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AMAZON multi-meters discounts AMAZON oscilloscope discounts Electric motors are devices that convert electrical energy into magnetic energy and finally into mechanical energy. The mechanical energy is generally transmitted from the rotor through a shaft that must be free to rotate in some type of bearing system (see ill. 43). The choice of the bearing system is key to the motor's performance and life. ill. 43 Electric motor components. The system generally consists of a shaft, a bearing, and a lubricant arranged in a fashion that maintains a film of lubricant between the shaft and the bearing surface. The components and system are typically chosen to meet the requirement of the specific application. This section first discusses the components, then systems, and finally the application. 11.1 Bearings Bearing systems are used to support the rotor and shaft assembly so that it remains in a certain constant position relative to the stator and so as to reduce the friction between the shaft and the end frames. The most common bearings used in motors are ball bearings and sleeve bearings. Ball bearings are typically constructed as shown in ill. 44. They consist of an inner race and outer race, balls, and a ball carrier. The races and balls are typically highly polished hardened steel. The ball carrier may be steel or plastic. The inner race supports the shaft and rotates with it. The outer race is held stationary in the end frame. The balls provide a low-friction method of allowing the inner race to roll with respect to the outer race as the shaft turns. The carrier maintains proper spacing of the balls to evenly distribute the load. Ball bearings are lubricated by injecting grease around the balls between the races. The grease may be contained by means of a shield or seal that fits between the races. Sealed bearings have a higher coefficient of friction, require more torque from the motor, and are more costly than shielded bearings. Therefore, they are used in applications such as pumps where it's necessary to keep moisture or corrosive agents out of the bearings. Ball bearings need to be preloaded to keep the balls from moving freely in the axial direction. The amount of preload is listed in the bearing manufacturer's data for each type of bearing. Preloading is generally accomplished by means of a coil spring or wavy washer. ill. 44 Ball bearing construction. Ball bearings are generally purchased by grade number. The higher grade numbers have tighter part tolerances and lower radial play, and are more costly. High-grade bearings are used in applications where radial rotor or shaft movement must be minimized. The ABEC grades and tolerances are shown in Table TABLE 3 Ball-Bearing; Grades; Maximum radial runout; Mean diameter tolerance. 11.2 Bearing Selection* There are several important considerations which must be evaluated simultaneously when choosing the proper bearing for a particular device. The following subsections briefly discuss some of the more important ones. Miniature and instrument ball bearings are normally made of either stainless steel or chrome alloy steel. The load ratings given are for chrome steel unless other-wise noted. Load ratings are affected by bearing material. Life calculations are affected by bearing material as well as lubrication selection. Type of Cage. Two types of pressed-steel ball cages are available for most bearings, H (crown type) and R (two-piece ribbon type). These two cage types are inter-changeable in most common applications. Cages made of molded and machined plastics are also available for some sizes (see ill. 45). ill. 45 Molded and machined plastic ball bearing cages. Shields and Seals. Shields are available for most sizes. These closures help to reduce the entrance of particulate contaminants into the bearing and reduce the amount of lubricant leakage. Radial clearance between the shield bore and the inner ring OD is approximately 0.002 to 0.005 in. The effect of shields on bearing torque or noise is insignificant. Contacting seals made of synthetic rubber (type DD), as shown in ill. 46, are available for most sizes. These seals provide the best protection from the entrance of contaminants or exit of lubricant, but they significantly increase operating torque. DD seals will withstand a slight amount of positive pressure differential. Noncontacting seals made of synthetic rubber (type SS) or reinforced polytetrafluoroethylene [PTFE (Teflon; type LL)], as shown in ill. 47, are also available for most chassis sizes. This type of seal offers better sealing than a metal shield, while keeping operating torque at the lowest possible levels. LL seals will contact the inner ring in some cases, but the nature of the seal material serves to keep torque at a minimum. Radial Play. Radial play is the free internal radial looseness between the balls and races. Radial play within a ball bearing is necessary to accommodate thermal expansion and the effects of interference fits and to control axial play. Table 4 suggests radial play ranges for some typical uses. Starting and Running Torque. The operating torque of a bearing can be described as starting and running torque. Starting torque is the torque required to begin rotation from a bearing at rest. Running torque is the torque required to rotate one ring at a known speed while keeping the other ring stationary. The main contributors to bearing torque are seal and lubrication type. ill. 46 Contacting seals (type DD, synthetic rubber) ill. 47 Noncontacting seals (type LL, reinforced Teflon). TABLE 4 Radial Play Ranges Typical application | Suggested radial play, in Small high-speed precision electric motors 0.0005-0.0008 Tape guides and belt guides, low speed 0.0002-0.0005 Tape guides and belt guides, high speed 0.0005-0.0008 Precision gear trains, low-speed