POWER TRANSMITTING CAPACITY OF SPUR GEARS 3001 The analytical method is valid for any gear design and is recommended for the following conditions: 1) ratio of net face width to pinion pitch diameter, F/D, is equal or greater than 2.0 (for double helical gears the gap is not included in the face width); 2) applications with over- hung gear elements; 3) applications with long shafts subject to large deflections or where deflections under load reduce width of contact; and 4) applications where contact does not extend across the full face of narrowest member when loaded. For designs that have high crowns to centralize tooth contact under deflected conditions, the factors C m and K m may be conservatively approximated by this method. For the most commonly encountered condition, contact across the entire face width under normal oper- ating load, the face load distribution factor expressions are (24) (25) If the total contact length under normal operating load is less than the face width, the expressions for the load distribution factor are (26) (27) where G = tooth stiffness constant, lb/in./in. of face (MPa), the average mesh stiffness of a single pair of teeth in the normal direction. The usual range of this value that is compati- ble with this analysis is 1.5–2.0 × 10 6 lb/in. 2 (1.0–1.4 × 10 4 MPa). The most conservative value is the highest. e t = total lead mismatch between mating teeth, in loaded condition, in. (mm). Z = length of action in transverse plane, from Equation (17) on page 2995, in. (mm). P b = transverse base pitch, in. (mm). The total mismatch, e t , is a virtual separation between the tooth profiles at the end of the face width which is composed of the static, no load separation plus a component due to the elastic load deformations. This total mismatch is influenced by all the items listed under Load Distribution Factors except the Hertzian contact stress and bending deformations of the gear teeth, which are accounted for by the tooth stiffness constant G. Evaluation of e t , is difficult but it is critical to the reliability of the analytical method. An iterative computer program may be used, but in critical applications full scale testing may be desirable. Allowable Stress Numbers, S ac and S at .—The allowable stress numbers depend on 1) material composition and cleanliness; 2) mechanical properties; 3) residual stress; 4) hardness and; and 5) type of heat treatment, surface or through hardened. An allowable stress number for unity application factor, 10 million cycles of load appli- cation, 99 per cent reliability and unidirectional loading, is determined or estimated from laboratory and field experience for each material and condition of that material. This stress number is designated S ac and S at .The allowable stress numbers are adjusted for design life cycles by the use of life factors. The allowable stress numbers for gear materials vary with material composition, cleanli- ness, quality, heat treatment, and processing practices. For materials other than steel, a range is shown, and the lower values should be used for general design purposes. Data for materials other than steel are given in the Standard. for spur gearing: C mf 1.0 Ge t F×× 2 W t × += and, for helical gearing: C mf 1.0 Ge t ZF××× 1.8 W t × += for spur gearing: C mf 2.0 Ge t F××× W t = and, for helical gearing: C mf 2.0 Ge t ZF×××× W t P b × = Machinery's Handbook 27th Edition Copyright 2004, Industrial Press, Inc., New York, NY 3002 POWER TRANSMITTING CAPACITY OF SPUR GEARS Allowable stress numbers for steel gears are established by specific quality control requirements for each material type and grade. All requirements for the quality grade must be met in order to use the stress values for that grade. Details of these quality requirements are given in the Standard. Reverse Loading: For idler gears and other gears where the teeth are completely reverse loaded on every cycle, 70 per cent of the S at values should be used. Fig. 4. Effective Case Depth for Carburized Gears, h e Case Depth of Surface–Hardened Gears.