Thumb for Mechanical Engineers 2011 Part 6 potx

140 Rules of Thumb for Mechanical Engineers F 2.0xsinr 2 +- d a= dbev = 2.0 x a x sin I? (24) The size of bevel gear teeth is calculated the same as with spur and helical gears except a size factor is included. When an average value of size factor is included in Equa- tion l l, the result is Equation 25. Equation 25 can be used to estimate the tooth size. The tapered tooth of the bevel gear changes in size across the face width. The diametral pitch calculated from Equation 25 is the size at the mid-face (mean diameter). The standard nomenclature for bevel gears defines the diametral pitch at the large end. Due to the different methods of manufacturing, it is not necessary to round the calculated pitch value to a whole number, but only to adjust the value to the size at the large end of the gear, per Equation 26. n x d x F x J x Sat Pd = 390,000 x P x cd NPbev = . Pdbev The number of teeth must be adjusted to a whole number, and then the pitch Pdbev must be adjusted so the number of teeth will fit the diameter. After rounding the number of gear teeth to a whole number, the gear diameter can be calculated: Bevel Gear Sizing Example. Estimate the size of bevel gearing for a steam turbine running 4,200 rpm to drive a centrifugal pump re- quiring 350 hp and running at 2,000 rpm. Most spiral bevel gears are carburized and hardened-this is best for the higher rpm in this application. For normal practice the pressure angle will be 20 degrees and the spiral angle 30 degrees. Start with the gear ratio: 4,200/2,000 = 2.1 to 1. From Table 3, for gears surface-hardened to 58 R,, the From Table 4, select C,. Based on smooth driving and For an initial value of G, assume a diameter d of 4.0 inch- Kall = 600. driven machines and extra life, choose C, = 1.25. es. Per Equation 5 calculate the pitch line velocity: v = .262 x 4.0 x 4,200 = 4,402 ft/min Use standard commercial practice, which would be to cut the gears, then carburize and harden, and then finish by lap- ping. From Table 5, use Cv = 1.26, and from Table 6, use C, = 1.30. Therefore, from Equation 4: C,j = 1.25 x 1.26 x 1.30 = 2.05 Now use Equations 19 and 20 to determine the face-to- diameter factor Fd based on the gear ratio of the bevels: 1 r = tan-’ - = 25.46 2.1 0.2 sin 25.46 Fd = = 0.47 Now the equivalent mean diameter d can be calculated using Equation 2 1. 100,000x50x2.05 2.1+1 d=[ .47 x 600 x 4,200 ( F)]”’ = 4.485 From Equation 20: F = 4.485 x 0.47 = 2.11 The cone distance a is calculated per Equation 23: - 6.27 1 2.109 + 4.485 a= 2x sin25.46 2 The actual pitch diameter of the bevel pinion is calculated from Equation 24: Gears 141 dbev = 2.0 x 6.271 x sin 25.46 = 5.392 Use Equation 5 to calculate the actual pitch line velocity: v = .262 x 5.392 x 4,200 = 5,933 Wmin This is close enough to 5,000 Wmjn to validate the orig- inal C, factor of 1.26. Proceed to Equation 25 to estimate the tooth size at the mean diameter. From Table 3, the value of Sa can be selected as 65,000 psi. Assume 22 teeth on the pinion and select J factors from the helical section since this gear set is spiral bevel, the equivalent of helical in a parallel shaft gear set. Jp = .520 and Jg = .560 Because the material strengths of the pinion and the gear are the same, the pinion with the smaller J factor will dictate the tooth size. 4,200 x 4.485 x 2.109 x .520 x 65,000 = 4.799 Pd = 390,000 x 350 x 2.05 From Equation 26 the actual diametral pitch at the large end of the pinion is: = 3.991 4.485 x 4.799 5.392 'dbev = From Equation 27: Nmev = 5.392 x 3.99 = 21.52 Obviously, the number of teeth must be an integer number, so use Nev = 22. Working Equation 26 backwards, de- fine the final diametral pitch: pdbev = - 22 = 4.08 5.392 From Equation 28: NGk., = 22 x 2.1 = 46.2 Use 46 teeth. From Equation 29: = 11.275 46 4.08 Dbev = - Thus, the size of the bevel pinion and gear and the num- bers of teeth on the gears are estimated. When this gear set is rated by AGMA standards, the results show 423 HP capacity which shows that the estimate is reasonable and conservative. Cylindrical Worm Gear Design Worm gears have a number of unique characteristics besides the arrangement of perpendicular shafts offset by the