electric motors, synchros, and servos 0.0002-0.0005 Static Cor and Dynamic Cr Loads. In evaluating three static load conditions, any forces exerted during assembly and test must be considered along with vibration and impact loads sustained during handling, testing, shipment, and assembly. Dynamic loading includes built-in preload, weight of supported members, and the effect of any accelerations due to vibration or motion changes. The static and dynamic radial load ratings are shown for each chassis size in the tables that follow. Speed of Operation. Although a very large bearing might be the best choice for long life due to its load-carrying capacity, it might very well fail early because of damage due to high centripetal forces or rubbing speeds generated by the rotational velocity. To determine whether a particular bearing will operate satisfactorily at the speed Nmax required in a particular device, multiply the value given for that bearing by Eq. (3.3) by the proper factor taken from Table 5. This table takes into account lubricant, retainer type, and ring rotation. Manufacturer's speed rating ≤ Nmax / fn (3.3) 11.3 Optimum Lubricant Selection of the lubricant is extremely important. Many lubricants are available for varying conditions and requirements. Unless torque is a problem, the selection of a grease is much preferred in pre-lubricating bearings since it's less susceptible to migration and leakage. Grease can multiply the inherent bearing torque by a factor of 1.2 to 5.0, depending on the type and quantity of grease in the bearing. Table 5 gives a partial listing of the most common greases. TABLE 5 fn Versus Cage Type, Lubricant Type, and Ring Rotation Ring rotation Metal cage, 2-piece or crown type Acetal cage [Crown type, Full-section type] Lubricant Inner Outer Inner Outer Inner Outer Petroleum oil 1.0 0.8 2.0 1.2 4.0 2.4 Synthetic oil 1.0 0.8 2.0 1.2 4.0 2.4 Silicone oil 0.8 0.7 0.8 0.7 0.8 0.7 Non-channeling grease 1.0 0.6 1.6 1.0 1.6 1.0 Channeling grease 1.0 0.8 2.0 1.2 2.4 1.3 Silicone grease 0.8 0.7 0.8 0.7 0.8 0.7 11.4 Ball-Bearing Components To assist in selecting the bearing with the proper components (ill. 48) for a particular design or use, an exploded view of a standard ball bearing with component callouts is shown in ill. 49. The part numbering system is shown in Table 6. To further illustrate the relative positioning of these components in the ball-bearing assembly, a cross section is shown in ill. 50. Basic Dimension Data. The dimensions and their associated symbols are shown in ill. 51 and defined here. These dimensions establish bearing size and other bearing parameters so that designers may choose the ball bearing most suited to their requirements. The symbols shown in ill. 51 and used throughout this section are defined as follows: d = inside diameter or bore D = outside diameter (OD) B = inner ring width C = outer ring width TABLE 6 Part Numbering System Group Factor Designation Description 1 Material DD Stainless-steel material which falls within the 400 series martensitic stainless-steel grouping. No code = chrome alloy steel (52100 or equivalent). 3 Basic size 418 Inch series first-one or two digits indicate OD 5532 in 16ths of an inch. The following two or three digits indicate the bore size in fractions of an inch, the first digit being the numerator and the second or the second and third digits being the denominator. Metric series first-two digits indicate OD in mm. Second two digits indicate ID in mm. Special size series ZB = integral shaft AS pulley-type assemblies, shaft assemblies, mechanical parts, tape guides, special pivot type, special bearings X following basic size indicates special ball complement assigned in numerical sequence, i.e., X1, X2, etc. 4 Features ZZEE Enclosures Z single metallic shield, removable ZZ double metallic shield, removable D single rubber seal, contact DD double rubber seal, contact L single glass-reinforced PTFE seal, noncontact LL double glass-reinforced PTFE seal, noncontact LZ glass-reinforced PTFE seal and shield with seal on flange side ZL shield and glass-reinforced PTFE seal with shield on flange side DZ rubber seal and shield SSD21 labyrinth seal, noncontact H single metallic shield, non-removable HH double metallic shield, nonremovable S single rubber seal, noncontact SS double rubber seal, noncontact Extended inner ring EE Both sides TABLE 6 Part Numbering System (Continue d) Group Factor Designation Description Group Factor Designation Description 5 Anderon Anderon meter test meter test MT motor quality and special GT extremely quiet, HDD spindle motor designs only No code noncritical application Special design SD special design bearing 6 Cage H H crown R ribbon J acetal crown type MN glass-fiber-reinforced molded nylon M7 molded nylon 7 ABEC A7 A1 ABEC 1 tolerance A7 A3 ABEC 3 A5 ABEC 5 A7 ABEC 7 Note: A1 miniature and instrument bearings of both the metric and inch configurations meet the tolerances of ABMA Standard 20 for ABEC 1 metric-series bearings. 8 Radial P25 P followed by two, three, or four numbers indicates the radial play limits in ten-thousandths of an inch. Example: P25 indicates radial play of 0.0002 to 0.0005 in. 9 Lubricant LY75 Lubricant letter codes are followed by a number LO1 to indicate specific type. LO oil LG = greases LY = other oils and greases LD = dry, no lubrication (DD material only) 10 Lube L X = 5-10% quantity L = 10-15% T = 15-20% No code = 25-35% H = 40-50% J = 50-60% F = 100% A = void volume Df = flange outside diameter Bf = flange width or thickness Li = inner ring reference diameter Lo = outer ring reference diameter r = maximum shaft of housing fillet radius that bearing corners will clear Z = number of balls DW = nominal diameter of balls Nmax = maximum speed, rpm fn = cage and lubricant factor (See Table 5.) ill. 48 Ball bearing components. ill. 49 Cross section of ball bearing. ill. 50 Bearing components. 