—The Standard provides formulas to guide the selection of minimum effective case depth at the pitchline for carburized and induction hardened teeth based on the maximum shear from contact loading. Fig. 5. Minimum Total Case Depth for Nitrided Gears, h c Machinery's Handbook 27th Edition Copyright 2004, Industrial Press, Inc., New York, NY POWER TRANSMITTING CAPACITY OF SPUR GEARS 3003 Fig. 5 shows values that have a long history of successful use for carburized gears and can be used for such gears. For nitrided gears, case depth is specified as the total case depth h c , which is defined as the depth below the surface at which the hardness has dropped to110 per cent of the core hardness. Minimum total case depths for nitrided gears are shown in Fig. 5. Momentary Overloads.—When the gear is subjected to less than 100 cycles of momen- tary overloads, the maximum allowable stress is determined by the allowable yield proper- ties rather than the bending fatigue strength of the material. Fig. 6 shows suggested values of the allowable yield strength S ay for through hardened steel. For case hardened gears, the core hardness should be used in conjunction with the table of metallurgical factors affect- ing the bending stress number for carburized gears shown in the Standard. Fig. 6. Allowable Yield Strength Number for Steel Gears, S ay The design should be checked to make sure that the teeth are not permanently deformed. Also, when yield is the governing stress, the stress correction factor K f is considered inef- fective and therefore taken as unity. Yield Strength.—For through hardened gears up to 400 BHN, a yield strength factor K y can be applied to the allowable yield strength taken from Fig. 6. This factor is applied at the maximum peak load to which the gear is subjected: (28) (28a) For conservative practice, K y is taken as 0.5 and for industrial practice, K y is 0.75. Hardness Ratio Factor C H .—The hardness ratio factor depends on (1) gear ratio and (2) Brinell hardness numbers of gear and pinion. When the pinion is substantially harder than the gear, the work hardening effect increases the gear capacity. Typical values of the hard- S ay K y × W max K a × K v P d F K s K m × JK f × ≥ S ay K y × W max K a K s K m ××× K v FmJK f ×××× ≥ Machinery's Handbook 27th Edition Copyright 2004, Industrial Press, Inc., New York, NY 3004 POWER TRANSMITTING CAPACITY OF SPUR GEARS ness ratio factor, C H , for through hardened gears are shown in Fig. 7. These values apply to the gear only, not to the pinion. When surface hardened pinions (48 HRC or harder) are run with through hardened gears (180–400 BHN), a work hardening effect is achieved. The C H factor varies with the surface finish of the pinion, K p and the mating gear hardness as shown in Fig. 8. Fig. 7. Hardness Ratio Factor, C H (Through Hardened) Life Factors C L and K L .