center distance. The input worm is basically a screw thread which makes one revolution to advance the gear wheel one tooth. This makes it possible to have very high gear ratios, especially since the worm can be made with mul- tiple start threads. Due to the sliding nature of the tooth contact, the efficiency can be poor, with typical values being 90% to 50% with the lower values in the high-ratio designs. This characteristic can be used to make a self-locking drive in which the output gear cannot drive the input worm. This is generally the case when the lead angle is 5 degrees or less. Caution must be exercised if this characteristic is desired, because the difference between dynamic and static coefficient of friction can cause a self-locking drive to unlock due to vibration or any slight initiation from the input. Since the efficiency can be low, it is best to think in terms of two different power ratings: the output power to drive the load and the input power which also includes the friction loss load. Input Power Rating. The rating equation has two parts: the first is the transmitted power and the second is the friction power loss in the mesh. W,Dn VWf P= +- 126,000 mg 33,000 The maximum tooth load is: W, = 900 x Do.' F,I&K, (31) 142 Rules of Thumb for Mechanical Engineers 0.4 0.3 0.2 0.1 This is based on a hardened and ground worm running with a centrifugal-cast bronze wheel with a physical face width of 6 inches or less. A wider face, up to 12 inches, would be derated up to 20%. For a chill-cast bronze wheel, derate by 2096, and by 30% for a sand-cast bronze wheel. F, is the effective face width, which is the actual face width but not exceeding % of the mean diameter of the worm. K, is the ratio correction factor taken from Figure 2. The velocity factor, K,, is a function of sliding velocity and can be read from Figure 3. The sliding velocity is: Velocity Functions 0.5 I I Series 2 0 500 1 Sliding Velocity Figure 3. Velocity functions. ll nd 12 cos h V= Ratio Correction Factor where y= lead angle of the worm thread at the mean diameter With the tooth load calculated, the friction force can be calculated: Figure 2. Ratio correction factor. PWt cos h cos 0 w, = (33) The design of the finished gear drive must allow for cooling the heat from the friction part of the input power. Worm drives frequently have cooling fans on the high-speed input shafts and cooling fins cast onto the housing. It should also be noted that surface finish is critical on the tooth surfaces, and lubricating oil properties are very important. Most gears are made of alloy steel. The main criteria for selecting material is the fact that the load capacity of the gear set is proportional to the hardness of the material. There are two major material categories: surface-hardened and through- hardened. Through-hardened alloy steel is normally limited to the range of 38 Rc maximum. One characteristic of through-hardened gear sets that might not be expected is the hardness relationship between the pinion and the gear. For best life and durability, the pinion should be at least 2 Rc points harder than the gear. When both members are the same size-one-to-one ratio-equal hardness works satisfactorily. Surface hardening can increase the surface to as much as 60 Rc while the softer core maintains a ductility and toughness. Of the various methods that can be used to surface-harden gears, three are most common. Carburizing is the most common method used to achieve the maximum hardness and gear load capacity. The greatest drawback with carburizing is the significant geometric distortion introduced during the quenching operation. This requires a finishing operation to restore the dimensional accuracy in almost all designs. As an alternative, nitriding can achieve surface hardness in the range of 50 to 60 Rc depending on the steel alloy used. The distortion is usually very low so that finishing is not generally required. The nitriding operation requires a long fur- nace he4 to 120 hours in proportion to the case depth- and therefore is normally limited to smaller case depth used for smaller-size teeth and may be impractical for gears with large-size teeth. Znduction hardening takes a number of forms and can be