3.11.5 Internal Bearing Geometry When designing ball bearings for optimum performance, internal bearing geometry is a critical factor. For any given bearing load, internal stresses can be either high or low, depending on the geometric relationship between the balls and raceways inside the ball-bearing structure. When a ball bearing is running under a load, force is transmitted from one bearing ring to the other through the ball set. Since the contact area between each ball and the rings is relatively small, even moderate loads can produce stresses of tens or even hundreds of thousands of pounds per square inch. Because internal stress levels have such an important effect on bearing life and performance, internal geometry must be carefully chosen for each application so bearing loads can be distributed properly. Raceway, Track Diameter, and Track Radius. The raceway in a ball bearing is the circular groove formed in the outside surface of the inner ring and in the inside surface of the outer ring. When the rings are aligned, these grooves form a circular track that contains the ball set. The track diameter and track radius are two dimensions that define the con-figuration of each raceway. Track diameter is the measurement of the diameter of the imaginary circle running around the deepest portion of the raceway, whether it be an inner or outer ring. This measurement is made along a line perpendicular to, and intersecting, the axis of rotation. Track radius describes the cross section of the arc formed by the raceway groove. It is measured when viewed in a direction perpendicular to the axis of the ring. In the context of ball-bearing terminology, track radius has no mathematical relationship to track diameter. The distinction between the two is shown in ill. 52. Radial and Axial Play. Most ball bearings are assembled in such a way that a slight amount of looseness exists between balls and raceways. This looseness is referred to as radial play and axial play. Specifically, radial play is the maximum distance that one bearing ring can be displaced with respect to the other, in a direction perpendicular to the bearing axis, when the bearing is in an unmounted state. Axial play, or end play, is the maximum relative displacement between the two rings of an unmounted ball bearing in the direction parallel to the bearing axis. ill 53 illustrates these concepts. Since radial play and axial play are both consequences of the same degree of looseness between the components in a ball bearing, they bear a mutual dependence. While this is true, both values are usually quite different in magnitude. In most ball-bearing applications, radial play is functionally more critical than axial play. If axial play is determined to be an essential requirement, control can be obtained through manipulation of the radial-play specification. ill. 51 Bearing dimensions and symbols. ill. 52 Distinction between track radius and track diameter. ill. 53 Distinction between radial and axial play. TABLE 7 Ball-Bearing Contact Angles The initial contact angle of the bearing is directly related to radial play-the higher the radial play, the higher the contact angle. Table 7 shows nominal contact angles, and Table 8 shows typical radial-play ranges. The contact angles in Table 7 are given for the mean radial play of the ranges shown- i.e., for P25 (0.0002 to 0.0005 in), the contact angle is given for 0.00035 in. Contact angle is affected by raceway curvature. For support of pure radial loads, a low level of radial play is desirable. Where thrust loading is predominant, higher radial-play levels are recommended. Radial play is affected by any interference fit between the shaft and bearing ID or between the housing and bearing OD. TABLE 8 Typical Radial-Play Ranges Description Radial-play range NMB code Tight 0.0001-0.0003 in P13 Normal 0.0002-0.0005 in P25 Loose 0.0005-0.0008 in P58 Raceway Curvature. Raceway curvature is an expression that defines the relation-ship between the arc of the raceway's track radius and the arc formed by the slightly smaller ball that runs in the raceway. It is simply the track radius of the bearing race-way expressed as a percentage of the ball diameter. This number is a convenient index of fit between the raceway and ball. ill 54 illustrates this relationship. Track curvature values typically range from approximately 52 to 58 percent. The lower-percentage, tight-fitting curvatures are useful in applications where heavy loads are encountered. The higher-percentage, loose curvatures are more suitable for torque-sensitive applications. Curvatures less than 52 percent are generally avoided because of excessive rolling friction that's caused by the tight conformity between the ball and raceway. Values above 58 percent are also avoided because of the high stress levels that can result from the small ball-to-raceway conformity at the contact area. ill. 54 Relationship of track radius to ball diameter. Contact Angle. The contact angle is the angle between a plane perpendicular to the ball-bearing axis and a line joining the two points where the ball makes contact with the inner and outer raceways. The contact angle of a ball bearing is determined by its free radial play-value, as well as its inner and outer track curvatures. The contact angle of thrust-loaded bearings provides an indication of ball position inside the raceways. When a thrust load is applied to a ball bearing, the balls will move away from the median planes of the raceways and assume positions somewhere between the deepest portions of the raceways and their edges. ill 55 illustrates the concept of contact angle by showing a cross-sectional view of a ball bearing that's loaded in pure thrust. ill. 55 Contract angle for bearing loaded in pure thrust. Free Angle and Angle of Misalignment. As