— These life factors adjust the allowable stress numbers for the required number of cycles of operation. In the Standard, the number of cycles, N, is defined as the number of mesh contacts under load of the gear tooth being analyzed. Allowable stress numbers are established for 10,000,000 tooth load cycles at 99 per cent reliability. The life cycle factors adjust the allowable stress numbers for design lives other than 10,000,000 cycles. The life factor accounts for the S/N characteristics of the gear material as well as for the gradually increased tooth stress that may occur from tooth wear, resulting in increased dynamic effects and from shifting load distributions that may occur during the design life of the gearing. A C L or K L value of 1.0 may be used beyond 10,000,000 cycles, where jus- tified by experience. Life Factors for Steel Gears: Insufficient data exist to provide accurate life curves for every gear and gear application. However, experience suggests life curves for pitting and strength of steel gears are as shown in Figs. 9 and 10. These figures do not include data for nitrided gears. The upper portions of the shaded zones are for general commercial applica- tions. The lower portions of the shaded zones are typically used for critical service applica- tions where little pitting and tooth wear are permissible and where smoothness of operation and low vibration levels are required. When gear service ratings are established by the use Machinery's Handbook 27th Edition Copyright 2004, Industrial Press, Inc., New York, NY WORM GEARING 3007 Worm Gearing Worm Gearing Classification.—Worm gearing may be divided into two general classes, fine-pitch worm gearing, and coarse-pitch worm gearing. Fine-pitch worm gearing is seg- regated from coarse-pitch worm gearing for the following reasons: 1) Fine-pitch worms and wormgears are used largely to transmit motion rather than power. Tooth strength except at the coarser end of the fine-pitch range is seldom an impor- tant factor; durability and accuracy, as they affect the transmission of uniform angular motion, are of greater importance. 2) Housing constructions and lubricating methods are, in general, quite different for fine- pitch worm gearing. 3) Because fine-pitch worms and wormgears are so small, profile deviations and tooth bearings cannot be measured with the same accuracy as can those of coarse pitches. 4) Equipment generally available for cutting fine-pitch wormgears has restrictions which limit the diameter, the lead range, the degree of accuracy attainable, and the kind of tooth bearing obtainable. 5) Special consideration must be given to top lands in fine-pitch hardened worms and wormgear-cutting tools. 6) Interchangeability and high production are important factors in fine-pitch worm gear- ing; individual matching of the worm to the gear, as often practiced with coarse-pitch pre- cision worms, is impractical in the case of fine-pitch worm drives. American Standard Design for Fine-pitch Worm Gearing (ANSI B6.9-1977).—This standard is intended as a design procedure for fine-pitch worms and wormgears having axes at right angles. It covers cylindrical worms with helical threads, and wormgears hobbed for fully conjugate tooth surfaces. It does not cover helical gears used as wormgears. Hobs: The hob for producing the gear is a duplicate of the mating worm with regard to tooth profile, number of threads, and lead. The hob differs from the worm principally in that the outside diameter of the hob is larger to allow for resharpening and to provide bot- tom clearance in the wormgear. Pitches: Eight standard axial pitches have been established to provide adequate coverage of the pitch range normally required: 0.030, 0.040, 0.050, 0.065, 0.080, 0.100, 0.130, and 0.160 inch. Axial pitch is used as a basis for this design standard because: 1) Axial pitch establishes lead which is a basic dimension in the production and inspection of worms; 2) the axial pitch of the worm is equal to the circular pitch of the gear in the central plane; and 3) only one set of change gears or one master lead cam is required for a given lead, regardless of lead angle, on commonly-used worm-producing equipment. Lead Angles: Fifteen standard lead angles have been