used with a wide range of case depths. Dis- tortion is usually low. This process requires careful development and, sometimes, tool development to assure con- sistent quality. Without proper development, the result may give good surface and core hardness but may have problems with ductility and fatigue life. Gears 143 Summary of Gear Types With so many types and arrangements of gearing available, a summary is provided below. Parallel Shaft This is the most common type of gear and, as the label implies, this type of gear set operates with the axes of rotation parallel to each other. The most common use of parallel shaft gears is to change the speed, and torque, of the driven shaft relative to the driving shaft. The driven shaft also rotates in the opposite direction (unless one of the gears is an internal gear). Unless the two gears are equal in diameter, the smaller diameter member is called a pinion. Spur. Spur gears are the most basic type of gear. The gear teeth are parallel to the axis of the shaft. Helical. As the name implies, the teeth on a helical gear have a lead angle relative to the axis of rotation and follow the curve of a helix across the face width of the gear. The tooth load is shared by more pairs of teeth and can be transferred from tooth to tooth more smoothly than the more simple spur gear. Therefore, when all other parame- ters are equal, a helical gear set can carry more load and run quieter with less vibration than a spur gear set. While the basic cost to manufacture a helical gear is usually no greater than a spur, there is a penalty in the form of a thrust com- ponent to the gear reaction loads that must be supported by the shaft and bearings. Double Helical (Herringbone). A way to counter the thrust loads of the helical gear is the double helical gear. This is accomplished by dividing the face width of the gear into two halves and using the opposite hand of helix for each half. In order to use conventional manufacturing machines, a cutter runout space must be provided between the two halves; this adds to the overall width of the gear and makes it bigger than the equivalent helical gear. With special cut- ting machines, the space between the two halves can be eliminated, and this type of gear is called herringbone. Some of the finishing methods used to improve the capacity and precision of gears cannot be used with herringbone gears. Since nothing in this world is perfect, the gear tooth circle is never perfectly concentric with the shaft axis of rotation, and this eccentricity contributes to vibration and dynamic load. This is of particular importance in the case of double helical gears because the two halves of the face each have specific runouts and combine to create an additional axial runout. While this is not generally a problem, it can require additional manufacturing effort. It is also impera- tive that the gear shaft bearing and coupling designs allow the two halves of the gear face to share the tooth load equally, as any external thrust loads that react through the gears will cause an overload in one half. Bevel Bevel gears have the teeth formed on a cone in place of a cylinder, and the axes of rotation intersect rather than being parallel lines. The most common arrangement has the axes intersecting at 90 degrees; however, other angles can be used, such as seen in Vee drives for boat transmissions. Straight Tooth. The straight tooth bevel is the spur gear of the bevel family. Being on a cone, the teeth are tapered in thickness from the inner end of the face to the outer end. Spiral Bevel. The spiral bevel gear is the equivalent of the helical gear on a cone. While the teeth on a straight bevel follow a ray line along the cone from one end of the face to the other, the spiral bevel tooth is modified in two ways. The tooth is set at a spiral angle, similar to the helix angle of the helical gear, and it is curved with the radius of the cutter head used to hold the blades that cut the teeth. Zerol. The zero1 is a special form of the spiral bevel that has the teeth curved with the cutter radius, but with a spiral angle of zero degrees. The curved tooth form gives some of the smoother-action characteristics of the spiral bevel; but with