a result of the previously described looseness, or play, which is purposely permitted to exist between the components of most ball bearings, the inner ring can be cocked or tilted a small amount with respect to the outer ring. This displacement is called the free angle of the bearing, and corresponds to the case of an unmounted bearing. The size of the free angle in a given ball bearing is determined by its radial play and track curvature values. ill 56 illustrates this concept. For the bearing mounted in an application, any misalignment present between the inner and outer rings (housing and shaft) is called the angle of misalignment. The misalignment capability of a bearing can have positive practical significance because it enables a ball bearing to accommodate small dimensional variations which may exist in associated shafts and housings. A maximum angle of misalignment of 1 /.4 is recommended before bearing life is reduced. Slightly larger angles can be accommodated, but bearing life won't be optimized. ill. 56 Free angle of bearing. 11.6 Materials Bearing Materials Chrome steel. A bearing steel used for standard ball-bearing applications in uses and in environments where corrosion resistance isn't a critical factor. The most commonly used ball-bearing steel in such applications is AISI 52100 or its equivalent. Due to its structure, this is the material chosen for extremely noise-sensitive applications. DD400 0.7% C; 13% Cr. A 400-series martensitic stainless steel combined with a heat-treating process was exclusively developed for use in miniature and instrument bearings. Bearings manufactured from DD meet the performance specifications of such bearings using AISI 440C martensitic stainless steel, and it's equal or superior in hardness, superior in low-noise characteristics, and at least equivalent in corrosion resistance. These material characteristic advantages make for lower torque, smoother running, and longer-life bearings. retainer, also referred to as the cage or separator, is the component part of a ball bearing that separates and positions the balls at approximately equal intervals around the bearing's raceway. There are two basic types: the crown or open-end design and the closed ball-pocket design. The most common retainer is the two-piece closed retainer, commonly called a ribbon retainer. The open-end design, or crown retainer, as shown in ill. 57, is of metal material. Crown retainers manufactured from molded plastics are available for some sizes. The metal retainer, constructed of hardened stainless steel, is very lightweight and has coined ball pockets which present a hard, smooth, low-friction contact surface. The closed-pocket design (two-piece construction) with clinching tabs, as out-lined in ill. 58, is a standard design for most miniature and instrument-sized ball bearings. The use of loosely clinched tabs is favorable for starting torque, and the closed-pocket design provides good durability required for various applications. Shields and seals are necessary to provide optimum ball-bearing life by retaining lubricants and preventing contaminants from reaching central work surfaces. Different types of closures can be supplied on the same bearing, and nearly all are removable and replace-able. They are manufactured with the same care and precision that goes into the ball bearings. The following are descriptions of the most common types of shields and seals available. Z- and H-type shields designate noncontacting metal shields. Z-type shields (ill. 59) are the simplest form of closure and , for most bearings, are removable. H-type shields (ill. 60) are similar to Z-types but are not removable. ill. 57 Standard one-piece crown retainer. ill. 58 Metal two-piece closed-pocket ribbon retainer. ill. 59 Two Z-type shields (removable). It is advantageous to use shields rather than seals in some applications because there are no interacting surfaces to create drag. This results in no appreciable increase in torque or speed limitations, and operation can be com-pared to that of open ball bearings. D-type contacting seals (ill. 61) consist of a molded Buna-N rubber lip seal with an integral steel insert. While this closure type provides excellent sealing characteristics, several factors must be considered for its application. The material normally used on this seal has a maximum continuous operating temperature limit of 25 0 F (12 0 C).Although it's impervious to many oils and greases, consideration must be given to lubrication selection. It is also capable of pro-viding a better seal than most other types by increasing the seal lip pressure against the inner ring OD. This can result in a higher bearing torque than with other types of seals and may cause undesirable seal lip heat buildup in high-speed applications. S-type noncontacting seals are constructed in the same fashion as the D-type seals. This closure type has the same temperature limitation of 25 0 F (12 0 C). It also is impervious to many oils and greases, but the same considerations should be noted on lubrication selection. The S-type seal (ill. 62) is uniquely designed to avoid contact on the inner land, significantly reducing torque over the D-type configuration. L-type seals (ill. 63) are fabricated from glass-reinforced Teflon. When assembled, a very small gap exists between the seal lip and the inner ring OD. It is common for some contact to occur between these components, resulting in an operating torque increase. The nature of the seal material serves to keep this torque increase to a minimum. In addition, the use of this material allows high operating temperatures with this configuration. ill. 60 Two H-type shields (nonremovable). ill. 61 Two D-type seals (contacting rubber). ill. 62 Two S-type seals (noncontacting rubber). ill. 63 Two L-type seals (nonflexed Teflon). ill. 64 Two SSD21-type seals (labyrinth-design seal). The SSD21-type seals (ill. 64) have the same operating characteristics as the D- and S-type seals, resulting in the same considerations of temperature limitation and lubricant selection. The SSD21-type seal is comprised of a noncontacting rubber seal combined with a labyrinth-design inner ring. The labyrinth-design configuration creates an extended path to the raceway, minimizing the tendency for contaminants to creep into the ball bearing. 