established to provide adequate coverage: 0.5, 1, 1.5, 2, 3, 4, 5, 7, 9, 11, 14, 17, 21, 25, and 30 degrees. This series of lead angles has been standardized to: 1) Minimize tooling; 2) permit obtaining geometric similarity between worms of different axial pitch by keeping the same lead angle; and 3) take into account the production distribution found in fine-pitch worm gearing applications. For example, most fine-pitch worms have either one or two threads. This requires smaller increments at the low end of the lead angle series. For the less frequently used thread num- bers, proportionately greater increments at the high end of the lead angle series are suffi- cient. Machinery's Handbook 27th Edition Copyright 2004, Industrial Press, Inc., New York, NY 3008 WORM GEARING Table 1. Formulas for Proportions of American Standard Fine-pitch Worms and Wormgears ANSI B6.9-1977 All dimensions in inches unless otherwise indicated. Pressure Angle of Worm: A pressure angle of 20 degrees has been selected as standard for cutters and grinding wheels used to produce worms within the scope of this Standard because it avoids objectionable undercutting regardless of lead angle. Although the pressure angle of the cutter or grinding wheel used to produce the worm is 20 degrees, the normal pressure angle produced in the worm will actually be slightly greater, and will vary with the worm diameter, lead angle, and diameter of cutter or grind- ing wheel. A method for calculating the pressure angle change is given under the heading Effect of Production Method on Worm Profile and Pressure Angle. LETTER SYMBOLS P=Circular pitch of wormgear P=axial pitch of the worm, P x , in the central plane P x =Axial pitch of worm P n =Normal circular pitch of worm and wormgear = P x cos λ = P cos ψ λ =Lead angle of worm ψ =Helix angle of wormgear n=Number of threads in worm N=Number of teeth in wormgear N=nm G m G =Ratio of gearing = N ÷ n Item Formula Item Formula WORM DIMENSIONS WORMGEAR DIMENSIONS a a Current practice for fine-pitch worm gearing does not require the use of throated blanks. This results in the much simpler blank shown in the diagram which is quite similar to that for a spur or heli- cal gear. The slight loss in contact resulting from the use of non-throated blanks has little effect on the load-carrying capacity of fine-pitch worm gears. It is sometimes desirable to use topping hobs for pro- ducing wormgears in which the size relation between the outside and pitch diameters must be closely controlled. In such cases the blank is made slightly larger than D o by an amount (usually from 0.010 to 0.020) depending on the pitch. Topped wormgears will appear to have a small throat which is the result of the hobbing operation. For all intents and purposes, the throating is negligible and a blank so made is not to be considered as being a throated blank. Lead Pitch Diameter D = NP ÷ π = NP x ÷ π Pitch Diameter Outside Diameter D o = 2C − d + 2a Outside Diameter Face Width Safe Minimum Length of Threaded Portion of Worm b b This formula allows a sufficient length for fine-pitch worms. DIMENSIONS FOR BOTH WORM AND WORMGEAR Addendum a = 0.3183P n Tooth thickness t n = 0.5P n Whole Depth h t = 0.7003P n + 0.002 Approximate normal pressure angle c c As stated in the text on page 3008, the actual pressure angle will be slightly greater due to the man- ufacturing process. φ n = 20 degrees Working Depth h k = 0.6366P n Clearance c = h t − h k Center distance C = 0.5 (d + D) lnP x = dlπλtan()÷= d o d 2a+= F Gmin 1.125 d o 2c+() 2 d o 4a–() 2 – ×= F W D o 2 D 2 –= Machinery's Handbook 27th Edition Copyright 2004, Industrial Press, Inc., New York, NY WORM GEARING 3009 Pitch Diameter Range of Worms: The minimum