chapter 1 evolution of a successful design

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JOURNAL BEARINGS 19.6 BEARINGS AND LUBRICATION tial bearings are used in situations where the load is mainly unidirectional. Partial journal bearings have been found to reduce frictional torque on the journal and provide convenient accessibility, and they do not, in many instances, require strict manufacturing tolerance. Partial journal bearings in which the bearing radius exceeds the journal radius are called clearance bearings, whereas partial journal bearings in which the bearing and the journal radii are equal are termed fitted bearings. Geometries in which two circular sectors are employed are called elliptical, or lemon, bearings (Fig. 19.3c).These bearings are really not elliptical at all but are fabricated by uniting two halves of a circular bearing which have had their mating faces machined so that the bearing has an approximately elliptical appearance. Lemon bearings are probably the most widely used bearings at low and moderate speeds. They are extensively used in turbine applications. Elliptical bearings in which the two cylindrical halves are laterally displaced along the major axis are termed offset bearings (Fig. 19.3d). The relative displacement of the center of each half of the bearing is called the preset. When the upper half of the bearing is displaced horizontally in the direction of rotation, the bearing has negative preset. It is found that load capacity increases with preset. Offset bearings have relatively high horizontal stiffness, which helps prevent dynamic instability. Further, offset bearings allow greater lubricant flow and so run cooler. Novel offset journal bearing designs for reducing power loss and wear in duty cycles which combine nonreversing loading with limited journal angular oscillation or in steady operation with counterrotation of journal and bearing under a constant load have been studied. In these applications, conventional journal bearings are found to develop extremely thin lubricant films, which in turn results in high friction and wear. Figure 19.3e depicts a journal bearing in which both the journal and the bearing are divided axially into segments with offset centerlines. This arrangement produces a dynamic rocking motion which promotes a thicker lubricating film. Accordingly the assembly has been called a rocking journal bearing. When a step is milled from the surface of the bearing (Fig. 19.3f), the resulting bearing is called a pressure dam, or step, bearing. The purpose of the step is to create additional hydrodynamic pressure on the top of the journal as the lubricant is rotated into the step. In turn, this pressure buildup enhances the load on the journal and therefore diminishes its susceptibility to vibration problems. Pressure dam bearings are very popular in the petrochemical industry. Bearing geometries consisting of three or more sectors (Fig. 19.3g and h) are termed lobed, or multilobed, bearings. Generally, bearings with more than three lobes are used only in gas bearing applications. Multilobed bearings act as a number of partial bearings in series.The cost of multilobed bearings is considered moderate. The multileaf journal bearing (Fig. 19.3i) is a variant of a multilobed bearing. It consists of a number of identical circular arcs, or leaves, whose centers are equally spaced around the generating circle. The operating characteristics of a multileaf bearing are practically independent of the direction of loading for bearings with eight or more leaves. In a floating-ring journal bearing (Fig. 19.3j), the lubricating film is divided in two by the addition of a “floating” ring between the journal and the bearing. Floatingring bearings have lower frictional losses and reduced heat generation and provide better stability. Hydrodynamic journal bearings may be distinguished by whether the bearing surface can pivot. The basic advantage of pivoting, or tilting-pad, journal bearings (Fig. 19.3k) over fixed-pad journal bearings is that they can accommodate, with little loss in performance, any shaft deflection or misalignment. Downloaded from Digital Engineering Library @ McGraw-Hill (www.digitalengineeringlibrary.com) Copyright © 2004 The McGraw-Hill Companies. All rights reserved. Any use is subject to the Terms of Use as given at the website.