11.7 Lubrication Lubricant Types. Oil is the basic lubricant for ball bearings. Previously, most lubricating oil was refined from petroleum. Today, however, synthetic oils such as diesters, silicone polymers, and fluorinated compounds have found acceptance because of improvements in properties. Compared to petroleum-based oils, diesters in general have better low-temperature properties, lower volatility, and better temperature/ viscosity characteristics. Silicones and fluorinated compounds possess even lower volatility and wider temperature/viscosity properties. Virtually all petroleum and diester oils contain additives that limit chemical changes, protect the metal from corrosion, and improve physical properties. Grease is an oil to which a thickener has been added to prevent oil migration from the lubrication site. It is used in situations where frequent replenishment of the lubricant is undesirable or impossible. All of the oil types mentioned in the next subsection can be used as grease bases to which are added metallic soaps, synthetic fillers, and thickeners. The operative properties of grease depend almost wholly on the base oil. Other factors being equal, the use of grease rather than oil results in higher starting and running torque and can limit the bearing to lower speeds. Oils and Base Fluids. Petroleum lubricants have excellent load-carrying abilities, but are usable only at moderate temperature ranges [ - 25 to 25 0 F (- 32 to 12 1 C)]. Greases that use petroleum oils for bases have a high dN capability. Greases of this type are recommended for use at moderate temperatures, light to heavy loads, and moderate to high speeds. While super-refined petroleum lubricants are usable at higher temperatures than petroleum oils [ - 65 to 35 0 F (- 54 to 17 7 C)], they still exhibit the same excellent load-carrying capacity. This further refinement eliminates unwanted properties, leaving only the desired chemical chains. Additives are introduced to increase the oxidation resistance, etc. The diesters are probably the most common synthetic lubricants. They don't have the film-strength capacity of petroleum products, but do have a wide temperature range [ - 65 to 35 0 F (- 54 to 17 1 C)] and are oxidation resistant. Synthetic hydrocarbons are finding a greater use in the miniature and instrument ball-bearing industry because they have proved to be a superior general-purpose lubricant. Silicone lubricants are useful over a wide temperature range [- 100 to 40 0 F (- 73 to 20 4 C)], but don't have the film strength of petroleum types and other synthetics. It has become customary in the instrument and miniature bearing industry, in recent years, to de-rate the dynamic load rating Cr of a bearing to one-third of its normal value if a silicone product is used. Per-fluorinated polyether oils and greases have found wide use where high temperature stability and /or chemical inertness are required. This specialty lubricant does not have the film strength of petroleum or diester products. However, it does have better film strength than silicone lubricants. Lubrication Methods. Grease packing to approximately one-quarter to one-third of a ball bearing's free volume is one of the most common methods of lubrication. Volumes can be controlled to a fraction of a percent for precision applications by special lubricators. In some instances, people have used bearings that were to be lubricated 100 percent full of grease. Excessive grease is as detrimental to a bearing as insufficient grease. It causes shearing, heat buildup, and deterioration through constant churning which can ultimately result in bearing failure. Centrifuging an oil-lubricated bearing removes excess oil and leaves only a very thin film on all surfaces. This method is used on very low torque bearings and can be specified for critical applications. Operating Speed. When petroleum or synthetic ester oils are used, the maxi-mum speed Nmax is dictated by the ball cage material and design or the centripetal ball loads rather than by the lubricant. For speed-limit values Nmax, the Nmax/ fn values shown in product listings must be multiplied by the fn values shown in Table 5. The following method may be used to select a lubricant. Step 1. Define the temperature range of the application, including the environ-mental temperature plus any heat rise from motors, etc. Refer to Table 9 and select the proper lubricant base for the maximum and minimum operating temperature. TABLE 9 Relationship Between Lubricants, dN Values, and Temperature Ranges Type dN Temperature range Silicone 200,000 - 100 to 40 deg F ( - 73 to 20 4 C) Diester 400,000 - 65 to 35 0 F ( - 54 to 17 1 C) Petroleum 600,000 - 25 to 25 0 F ( - 32 to 12 1 C) When selecting a base fluid type, the fluid with the greatest film support is the preferred choice. Refer to the description of lubricant types for individual capabilities. Step 2. Determine the speed of the bearing and calculate the dN value (see next subsection, Speed Factor). elect the lubricant type that will operate within the dN speed factor, referring to Table 9. Step Knowing the dN value, determine the proper viscosity of the lubricating oil or the base oil of the grease (ill. 65). Since grease is approximately 80 per-cent oil, it's necessary to determine the viscosity of the oil for any high-speed application. Improper selection can result in rapid