no spiral angle, the thrust reaction is not transmitted to the bearings. Hypoid. The hypoid gear set is very similar to the spiral bevel set except the input pinion axis is offset so that the axes no longer intersect. More sliding is introduced in the tooth contact which results in a slight reduction in efficiency, but some geometric shaft arrangement problems can be solved. Worm Worm gear sets have their input and output axes perpendicular and offset by the center distance. While this 144 Rules of Thumb for Mechanical Engineers arrangement may be an advantage for some applications, the worm gear type is more frequently chosen for other characteristics. Very high gear ratios can be achieved in a single gear stage. However, efficiency goes down as ratio goes up. This is sometimes used to advantage since high- ratio worm sets are the only gears normally designed to be self-locking. A self-locking set acts as a brake, and the gears lock if the output shaft tries to drive the input. Buying Gears and Gear Drives One of the most important considerations in purchasing gears is to work with a reliable and experienced vendor. A good source of information on suppliers is the American Gear Manufacturers Association (see References). It is also important to inform the vendor of all possible data about the requirements, application, and use planned for the gears or drive. Keep the specification of detailed gear data to a minimum and allow the vendor to apply his experience to help you get the best possible product. However, the most detailed possible design information should be required to be submitted with the vendor’s quotation. The idea is to give the vendor freedom to offer the most appropriate product but to require detailed data with the quotation for evalua- tion in selecting the best offering. Many times, a second quotation will be in order. REFERENCES 1. American Gear Manufacturers Association, AGMA and IS0 Standards, 1500 King St., Suite 201, Alexandria, VA 223 14. 2. Dudley, Darle W., Practical Gear Design. New York McGraw-Hill, Inc., 1984. 3. Drago, Raymond J., Fundamentals of Gear Design. 4. Townsand, Dennis P., Dudley’s Gear Handbook New Stoneham, MA: Butterworth Publishers, 1988. York McGraw-Hill, Inc., 199 1. Bearings C . Richard Lenglade. Jr., Development Engineer. Allison Engine Company Types of Bearings 146 Ball Bearings 146 Roller Bearings 147 Standardization 149 Materials 15 1 Rating and Life e 152 ABMA Definitions 152 Fatigue Life 153 Life Adjustment Factors 154 Load and Speed Analysis 156 Equivalent Loads 156 Contact Stresses 157 Preloading 157 Special Loads 158 Effects of Speed 159 Lubrication 160 General 160 Oils 161 Greases 161 Lubricant Selection 162 Lubricating Methods 163 Relubrication 164 Cleaning, Preservation, and Storage 165 Mounting 166 Shafting 166 Housings 169 Bearing Clearance 172 Seals 174 Sleeve Bearings 175 References 177 145 146 Rules of Thumb for Mechanical Engineers TYPES OF BEARINGS There are two general categories of bearings: rolling element bearings and journal bearings. Most of this chapter is devoted to rolling element bearings because, for most industrial equipment, these are the most common bearings in usage. On the other hand, journal bearings have their place on some types of equipment, and are covered briefly at the end of the chapter. Rolling element bearings consist of four basic components: the inner ring, the outer ring, the cage or separator or retainer, and the rolling elements, either balls or rollers. The inner ring is mounted on the shaft with the rolling elements between it and the outer ring, which goes in the housing. Rolling element bearings can be grouped into two basic types: ball bearings and roller bearings. Each type has its advantages and disadvantages which are described below in the discussion for each type of bearing. ~~ ~~ Ball Bearings Ball bearings have a number of advantages over roller bearings, but they also have some disadvantages. Advan- tages are: Low friction Low heat generation Higher speeds Low cost Take both radial and thrust loads Less sensitive to mounting errors Disadvantages are: Lower life Lower load capacity There are many different types of ball bearings, each designed for a particular