A foil journal bearing (Fig. 19.3l) consists of a very thin compliant bearing surface resting atop a series of corrugations.When it is compared to a conventional gas bearing, the foil bearing has a thicker film, higher load capacity, lower power loss, better stability, and superior endurance to high operating temperatures. 19.2.2 Journal Shapes JOURNAL BEARINGS 19.7 Although the journal is generally assumed to be perfectly circular, wear effects or poor manufacture can lead to journals with the shapes shown in Fig. 19.4a, b, and c. In addition, the possibility of developing pressure by grooving the surface of the journal has been investigated.Three grooved patterns that were found to yield good stability characteristics are shown in Fig. 19.4c, d, and e. (a) (b) (d) (e) FIGURE 19.4 Journal shapes. (a) Hourglass; (b) barrel; (c) tapered; (d) herringbone; (e) partly grooved symmetrical pattern: (f) partly grooved asymmetrical pattern. (Parts (d), (e), and (f) are from [19.1].) 19.3 BEARING MATERIALS AND SELECTION CRITERIA 19.3.1 Bearing Materials JOURNAL BEARINGS The ideal journal bearing material would have the following characteristics: 1. High compressive strength to withstand the applied radial loading 2. High fatigue strength to endure any cyclic changes in load direction and/or load intensity 3. Compatibility with the journal material to minimize surface scoring and bearing seizure whenever the journal and bearing surfaces come into contact (e.g., during startup) 4. Embedability to permit foreign particles in the lubricant to penetrate the bearing surface to avoid scoring and wear Downloaded from Digital Engineering Library @ McGraw-Hill (www.digitalengineeringlibrary.com) Copyright © 2004 The McGraw-Hill Companies. All rights reserved. Any use is subject to the Terms of Use as given at the website. (c) (f)
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Source: STANDARD HANDBOOK OF MACHIN
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EVOLUTION OF A SUCCESSFUL DESIGN EV
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Flywheels are an important machine
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EVOLUTION OF A SUCCESSFUL DESIGN sp
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GLOSSARY OF SYMBOLS CHAPTER 2 A THE
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(a) A THESAURUS OF MECHANISMS 2.5 (
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(c) A THESAURUS OF MECHANISMS A THE
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(g) (l) (c) (e) A THESAURUS OF MECH
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(a) (g) (c) A THESAURUS OF MECHANIS
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(a) (c) (e) A THESAURUS OF MECHANIS
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(e) (a) (i) A THESAURUS OF MECHANIS
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(b) A THESAURUS OF MECHANISMS (d) A
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(a) (f) A THESAURUS OF MECHANISMS A
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(a) A THESAURUS OF MECHANISMS A THE
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(c) (a) A THESAURUS OF MECHANISMS A
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(c) (f) (i) A THESAURUS OF MECHANIS
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A THESAURUS OF MECHANISMS (a) A THE
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(a) (d) (j) (g) A THESAURUS OF MECH
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CHAPTER 3 LINKAGES Richard E. Gusta
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(a) (c) where l = number of links (
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(a) (c) LINKAGES LINKAGES 3.5 FIGUR
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Since both sides of (3.5) can be di
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The driven link d will be at angle
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LINKAGES FIGURE 3.8 Two positions o
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LINKAGES LINKAGES 3.13 FIGURE 3.10
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LINKAGES FIGURE 3.12 Determining th
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of motion. Proper care in selection
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4. With O B as vertex, set off a li
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3. Determine lines S ab(ab) and S a
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LINKAGES LINKAGES 3.23 3.13 John J.
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CHAPTER 4 CAM MECHANISMS Andrzej A.