recommended worm pitch diameter is 0.250 inch and the maximum is 2.000 inches.Pitch diameters for all possible combinations of lead and lead angle, together with the number of threads for each lead, are given in Table 2a and 2b. Tooth Form of Worm and Wormgear: The shape of the worm thread in the normal plane is defined as that which is produced by a symmetrical double-conical cutter or grinding wheel having straight elements and an included angle of 40 degrees. Because worms and wormgears are closely related to their method of manufacture, it is impossible to specify clearly the tooth form of the wormgear without referring to the mat- ing worm. For this reason, worm specifications should include the method of manufacture and the diameter of cutter or grinding wheel used. Similarly, for determining the shape of the generating tool, information about the method of producing the worm threads must be given to the manufacturer if the tools are to be designed correctly. The worm profile will be a curve that departs from a straight line by varying amounts, depending on the worm diameter, lead angle, and the cutter or grinding wheel diameter. A method for calculating this deviation is given in the Standard. The tooth form of the wormgear is understood to be made fully conjugate to the mating worm thread. Proportions of Fine-pitch Worms and Wormgears.—Hardened worms and cutting tools for wormgears should have adequate top lands. To automatically provide sufficient top lands, regardless of lead angle or axial pitch, the addendum and whole depth propor- tions of fine-pitch worm gearing are based on the normal circular pitch. Tooth proportions based on normal pitch for all combinations of standard axial pitches and lead angles are given in Table 3. Formulas for the proportions of worms and worm gears are given in Table 1. Example 1:Determine the design of a worm and wormgear for a center distance of approximately 3 inches if the ratio is to be 10 to 1; axial pitch, 0.1600 inch; and lead angle, 30 degrees. From Table 2a and 2b it can be determined that there are eight possible worm diameters that will satisfy the given conditions of lead angle and pitch. These worms have from 3 to 10 threads. To satisfy the 3-inch center distance requirement it is now necessary to determine which of these eight worms, together with its mating wormgear, will come closest to making up this center distance. One way of doing this is as follows: First use the formula given below to obtain the approximate number of threads neces- sary. Then from the eight possible worms in Table 2a and 2b, choose the one whose num- ber of threads is nearest this approximate value: Approximate number of threads needed for required center distance = Approximate number of threads = Of the eight possible worms in Table 2a and 2b, the one having a number of threads near- est this value is the 10-thread worm with a pitch diameter of 0.8821 inch. Since the ratio of gearing is given as 10, N may now be computed as follows: N = 10 × 10 = 100 teeth (from Table 1) Other worm and wormgear dimensions may now be calculated using the formulas given in Table or may be taken from the data presented in Table 2a, 2b, and 3. l = 1.600 inches (from Table 2b) 2π required center distance× P x λcot m G +() 2 3.1416 3×× 0.1600 1.7320 10+()× 10.04 threads= Machinery's Handbook 27th Edition Copyright 2004, Industrial Press, Inc., New York, NY 3010 WORM GEARING d = 0.8821 inch (from Table 2b) D = 100 × 0.1600 + 3.1416 = 5.0930 inches (from Table 1) C = 0.5(0.8821 + 5.0930) = 2.9876 inches (from Table 1) P n = 0.1386 inch (from Table 3) a =0.0441 inch (from Table 3) h t = 0.0990 inch (from Table 3) h k = 0.6366 × 0.1386 = 0.0882 inch (from Table 