type of application. The most common type of ball bearing is the Conrad or deep groove type (Figure 1). It is suitable for radial loads, thrust loads in both Figure 1. Deep groove (Conrad) ball bearing. (Courtesy SKF USA, Inc.) directions, or a combination of both. This bearing uses either a two-piece riveted cage or a snap-on polymeric cage. This feature of the bearing tends to limit its top end speed where a one-piece cage is needed, but it is suitable for most industrial machine speeds. Another common type of ball bearing is the angular contact ball bearing (Figure 2). This bearing is designed pri- marily for thrust loads but can take limited radial loads if sufficient thrust loads are also present. The thrust load must be in one direction only on single bearings. This bearing has the advantage of higher capacity and longer life than a deep groove bearing because one of the rings is counterbored, allowing more balls to be assembled in the bearing. Another advantage for very high speeds is that a one piece cage can be used, if necessary. Angular contact bearings are available in several different contact angles, depending on how much thrust will be present relative to the radial load. Figure 2. Angular contact ball bearing. (Courtesy SKF USA, Inc.) Because single angular contact bearings can take thrust in only one direction, they are often used in pairs. This is sometimes called a duplex bearing, or a duplex set (Figure 3). The two single bearings are mounted with their coun- terbores in 'opposite directions, allowing thrust in both directions. Duplex bearings can also be conveniently pre- loaded as a set to provide very rigid and accurate shaft position control and stiffness. 0ack.lo-back arrangement Face-lo-lace arrangement Figure 3. Duplex sets of angular contact ball bearings. (Courtesy SKF USA, Inc.) A variation of the angular contact bearing is the split inner ring bearing (Figure 4). This is a ball bearing with the inner ring split circumferentially, allowing a single row bearing to take thrust in either direction. These bearings are used mostly in the aircraft industry due to their cost. Figure 4. Split inner ring ball bearing. (Courtesy SKF USA, Inc.) Bearings 147 Self-aligning, double-row ball bearings are a somewhat specialized two-row bearing (Figure 5). The outer ring raceway is a portion of a curve with only the inner ring having grooves for the balls to ride in. This allows the bearing to be internally self-aligning, and can com- pensate for considerable mounting or even dynamic misalignment in the shaft/housing system. Its major disadvantage is that because of the flat outer raceway, the load capacity is not very high. Figure 5. Self-aligning ball bearing. (Courtesy SKF USA, Inc.) Finally, thrust-type ball bearings are bearings with a 90" contact angle (Figure 6). They cannot take any radial load, but can take considerable thrust load and high speeds. They are somewhat of a specialty bearing due to the special mounting systems required. -u Figure 6. Thrust ball bearing. (Courtesy SKF USA, Inc.) Roller Bearings Roller bearings are usually used for applications re- quiring greater load carrying capacity than a ball bearing. Roller bearings are generally much stiffer structurally and provide greater fatigue life than do ball bearings of a com- parable size. Their advantages and disadvantages tend to be the opposite of ball bearings. Advantages are: . Greater load capacity . Greater fatigue life 148 Rules of Thumb for Mechanical Engineers Some types take both radial and thrust loads * Some types less sensitive to mounting errors cate the shaft as long as there is no external thrust load. This feature is used in gear trains by using two cylindrical bearings to support the spur gear shaft with no ball bearing. The typical cylindrical roller bearing is free to float axially. It Disadvantages are: Higher friction Higher heat generation Moderate speeds has two roller guiding ribs on one ring and none on the other. Then a ball bearing or other thrust type bearing is used on the other end of the shaft to locate it. Spherical Roller Bearings Higher cost There are three basic types of roller bearings: cylindrical or straight roller bearings, spherical roller