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FIGURE 4.3 Plate cams with oscillat
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FIGURE 4.6 Types of follower motion
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CAM MECHANISMS CAM MECHANISMS 4.7 F
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CAM MECHANISMS CAM MECHANISMS 4.9 F
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the motion specification for the ri
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CAM MECHANISMS CAM MECHANISMS 4.13
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CAM MECHANISMS 4.15 This is called
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It is a common rule of thumb to ass
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For F23 = 0, roller and cam lose th
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FIGURE 4.20 Programming of cam syst
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n5 = N2N4 n2 (5.5) N3N5 and the t
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5.3 PLANETARY GEAR TRAINS GEAR TRAI
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GEAR TRAINS 5.7 termini of the velo
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GEAR TRAINS TABLE 5.1 Solution by T
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GEAR TRAINS 5.11 TABLE 5.2 Characte
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FIGURE 5.9 (a) Planetary train; (b)
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GEAR TRAINS GEAR TRAINS 5.15 (a) (b
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SPRINGS SPRINGS 6.5 close-wound: wo
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6.3.2 Heat Treatment of Springs Hea
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SPRINGS 6.9 Downloaded from Digital
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SPRINGS 6.11 Downloaded from Digita
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Helical compression springs are str
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Types of Ends. Four basic types of
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TABLE 6.5 Typical Properties of Spr
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SPRINGS SPRINGS 6.19 The rate equat
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SPRINGS SPRINGS 6.21 FIGURE 6.9 End
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The fatigue life estimates in Table
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The spring rate for a rectangular-w
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SPRINGS 6.27 FIGURE 6.15 Stress cor
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6.5 HELICAL EXTENSION SPRINGS 6.5.1
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6.5.2 Initial Tension Initial tensi
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SPRINGS FIGURE 6.20 Common end conf
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Dynamic considerations discussed pr
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SPRINGS SPRINGS 6.37 TABLE 6.14 Tol
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The number of coils is equal to the
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SPRINGS SPRINGS 6.41 TABLE 6.17 Com
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6.7.1 Nomenclature a OD/2, mm (in)
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Ef Sc = C1 h − 2 2 (1 −µ )Ma
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SPRINGS SPRINGS 6.47 FIGURE 6.31 Lo
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Because of production variations in
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7. h = 1.41t = 1.41(0.054) = 0.076
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and Design equations are SPRINGS SP
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6.9 FLAT SPRINGS 6.9.1 Introduction
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SPRINGS fEb P = ot (6.52) where K =
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SPRINGS SPRINGS 6.59 TABLE 6.25 Max
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If unknown, let b/t = 100/1, D2 = 1
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6.11.2 Design Equations: Rectangula
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SPRINGS SPRINGS 6.65 FIGURE 6.50 Ty
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SPRINGS SPRINGS 6.67 FIGURE 6.51 Av
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used only for centerless ground all
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CHAPTER 7 FLYWHEELS Daniel M. Curti
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The energy-storage capacity of a fl
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FLYWHEELS Since the torque-angle cu
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7.2.3 Coefficient of Energy Variati
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7.2.5 Speed-Dependent Torques FLYWH
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FLYWHEELS FLYWHEELS 7.11 range [7.7
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and 205 rpm = 2π(205)/60 = 21.468
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Example 6. An alloy of density 26.6
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cos (π/8) f4 = =0.340 (7.42) 7.11
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For a constant-thickness disk witho
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factor F s = 1.0 for the fully stre
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FLYWHEELS FLYWHEELS 7.23 (g) (h) (i
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many composite flywheels shred, for
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CLUTCHES AND BRAKES CLUTCHES AND BR
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shaft begins to rotate.At the desig
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FIGURE 8.7 Classification of brakes
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8.1.4 Selecting a Brake CLUTCHES AN
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TABLE 8.2 Selecting the Right Brake
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Similarly, In general, CLUTCHES AND
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IPIL(ΩP −ΩL) tS = (8.10) T(I
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The rate at which heat is generated
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Here ωo = 2π 60 (315) = 33 rad/
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for preliminary design estimates on
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CLUTCHES AND BRAKES FIGURE 8.16 Equ
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drum’s rotation, use the positive
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pmax Px = (θ2 −θ1 + sin θ1 cos
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8.5.2 Centrifugal Clutches CLUTCHES
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CLUTCHES AND BRAKES FIGURE 8.19 For
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3. Next the moment arm length for
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CLUTCHES AND BRAKES 8.41 to the dis
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CLUTCHES AND BRAKES Torque Capacity
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For finding the average contact pre
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CLUTCHES AND BRAKES (a) (b) (c) FIG
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CHAPTER 9 SPUR GEARS Joseph E. Shig
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SPUR GEARS SPUR GEARS 9.5 FIGURE 9.