1) c = 0.0990−0.0882 = 0.0108 inch (from Table 1) t n = 0.5 × 0.1386 = 0.0693 inch (from Table 1) d 0 = 0.8821 + (2 × 0.0441) = 0.9703 inch (from Table 1) D 0 = (2 × 2.9876) − 0.8821 + (2 × 0.0441) = 5.1813 (from Table 1) Example 2:Determine the design of a worm and wormgear for a center distance of approximately 0.550 inch if the ratio is to be 50 to 1 and the axial pitch is to be 0.050 inch. Assume that n = 1 (since most fine-pitch worms have either one or two threads). The lead of the worm will then be nP x = 1 × 0.050 = 0.050 inch. From Table 2a and 2b it can be deter- mined that there are six possible lead angles and corresponding worm diameters that will satisfy this lead. The approximate lead angle required to meet the conditions of the exam- ple can be computed from the following formula: Using letter symbols, this formula becomes: Of the six possible worms in Table 2a and 2b, the one with the 3-degree lead angle is clos- est to the calculated 2°59′ lead angle. This worm, which has a pitch diameter of 0.3037 inch, is therefore selected. The remaining worm and wormgear dimensions may now be determined from the data in Table 2a, 2b and 3 and by computation using the formulas given in Table 1. N = 50×1=50 teeth (from Table 1) d = 0.3037 inch (from Table 2b) D = 50×0.050 ÷ 3.1416 = 0.7958 inch (from Table 1) C = 0.5(0.3037 + 0.7958) = 0.5498 inch (from Table 1) P n = 0.0499 inch (from Table 3) a =0.0159 inch (from Table 3) h t = 0.0370 inch (from Table 3) h k = 0.6366 × 0.0499 = 0.0318 inch (from Table 1) c = 0.0370−0.0318 = 0.0052 inch (from Table 1) t n = 0.5 × 0.0499 = 0.0250 inch (from Table 1) d 0 = 0.3037 + (2 × 0.0159) = 0.3355 inch (from Table 1) D 0 = (2 × 0.5498) − 0.3037 + (2 × 0.0159) = 5.1813 (from Table 1) F G 1.125 0.9703 2 0.0108×+() 2 0.9703 4 0.0441×–() 2 – 0.6689 inch== F W 5.1813 2 5.0930 2 – 0.9525 inch== Cotangent of approx. lead angle 2π approximate center distance required× assumed number of threads axial pitch× m G –= F Gmin 1.125 0.3355 2 0.0052×+() 2 0.3355 4 0.0159×–() 2 – 0.2405 inch== F W 0.8277 2 0.7958 2 – 0.2276 inch== Machinery's Handbook 27th Edition Copyright 2004, Industrial Press, Inc., New York, NY WORM GEARING 3015 (4) (5) (6) (7) (8) In these formulas, ρ ni = radius of curvature of normal thread profile for involute thread; r=pitch radius of worm; Φ n =normal pressure angle of cutter or grinding wheel; λ =lead angle of worm; ρ n =radius of curvature of normal thread profile; R=radius of cutter or grinding wheel; ∆Φ =difference between the normal pressure angle of the thread and the normal pressure angle of the cutter or grinding wheel in minutes (see diagram). Sub- scripts c and w are used to denote the cutter and grinding wheel diameters, respectively; n=number of threads in worm; a=addendum of worm; q=slant height of worm addendum; y=amount normal worm profile departs from a straight side (see diagram). Sub- scripts c and w are used to denote the cutter and grinding wheel diameters, respectively; s=effect along slant height of worm thread caused by change in pressure angle ∆Φ ∆y=difference in y values of two cutters or grinding wheels of different diameter (see diagram); ∆s=effect of ∆Φ c − ∆Φ w along slant height of thread (see diagram). Example 3:Assuming the worm dimensions are the same as in Example 1, determine the corrections for two worms, one milled by a 2-inch diameter cutter, the other ground by a 20-inch diameter wheel, both to be assembled with identical wormgears. To make identical worms when using a 2-inch cutter and a 20-inch wheel, the pressure angle of either the cutter or the wheel must be corrected by an amount corresponding to ∆s and the profile of the cutter or wheel must be a curve which departs from a straight line by an amount ∆y. The calculations are as follows: For the 2-inch diameter cutter, using Formula (1) to (6), (1) (2) (3) (4) qaφ n sec inches= y q 2 2ρ n inches= s 0.000582q∆φ inches= ∆yy w y c – inches= ∆ss c s w inches–= ρ ni 0.4410 0.3420× 0.5000 2 0.6033 inch== ρ n 0.6033 0.4410 0.6033× 1 0.8660 2 × + 0.9581 inch== ∆φ c 5400 0.4410 0.5000 3 ×× 10 1 0.8660 2 × 0.4410+() 24.99 inches== q 0.0441 1.0642× 0.0469 inch== Machinery's Handbook 27th Edition Copyright 2004, Industrial Press, Inc., New York, NY [...]