bearings, and tapered roller bearings. As with the ball bearings, each has its strengths and weaknesses. Cylindrical Roller Bearings Cylindrical roller bearings have the lowest frictional characteristics of all other roller bearings, which makes them more suitable for high speed operation. They also have the highest radial load carrying capacity. They are not designed for carrying axial loads, although some configurations can handle very small axial loads, such as shaft po- sitioning, when there is no external thrust load. Cylindrical roller bearings are also very sensitive to misalignment. Often, their rollers have a partial or even a full crown to help this situation. Cylindrical roller bearings are available in a variety of rib configurations (Figure 7). These are illustrated below. In general, there must be at least two ribs on one of the rings. One or two ribs on the other ring allow the bearing to lo- Spherical roller bearings (Figure 8) are so named because the cross-section of one of the raceways, usually the outer raceway, makes up a portion of a sphere. The rollers of this type of bearing are barrel shaped and usually symmetrical but sometimes off-center or asymmetrical. The bearings are available in both single- or double-row configurations, but the double-row design is by far the most common. Spher- ical roller bearings are capable of carrying high radial loads or, in the double-row versions, a combination of radial and axial loads. The single-row design cannot take any thrust loading. The great advantage of the spherical roller bearing over the ball bearing or cylindrical roller bearing is its ability to take considerable amounts of misalignment without reduction of capacity. The misalignment can be either static or dynamic, and as much as 3 to 5 degrees depending on the internal geometry of the bearing. It can also take much more thrust load than a ball bearing of the same size. Its biggest disadvantage is that it is the most difficult bearing type to manufacture. It costs several times as much as a Roller Bearing Types Bearing Type Inner Ring Sides F I a n g e s Outer Ring Flanges Sides Both I None I Side Sides Fixed Separable Side Both Sides Sides I Figure 7. Cylindrical roller bearing rib configurations [16]. (Courtesy SKF USA, Inc.) Bearings 149 Bearing on Bearing on Bearing with adapter sleeve withdrawal sleeve cylindrical bore Fylum 8. Spherical roller bearings. (Coutfesy SKF USA, Inc.) cylindrical roller bearing with the same load capacity. The other significant disadvantage is that it has more friction and heat generation than any other type of bearing. Tapered Roller Bearings Tapered roller bearings (Figure 9) are similar to cylindrical roller bearings except that the roller is tapered from one end to the other and the raceways are angled to match the roller taper. Unlike cylindrical roller bearings, they can take large thrust loads or a combination of radial and thrust loads. Tapered roller bearings can be mounted on the shaft in pairs, taking thrust in both directions and completely controlling the shaft location. They also have more load capacity than a spherical roller bearing of the same size and are much less difficult to manufacture, providing a significant cost advantage. The biggest disadvantage of the tapered roller bearing is its tapered design. In operation, the raceway forces push the roller to one end of the bearing so that there must be a guide flange present to keep the roller in the bearing. This sliding contact causes friction and heat generation and makes the bearing generally unsuitable for high speeds. The other disadvantage of this bearing is that it is sensitive to misalignment, just like a cylindrical roller bearing. In general, tapered roller bearings have the same .001” per inch re- quirement for full load capacity. Because the tapered roller bearing has evolved a little dif- ferently than other types of roller bearings, its part termi- nology is different. Inner rings are frequently called cones and outer rings are called cups. Figure 9. Tapered roller bearings. (Courtesy SKF USA, hc.) Standardization Bearings are one of the earlier manufactured items to have become standardized. Today, almost all bearings are made