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9.3 FORCE ANALYSIS In Fig. 9.3 a ge
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SPUR GEARS SPUR GEARS 9.9 TABLE 9.5
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Allowable Bending Stress Number. Th
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10.1 INTRODUCTION / 10.1 10.2 TYPES
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HELICAL GEARS HELICAL GEARS 10.3 (a
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helical gears are employed, a limit
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contact ratio and the axial overlap
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10.5.1 Strength and Durability HELI
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HELICAL GEARS FIGURE 10.4 Dynamic f
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HELICAL GEARS HELICAL GEARS 10.13 (
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HELICAL GEARS housing which support
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If the tooth contact pattern at nor
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for external gears; for internal ge
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HELICAL GEARS HELICAL GEARS 10.21 G
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HELICAL GEARS HELICAL GEARS 10.23 F
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Letting Ze = distance on line of ac
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HELICAL GEARS HELICAL GEARS 10.39 H
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HELICAL GEARS 10.41 Downloaded from
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HELICAL GEARS HELICAL GEARS 10.51 p
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HELICAL GEARS HELICAL GEARS 10.53 W
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HELICAL GEARS HELICAL GEARS 10.55 E
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In most cases the blank temperature
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CHAPTER 11 BEVEL AND HYPOID GEARS T
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FIGURE 11.3 Zerol bevel set. (Gleas
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FIGURE 11.7 Hypoid gear nomenclatur
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The taper you select depends in som
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BEVEL AND HYPOID GEARS BEVEL AND HY
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Hypoid gears are recommended where
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TABLE 11.1 Material Factors C M BEV
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11.4.7 Hypoid Offset BEVEL AND HYPO
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FIGURE 11.16 Selection of spiral an
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should be hardened and tempered abo
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BEVEL AND HYPOID GEARS BEVEL AND HY
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TABLE 11.7 Mean Addendum Factor BEV
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11.5.4 AGMA References † The foll
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BEVEL AND HYPOID GEARS TABLE 11.10
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BEVEL AND HYPOID GEARS TABLE 11.10
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BEVEL AND HYPOID GEARS area.The rec
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TABLE 11.12 Load-Distribution Facto
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BEVEL AND HYPOID GEARS FIGURE 11.21
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BEVEL AND HYPOID GEARS 11.43 Downlo
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TABLE 11.15 Allowable Bending Stres
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BEVEL AND HYPOID GEARS 11.7.3 Axial
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Pressure lubrication is recommended
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CHAPTER 12 WORM GEARING K. S. Edwar
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FIGURE 12.2 Photograph of a worm-ge
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12.3 VELOCITY AND FRICTION Figure 1
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where the subscripts are t for the
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This force is the negative x direct
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24, from Table 12.3, K m = 0.823 by
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12.7 DESIGN STANDARDS The American
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TABLE 12.6 Dimensions of the Gear a
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Table 12.9 presents a system for st
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12.8.3 Gear Ratio The gear ratio is
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12.8.10 Worm Length The effective l
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CHAPTER 13 POWER SCREWS Rudolph J.
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FIGURE 13.1 Power screw assembly us
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POWER SCREWS POWER SCREWS 13.5 TABL
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POWER SCREWS POWER SCREWS 13.7 The
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which can be solved for a circular
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TABLE 13.3 Sizes and Capacities of
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REFERENCES POWER SCREWS POWER SCREW
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CHAPTER 14 BELT DRIVES Wolfram Funk
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Disadvantages: ● Limited power tr
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FIGURE 14.2 Multiple-pulley drives.
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Because µβwπ Fu = F2′(m − 1)
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The maximum power transmission capa
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(a) (c) BELT DRIVES BELT DRIVES 14.