deterioration of the base oil and failure of the unit. Step 4. Once you have determined these factors, the lubricant selection has been narrowed to the type of base oil, the operating temperature, and the oil viscosity range for a particular dN value (see next subsection, Speed Factor). Next, determine whether a grease or oil is needed for the application. Then, individual lubricants should be examined to determine their suitability for the application. Speed Factor. The maximum usable operating speed of a grease lubricant is dependent on the type of oil. The speed factor is a function of the bore of the bearing d, mm, and the speed of the bearing N, rpm: speed factor = d x N x dN (3.4) There are many lubricants available for ball bearings. Refer to Table 10. Dynamic Load Ratings and Fatigue Life Dynamic Radial Load Rating. The dynamic radial load rating Cr for a radial ball bearing is a calculated, constant radial load which a group of identical bearings can theoretically endure for a rating life of 1 million revolutions. The dynamic radial load rating is a reference value only. The base rating-life value of 1 million revolutions has been chosen for ease of calculation. Since applied loading equal to the basic load rating tends to cause permanent deformation of the rolling surfaces, such excessive loading isn't normally applied. Typically, a radial load that corresponds to 15 percent or more of the dynamic radial load rating is considered heavy loading for a ball bearing. In cases where loading of this degree is required, consult a bearing manufacturer's application engineer for information regarding bearing life and lubricant recommendations. ill. 65 Speed factor. TABLE 10 Commonly Used Lubricant Types Code | Brand name | Basic oil type | Operating temperature | Uses LO1 LG20 LG39 LY48 LY75 LY83 LY121 Windsor L245X low-(MIL-L-6085A) Exxon Beacon 325 Exxon Andok C Mobil 28 (MIL-G-81322) Chevron SRI-2 Shell Alvania X2 Kyodo Multemp smooth-SRL Ester oil Channeling grease: mineral oil and sodium soap thickener, Channeling grease: mineral oil and sodium soap thickener Synthetic oil and clay thickener , Mineral oil and urea soap high-thickener, Mineral oil and general-lithium, soap thickener Ester oil and lithium soap thickener -60 to + 25 0 F ( 51 to 12 1 C) -60 to + 25 0 F ( 51 to 12 1 C) . 20 to + 25 0 F ( 29 to 12 1 C) . 65 to + 35 0 F ( 54 to 17 7 C) . 20 to + 35 0 F ( 29 to 17 7 C) . 30 to 25 0 F ( 29 to 12 1 C) 40 to 30 0 F (3 to 14 9 C) Low-torque, low-(speed instrument oil; rust preventative, General-purpose grease., Low migration; general office equipment applications Good heat resistance with low torque; throttle body applications Good heat resistance, high-thickener speed grease; power tool and vacuum cleaner motor applications Long-life, general-lithium use grease; power tool applications Low-noise, smooth-SRL running grease; general motor applications Rating Life. The rating life L10 of a group of apparently identical ball bearings is the life in millions of revolutions, or number of hours, that 90 percent of the group will complete or exceed. For a single bearing, L10 also refers to the life associated with 90 percent reliability. The median life L50, the life which 50 percent of the group of ball bearings will complete or exceed, is usually not greater than 5 times the rating life. Rating life is calculated as follows: L10 = ( Cr/Pr)^ 3 where L10 = rating life Cr = dynamic radial load rating, kgf Pr = dynamic equivalent radial load, kgf The dynamic radial load rating Cr can be found from product listings. The dynamic equivalent load must be calculated according to the following procedure: Pr = XFr + YFa (3.5) where X and Y are obtained from Table 11 Fr = radial load on the bearing during operation, kgf Fa = axial load on the bearing during operation, kgf TABLE 11 Axial Load Variables The L10 life can be converted from millions of revolutions to hours using the rotation speed. This can be done as follows: L10, millions of revolutions × 1,000,000/rpm × 60 = L10, hours (3.6) To convert pounds to kilograms force, divide by 0.45359: kgf = lb / 0.4359 kgf/lb Life Modifiers. For most cases, the L10 life obtained from the equation discussed previously will be satisfactory as a bearing performance criterion. However, for particular applications, it might be desirable to consider life calculations for different reliabilities and /or special bearing properties and operating conditions. Reliability adjustment factors, bearing material adjustment, and special operating conditions are discussed in the following subsections. Bearing Material. Manufacturers recommend that radial load ratings published for chrome steel be reduced by 20 percent for stainless steel. This is a conservative approach to ensure that bearing capacity isn't exceeded under the most adverse conditions. This is incorporated in the a2 modifier, as shown in Table 12. Reliability Modifier. Where a more conservative approach than conventional rating life L10 is desired, the American Bearing Manufacturers Association (ABMA) offers a means for such estimates. Table 12 provides selected modifiers a2 for calculating failure rates down to 1 percent ( L1). TABLE 12 Reliability Versus Material Life Modifier a2 Required reliability, %| Ln | Value of a2 (Chrome DD) 90 L10 1.00 0.50 95 L5 0.62 0.31 96 L4 0.53 0.27 97 L3 0.44 0.22 98 L2 0.33 0.17 99 L1 0.21 0.11 Other Life Adjustments. The conventional rating life often has to be modified as a consequence of application abnormalities, whether they be intentional or unknown. Seldom are loads ideally applied. The following conditions all have the practical effect of modifying the ideal, theoretical rating life L10. Vibration and /or shock-impact loads Angular misalignment High-speed effects Operation at elevated temperatures Fits Internal design Oscillatory Service Life. Frequently, ball