to a strict standard, for many features, that is the same around the world in many aspects, especially in the areas of boundary plan and tolerances. A standardized set of definitions has been developed by the American Bearing Manufacturers Association (ABMA) for the various bearing components and some of their key dimensions and tolerances. To better understand the discussions that follow and to better communicate with bearing suppliers, some of these definitions as given in ANSUAFBMA Standard 1- 1990 [4] are included here. Inner ring: A bearing ring incorporating the raceway(s) on Cone: An inner ring of a tapered roller bearing. its outside surface. Outer ring: A bearing ring incorporating the raceway(s) on Cup: An outer ring of a tapered roller bearing. Cage: A bearing part which partly surrounds all or several of the rolling elements and moves with them. Its pur- pose is to space the rolling elements and generally also to guide and/or retain them in the bearing. its inside surface. Separator: Another word for cage. Retainer: Another word for cage. Rolling element: A ball or roller which rolls between raceways. Raceway: A surface of a load supporting part of a rolling bearing, suitably prepared as a rolling track for the rolling elements. Bearing bore diameter (bore): The bore or I.D. of the inner ring of a rolling bearing. [...]... 16 16 16 18 18 18 20 20 20 22 24 24 28 :ies 1 0 O.D Widtl 26 8 28 8 32 9 10 35 12 42 12 44 12 47 12 52 55 13 58 13 14 62 15 68 75 16 16 80 90 18 95 18 100 18 110 20 115 20 125 22 22 130 140 24 24 145 L50 24 26 L60 28 170 28 180 33 200 210 33 l25 35 0 2 0 D Width 30 9 10 32 1 1 35 12 40 14 47 14 50 15 52 58 16 62 16 17 65 72 17 18 80 85 19 90 20 100 21 22 110 120 23 125 24 63 0 25 140 26 150 28 160 30... 60 65 70 75 80 85 90 95 100 105 110 120 130 140 150 1 9 O.D Width O.D 19 21 24 26 32 34 37 40 42 44 47 52 58 65 72 78 85 90 95 100 11 1) 115 120 125 130 140 150 165 175 190 5 5 5 5 7 7 7 7 7 7 7 7 7 7 9 10 10 10 10 10 13 13 13 13 13 16 16 18 18 20 22 24 28 30 37 39 42 45 47 52 55 62 68 72 EO 85 90 LOO 105 110 120 125 130 140 145 150 165 180 190 110 Width 6 6 7 7 9 9 9 9 9 10 10 12 12 12 13 13 13 16. .. 32 180 34 36 190 38 200 215 40 40 230 250 42 270 45 0 3 0 D Widtl 35 1 1 37 12 42 13 47 14 52 15 16 56 62 17 68 18 72 19 75 20 80 21 90 23 100 25 110 27 120 29 130 31 140 33 150 35 160 37 170 39 180 41 190 43 200 45 215 47 225 49 240 50 260 55 280 58 300 62 320 65 0 4 O.D Widtl 37 42 52 62 72 12 13 15 17 19 80 21 90 23 100 110 120 130 140 150 160 180 190 200 210 225 240 250 260 280 310 340 360 380 25... load for condition n N, = speed for condition n = fraction of time at condition n b = 3 for ball bearings; 10/3 for roller bearings To obtain the most accurate life estimate for a duty cycle, or for a system of bearings, where the life is known for each duty cycle condition or for each bearing, Minor's Rule can be used The formula for this is as follows: where: (Ll& = calculated Ll0 life of condition for. .. (Axial) Load 1 1 All Sizes h6 All Sizes 96 Inner Ring need Heaw I All Sizes 1 1 I j6 I 1 I 96 All Sizes h6 All Sizes Consult Bearing Manufacturer (1) Tolerance Classificationsshown are for solid steel shaff Numerical values are listed in Table 15 For hollow or nonferrous shafts, tighter fits may be needed (2) If greater accuracy is needed, substitute j5, k5 and m5 for j6 k6, and m6 respectively Source:... Temperature Limits for Oils, O F Petroleum o mineral r Supemfined petroleum Synthetic hydrocarbon Synthetic esters Silicones Polyphenolether Perfluorinated compounds 300 350 400 400 500 500 60 0 Table 12 Approximate Temperature Limits for Grease Thickeners, O F Calcium Aluminum Barium Sodium Lithium Synthetics 170 180 225 250 300 . Lubricant Selection 162 Lubricating Methods 163 Relubrication 164 Cleaning, Preservation, and Storage 165 Mounting 166 Shafting 166 Housings 169 Bearing Clearance . 45 47 52 55 62 68 72 EO 85 90 LOO 105 110 120 125 130 140 145 150 165 180 190 110 6 6 7 7 9 9 9 9 9 10 10 12 12 12 13 13 13 16 16 16 18 18 18 20. 16 16 17 17 18 19 20 21 22 23 24 25 26 28 30 32 34 36 38 40 40 42 45 03 0. D. Widtl 35 37 42 47 52 56 62 68 72 75 80 90 100 110 120 130 140 150 160

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