tan PHI where TF = Tangential Force HP = Input horsepower to pinion or gear Dp = Pitch diameter of pinion or gear RPM = Speed of pinion or gear f = Pinion or gear tooth pressure angle

Helical Gears

The shafts are in the same plane and parallel but the teeth are cut at an angle to the centerline of the shafts. Helical teeth have an increased length of contact, run quieter and have a greater strength and capacity than spur gears. Normally the angle created by a line through the center of the tooth and a line parallel to the shaft axis is 45 degrees. However, other angles may be found in machinery. Helical gears also have a preload by design; the critical force to be considered, however, is the thrust load (axial) generated in normal operation; see Fgr. 12.

S_TF = (TF + tan PHI) / cos lambda

where TF = Tangential Force STF = Separating Force TTF = Thrust Force HP = Input horsepower to pinion or gear Dp = Pitch diameter of pinion or gear RPM = Speed of pinion or gear Phi = Pinion or gear tooth pressure angle lambda = Pinion or gear helix angle

Herringbone Gears

These are commonly called double helical because they have teeth cut with right and left helix angles. They are used for heavy loads at medium to high speeds.

They don't have the inherent thrust forces that are present in helical gear sets.

Herringbone gears, by design, cancel the axial loads associated with a single helical gear. The typical loads associated with herringbone gear sets are the radial side-load created by gear mesh pressure and a tangential force in the direction of rotation.

Internal Gears

Internal gears can be run only with an external gear of the same type, pitch, and pressure angle. The preload and induced load will depend on the type of gears used. Refer to spur or helical for axial and radial forces.


One of the primary causes of gear failure is the fact that, with few exceptions, gear sets are designed for operation in one direction only. Failure is often caused by inappropriate bi-directional operation of the gearbox or backward installation of the gear set. Unless specifically manufactured for bi-directional operation, the ''non-power'' side of the gear's teeth is not finished. Therefore, this side is rougher and does not provide the same tolerance as the finished ''power'' side.

Note that it has become standard practice in some plants to reverse the pinion or bull gear in an effort to extend the gear set's useful life. While this practice permits longer operation times, the torsional power generated by a reversed gear set is not as uniform and consistent as when the gears are properly installed.

Table 1 Common Failure Modes of Gearboxes and Gear Sets


Bent Shaft Broken or Loose Bolts or Setscrews Damaged Motor Elliptical Gears Exceeds Motor's Brake Horsepower Rating Excessive or Too Little Backlash Excessive Torsional Loading Foreign Object In Gearbox Gear Set Not Suitable for Application Gears Mounted Backward on Shafts Incorrect Center-to-Center Distance Between Shafts Incorrect Direction of Rotation Lack of or Improper Lubrication Misalignment of Gears or Gearbox Overload Process Induced Misalignment Unstable Foundation Water or Chemicals in Gearbox Worn Bearings Worn Coupling



Gear Failures Variations in Torsional Power Insufficient Power Output Overheated Bearings Short Bearing Life Overload on Driver High Vibration High Noise Levels Motor Trips

Gear overload is another leading cause of failure. In some instances, the over load is constant, which is an indication that the gearbox is not suitable for the application. In other cases, the overload is intermittent and only occurs when the speed changes or when specific production demands cause a momentary spike in the torsional load requirement of the gearbox.

Misalignment, both real and induced, is also a primary root cause of gear failure.

The only way to ensure that gears are properly aligned is to hard blue the gears immediately following installation. After the gears have run for a short time, their wear pattern should be visually inspected. If the pattern does not conform to vendor's specifications, alignment should be adjusted.

Poor maintenance practices are the primary source of real misalignment problems. Proper alignment of gear sets, especially large ones, is not an easy task.

Gearbox manufacturers don't provide an easy, positive means to ensure that shafts are parallel and that the proper center-to-center distance is maintained.

Induced misalignment is also a common problem with gear drives. Most gear boxes are used to drive other system components, such as bridle or process rolls.

If misalignment is present in the driven members (either real or process induced), it also will directly affect the gears. The change in load zone caused by the misaligned driven component will induce misalignment in the gear set. The effect is identical to real misalignment within the gearbox or between the gearbox and mated (i.e., driver and driven) components.

Visual inspection of gears provides a positive means to isolate the potential root cause of gear damage or failures. The wear pattern or deformation of gear teeth provides clues as to the most likely forcing function or cause. The following sections discuss the clues that can be obtained from visual inspection.


Fgr. 35 illustrates a gear that has a normal wear pattern. Note that the entire surface of each tooth is uniformly smooth above and below the pitch line.


Fgrs. 36 through 39 illustrate common abnormal wear patterns found in gear sets. Each of these wear patterns suggests one or more potential failure modes for the gearbox.

Fgr. 35 Normal wear pattern.

Fgr. 36 Wear pattern caused by abrasives in lubricating oil.


Abrasion creates unique wear patterns on the teeth. The pattern varies, depending on the type of abrasion and its specific forcing function. Fgr. 36 illustrates severe abrasive wear caused by particulates in the lubricating oil. Note the score marks that run from the root to the tip of the gear teeth.

Chemical Attack or Corrosion

Water and other foreign substances in the lubricating oil supply also cause gear degradation and premature failure. Fgr. 37 illustrates a typical wear pattern on gears caused by this failure mode.

Fgr. 37 Pattern caused by corrosive attack on gear teeth.

Fgr. 38 Pitting caused by gear overloading.


The wear patterns generated by excessive gear loading vary, but all share similar components. Fgr. 38 illustrates pitting caused by excessive torsional loading. The pits are created by the implosion of lubricating oil. Other wear patterns, such as spalling and burning, can also help to identify specific forcing functions or root causes of gear failure.

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