bearings don't operate with one ring rotating unidirectionally. Instead, they execute a partial revolution, reverse motion, and then repeat this cycle, most often in a uniform manner. Efforts to forecast a reliable fatigue life by simply relating oscillation rate to an "equivalent" rotational speed are invalid. The actual fatigue life of bearings operating in the oscillatory mode is governed by four factors: applied load, angle of oscillation, rate of oscillation, and lubricant. Lubricant Life. In many instances a bearing's effective life is governed by the lubricant's life. This is usually the case where applications involve very light loads and /or very slow speeds. With light loads and /or slow speeds, the conventional fatigue-life forecast will be unrealistically high. The lubricant's ability to provide sufficient film strength is sustained only for a limited time. This is governed by the following factors:
11.8 Static Capacity Static Radial Load Rating. The static radial load rating Cor is the radial load which a non-rotating ball bearing will support without damage, continuing to provide satisfactory performance and life. The static radial load rating is dependent on the maximum contact stress between the balls and either of the two raceways. The load ratings shown were calculated in accordance with the ABMA standard. The ABMA has established the maximum acceptable stress level resulting from a pure radial load in a static condition to be 4.2 GPa (609,000 lb/in 2 ). Static Axial Load Capacity. The static axial load capacity is the axial load which a non-rotating ball bearing will support without damage. The axial static load capacity varies with bearing size, bearing material, and radial play. Radial static load ratings and thrust static load ratings in excess of published Cor values have practical applications where smoothness of operation and /or low noise are not of concern. Properly manufactured ball bearings, when used under controlled shaft and housing fitting practices, can sustain significantly greater permanent deformation, such as brinells, than deformations associated with normal static load ratings. 11.9 Preloading Ball-bearing systems are preloaded for the following reasons: To eliminate radial and axial looseness To reduce operating noise by stabilizing the rotating mass To control the axial and radial location of the rotating mass and to control movement of this mass due to external force influences To reduce the repetitive and non-repetitive runout of the rotational axis To reduce the possibility of damage due to vibratory loading To increase stiffness Spindle motors and tape guides are examples of applications where preloaded bearings are used to accurately control shaft position when external loads are applied. As the name implies, a preloaded assembly is one in which a bearing load (normally a thrust load) is applied to the system so the bearings are already carrying a load before any external load is applied. There are essentially two ways to preload a ball-bearing system, by using a spring or by using a solid stack of parts. Ill. 66 Spindle assembly using compression coil spring, with shaft rotation. Spring Preloading. For many applications, one of the simplest and most effective methods of applying a preload is by means of a spring. This can consist of a coil spring or a wavy washer which applies a force against the inner or outer ring of one of the bearings in an assembly. When a spring is used, it's normally located on the non-rotating component; i.e., with shaft rotation, the spring should be located in the housing against the outer rings. Springs can be very effective when differential thermal expansion is a problem. In the spindle assembly shown in ill. 66, when the shaft becomes very hot and expands in length, the spring will move the outer ring of the left bearing and thus maintain system preload. Care must be taken to allow for enough spring movement to accommodate the potential shaft expansion. Since, in a spring, the load is fairly consistent over a wide range of compressed length, the use of a spring for preloading negates the necessity for holding tight location tolerances on machined parts. E.g., retaining rings can be used in the spindle assembly, thus saving the cost of locating shoulders, shims, or threaded members. Normally, a spring preload wouldn't be used where the assembly is required to withstand reversing thrust loads. Solid-Stack Preloading. When precise location control is required, as in a precision motor (ill. 67), or a flanged tape guide (ill. 68), a solid preloading system is indicated. A solid-stack "hard" or "rigid" preload can be achieved in a variety of ways. Theoretically, it's possible to preload an assembly by tightening a screw, as shown in ill. 68, or inserting shims, as shown in ill. 69, to obtain the desired rigidity. It should be noted that care must be taken when using a solid-stack preloading system with miniature and instrument bearings. Overload of the bearings must be avoided so that the bearings are not damaged during this process. Ill. 67 Rotor outer-ring spacer, with stator mount as inner ring. Ill. 68 Typical tape-guide design using screw and washer for solid preloading by clamping inner rings, with outer-ring rotation. Ill. 69 Shims to apply preload. Preload Levels. Preloading is an effective means of positioning and control-ling stiffness because of the nature of the ball-raceway contact. Under light loads, the ball-raceway contact area is very small, and so the amount of yield or definition is substantial with respect to the amount of load. As the load is increased, the ball-raceway contact area increases in size (the contact is in the shape of an ellipse) and so provides increased stiffness or reduced yield per unit of applied load. This is illustrated in the single-bearing deflection curve shown