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BELT DRIVES FIGURE 14.11 Pretension
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BELT DRIVES 14.17 Taking into consi
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FIGURE 14.15 Multiple-ply belt. BEL
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4. The manufacturer can supply belt
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BELT DRIVES FIGURE 14.18 Section of
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The calculation of belt velocity (p
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BELT DRIVES The flexibility of the
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BELT DRIVES BELT DRIVES 14.29 FIGUR
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transmissible power P is determined
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BELT DRIVES BELT DRIVES 14.33 The a
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BELT DRIVES BELT DRIVES 14.35 (a) (
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(twisting of tension member and its
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The power transmission capacity of
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FIGURE 14.33 Comparison of system c
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CHAPTER 15 CHAIN DRIVES John L. Wri
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esponding chains in Ref. [15.1], bu
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TABLE 15.1 Roller-Chain Dimensions
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CHAIN DRIVES 15.3 SELECTION OF ROLL
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sprockets may be required. Idler sp
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CHAIN DRIVES TABLE 15.2 Service Fac
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Determine Lubrication Type. The typ
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CHAIN DRIVES CHAIN DRIVES 15.15 TAB
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CHAIN DRIVES inexpensive, but they
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TABLE 15.9 Offset Sidebar Chain Dim
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CHAIN DRIVES CHAIN DRIVES 15.21 Cha
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CHAIN DRIVES CHAIN DRIVES 15.23 TAB
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15.6.3 Horsepower Ratings of Offset
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15.7.3 Silent-Chain Sprockets CHAIN
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In a drive where the shaft centers
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CHAIN DRIVES silent-chain size when
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7. Misalignment Amount and type of
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COUPLINGS COUPLINGS 16.5 TABLE 16.1
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alignment up to 45°. The maximum a
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FIGURE 16.4 Portion of shaft showin
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COUPLINGS 16.3.2 Chain, Grid, and B
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COUPLINGS 16.13 FIGURE 16.11 (a) Si
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COUPLINGS 16.15 FIGURE 16.16 Hydrau
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operating radius corresponding to a
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FIGURE 16.24 A bellows coupling. (S
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COUPLINGS COUPLINGS 16.21 pression
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FIGURE 16.30 Bonded type of a shear
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16.5 UNIVERSAL JOINTS AND ROTATING-
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COUPLINGS COUPLINGS 16.27 FIGURE 16
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COUPLINGS The correction factor Ks
Page 519 and 520: COUPLINGS COUPLINGS 16.31 Rzeppa Un
Page 521 and 522: method of attachment allows a much
Page 523 and 524: CHAPTER 17 SHAFTS Charles R. Mischk
Page 525 and 526: Geometric fidelity is important to
Page 527 and 528: Solution. Equation (17.1) is used.
Page 529 and 530: SHAFTS Based on this, the designer
Page 531 and 532: SHAFTS 17.9 The dual-entry slope dy
Page 533 and 534: SHAFTS SHAFTS 17.11 The prediction
Page 535 and 536: 17.4 DISTORTION DUE TO TORSION Angu
Page 537 and 538: A press fit induces a surface press
Page 539 and 540: factor corrected for notch sensitiv
Page 541 and 542: The methods of Secs. 17.2 and 17.3
Page 543 and 544: REFERENCES 17.1 Joseph E. Shigley a
Page 545 and 546: Source: STANDARD HANDBOOK OF MACHIN
Page 547 and 548: Source: STANDARD HANDBOOK OF MACHIN
Page 549 and 550: ROLLING-CONTACT BEARINGS 18.5 FIGUR
Page 551 and 552: FIGURE 18.9 Tapered-roller thrust b
Page 553 and 554: ROLLING-CONTACT BEARINGS where C s
Page 555 and 556: ROLLING-CONTACT BEARINGS For exampl
Page 557 and 558: ROLLING-CONTACT BEARINGS ROLLING-CO
Page 559 and 560: age on the bearing per revolution a
Page 561 and 562: ROLLING-CONTACT BEARINGS TABLE 18.5
Page 563 and 564: deep-groove ball bearings. Self-ali
Page 565 and 566: CHAPTER 19 JOURNAL BEARINGS Theo G.
Page 567 and 568: Λ Bearing number = (6µω/p a)(R/C
Page 569: TABLE 19.1 Characteristics of Lubri
Page 573 and 574: TABLE 19.2 Physical Properties of J
Page 575 and 576: TABLE 19.4 Performance Ratings from
Page 577 and 578: Next the candidate materials are gi
Page 579 and 580: This equation can be interpreted as
Page 581 and 582: JOURNAL BEARINGS where θ 1 and θ
Page 583 and 584: In general, volume flow rate per un
Page 585 and 586: JOURNAL BEARINGS JOURNAL BEARINGS 1
Page 587 and 588: It was proposed that the film press
Page 597 and 598: L b 1 Sb2 + b3 (L/D) TABLE 19.15 S
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