in ill. 70. When two bearings are preloaded together and subjected to an external thrust load, the axial-yield rate for the pair is drastically reduced because of the preload and the interaction of the forces exerted by the external load and the reactions of the two bearings. As can be seen by the lower curve in ill. 70, the yield rate for the preloaded pair is essentially linear. Ill. 70 Single-bearing deflection curve. Miniature and instrument bearings are typically built to accept light preloads normally ranging from 0.25 lb to not more than 10 lb. TABLE 13 Recommended Fits Typical application | Shaft fit | Shaft diameter | Housing fit | Housing diameter General application-inner ring rotation (inner ring press fit, outer ring loose fit); General application-outer ring rotation (inner ring loose fit, outer ring press fit); Tape-guide roller Drive motor (spring preload) Precision synchro or servo Potentiometer Encoder spindle Gear reducer Light-duty mechanism Clutches, brakes (inner race floats) Pulleys, rollers, cam followers (outer race rotates) 11.10 Assembly and Fitting Procedure The operating characteristics of a system can be drastically affected by the way in which the ball bearings are handled and mounted. A bearing which has been dam-aged due to excessive force or shock loading during assembly, or which is fitted too tight or too loose, may cause the device to perform in a substandard manner. By following a few general guidelines during the design of mating parts and by observing some basic cautions in the assembly process, the possibility of producing malfunctioning devices can be considerably reduced. Table 13 lists recommended fits for most normal situations. There are four cautions which must be observed. 1. When establishing shaft or housing sizes, the effect of differential thermal expansion must be accounted for. Table 13 assumes stable operating conditions, so if thermal gradients are known to be present or dissimilar materials are being used, the room temperature fits must be adjusted so that the proper fit's attained at operating temperature. 2. When miniature and instrument ball bearings are interference-fitted (either intentionally or as a result of thermal gradients), the bearing radial play can be estimated to be reduced by an amount equal to 80 percent of the actual diametrical interference fit. This 80 percent figure is conservative, but is of good use for design purposes. Depending on the materials involved, this factor will typically range from 50 to 80 percent. The following is an example of calculating loss of radial play: Radial play of bearing: 0.0002 in Total interference fit: 0.0003 in (tight) 80 percent of interference fit (0.0003 in × 80%) 0.00024 in Theoretical resultant radial play of bearing 0.00004 in (tight) Theoretically, this bearing could be operating with negative radial play. A bearing operated in an excessive negative radial-play condition will perform with reduced life. However, the preceding calculation is for design only, and does not take into account housing material, shaft material, or surface finish of the housing or shaft surfaces. As an example, if the finish of the shaft surface ring and shaft will be absorbed by the deformation of the shaft surface, this will serve to reduce the overall interference fit. Thus, the radial play of the bearing won't be reduced as much as is shown in the preceding calculation. Table 13 is based on the use of bearings of ABEC 5 or better tolerance level. If the outer or inner ring face is to be clamped or abutted against a shoulder, care must be taken to make sure that this shoulder configuration provides a good mounting surface: The shoulder face must be perpendicular to the bearing mounting seat. The maximum recommended permissible angle of misalignment is 1/4 ° . The corner between the mounting diameter and the face must have an under-cut or a fillet radius r no larger than that shown in ill. 51. The shoulder diameter must meet the requirements shown in Table 14. 4. Assembly technique is extremely critical. After the design is finalized and assembly procedures are being formulated, the bearing static capacity Cor becomes extremely important. It is easy, for instance, to exceed the 3-lb capacity of a DDRI-2 during assembly. After assembly to the shaft, damage can be done either by direct pressure or by moment load while the bearing and shaft subassembly is being forced into a tight housing. A few simple calculations will underscore this point. Adequate fixturing should always be provided for handling and assembling precision bearings. This fixturing must be designed so that when assembling the bearing to the shaft, force is applied only to the inner ring, and when assembling into the housing, force is applied only to the outer ring. Further, the fixturing must preclude the application of any moment or shock loads which would be transmitted through the bearing. Careful attention to this assembly phase of the total design effort can prevent many problems and provide savings when production starts. TABLE 14 Recommended Shoulder Diameter Basic size | Minimum shaft shoulder diameter, in | Maximum housing shoulder diameter, in DDRI-2 0.060 0.105 DDRI-2 1/.2 0.071 0.132 DDRI-3 0.079 0.164 DDRI-4 0.102 0.226 DDRI-3332 0.114 0.168 DDRI-5 0.122 0.284 DDRI-418 0.148 0.226 DDRI-518 0.153 0.284 DDRI-618 0.153 0.347 DDR-2 0.179 0.325 DDRI-5532 0.180 0.288 DDR-1640 0.210 0.580 DDRI-5632 0.210 0.288 DDRI-6632 0.216 0.347 DDR-3 0.244 0.446 DDR-1650 0.250 0.580 DDR-1950 0.250 0.700 DDR-1960 0.290 0.700 DDRI-614 0.272 0.352 DDRI-814 0.284 0.466 DDR-4 0.310 0.565 DDRI-1214 0.322 0.678 DDR-2270 0.325 0.810 DDR-2280 0.370 0.810 DDRI-8516 0.347 0.466 DDRI-1038 0.435 0.565 DDRI-1438 0.451 0.799 DDRI-1212 0.560 0.690 DDRI-1458 0.665 0.835 DDRI-1634 0.790 0.960 |
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