... 0.01 0. 02 0.03 0.04 0.05 0.06 0.07 0.08 0. 09 0.1 Lead angle of worm in degrees 5 Deg 10 Deg 15 Deg 20 Deg 25 Deg 30 Deg 35 Deg 40 Deg 45 Deg 89. 7 94 .5 96 .1 97 .0 97 .4 97 .7 97 .9 98.0 98 .0 81.3 89. 5 92 . 6 94 .1 95 .0 95 .5 95 .9 96.0 96 .1 74.3 85.0 89 .2 91 .4 92 . 6 93 .4 93 .9 94.1 94 .2 68.4 80 .9 86.1 88.8 90 .4 91 .4 91 .9 92 . 2 92 . 3 63.4 77 .2 83.1 86.3 88 .2 89. 4 90 .1 90 .4 90 .5 59. 0 73.8 80.4 84.0 86.1 87.4 88 .2 88.6... 1.53 1. 72 4-Inch Center Distance 9. 40 16 .90 18.70 6. 09 11.50 12. 70 4 . 29 8.15 9. 06 3 .28 6 .26 6 .96 2. 65 5.05 5. 62 2 .22 4 .24 4.71 1.67 3. 19 3.55 1.34 2. 56 2. 85 1. 12 2.14 2. 38 1150 1750 5.48 3. 62 2.57 1 .97 1. 59 1.33 1.00 0.81 6.87 4.54 3 .22 2. 48 2. 00 1.68 1 .26 1.01 9. 06 6.16 4.38 3.37 2. 72 2 .28 1. 72 1.38 1.15 11.30 7. 72 5. 49 4 .22 3.41 2. 86 2. 15 1.73 1.44 15 .90 10.70 7. 62 5.86 4.73 3 .96 2. 99 2. 40 2. 00 19. 70... 15.00 25 .30 28 .00 12. 20 20 .50 22 .60 10 .20 17 .20 19. 00 7.66 12. 90 14.30 6.15 10.40 11.50 5.13 8.66 9. 58 4.41 7.43 8 .23 8-Inch Center Distance 59. 60 95 .20 104.00 40 .90 67.40 74.40 29 .00 48 .20 53 .20 22 .20 37.00 40 .90 18.00 29 .90 33.10 15.10 25 .10 27 .80 11.30 18 .90 20 .90 9. 10 15 .20 16.80 7. 59 12. 70 14.00 6. 52 10 .90 12. 00 1150 1750 39. 60 27 .20 19. 30 14 .90 12. 00 10.10 7.58 6. 09 5.08 4.36 47.40 33.30 23 .80... 178.00 73 .20 117.00 1 29 .00 52. 10 83.70 91 .90 40.00 64.40 70.70 32. 30 52. 10 57 .20 27 .10 43.70 48.00 20 .40 32. 90 36.10 16.40 26 .40 29 .00 13.70 22 .10 24 .20 11.70 18 .90 20 .80 12- Inch Center Distance 167.00 25 7.00 27 1.00 118.00 186.00 20 2.00 83 .90 133.00 145.00 64.40 1 02. 00 111.00 52. 10 82. 70 90 .10 43.70 69. 40 75.70 32. 90 52. 20 57.00 26 .40 41 .90 45.80 22 .10 35.00 38 .20 18 .90 30.10 32. 80 1150 1750 194 .00 144.00... 3.05 0 .91 1. 89 2. 16 0.70 1.44 1.65 0.56 1.16 1.33 0.47 0 .98 1. 12 0.35 0.73 0.84 0 .28 0. 59 0.68 3-Inch Center Distance 3.60 6 .99 7. 79 2. 31 4.65 5 .26 1. 62 3 . 29 3.74 1 .24 2. 52 2.87 1.00 2. 04 2. 31 0.84 1.71 1 .94 0.63 1 .28 1.46 0.51 1.03 1.17 0. 42 0.86 0 .98 3.5-Inch Center Distance 6.60 12. 30 13.70 4 .24 8 .27 9 .21 2. 99 5.85 6.54 2. 28 4. 49 5.03 1.84 3. 62 4.06 1.54 3.04 3.40 1.16 2. 29 2. 56 0 .93 1.83 2. 06 0.78... 14.80 12. 40 9. 34 7.50 6 .26 5.38 57.50 38.80 27 .60 21 .20 17 .20 14.40 10.80 8.70 7 .26 6 .23 66.80 46.40 33 .20 25 .60 20 .70 17.40 13.10 10.50 8. 79 7.55 83.60 58.70 41 .90 32. 20 26 .00 21 .80 16.40 13 .20 11.00 9. 46 96 .40 68.40 49 .20 37 .90 30.70 25 .80 19. 40 15.60 13.00 11 .20 116.00 85.10 61 .20 47.00 38.00 31 .90 24 .00 19. 30 16.10 13.80 134.00 98 .70 71.00 54.70 44.30 37 .20 28 .00 22 .50 18.80 16.10 Copyright 20 04,... 11.80 21 .10 23 .40 8.36 15.00 16.70 6.40 11.60 12. 80 5.16 9. 34 10.30 4. 32 7.83 8.67 3 .25 5 .90 6.53 2. 61 4.73 5 .24 2. 18 3 .95 4.37 1.87 3. 39 3.76 6-Inch Center Distance 27 .30 45.30 50.10 17.80 30.40 33.70 12. 60 21 .60 23 .90 9. 64 16.60 18.40 7.78 13.40 14 .90 6. 52 11.30 12. 50 4 .90 8.47 9. 38 3 .93 6.80 7.53 3 .28 5.68 6 . 29 2. 82 4.87 5.40 7-Inch Center Distance 41.60 67.30 73 .90 27 .70 46 .20 51 .20 19. 60 32. 90 36.50... 70:1 7 .21 4.63 3 .25 2. 49 2. 01 1.68 1 .27 1. 02 0.85 0.73 5:1 10:1 15:1 20 :1 25 :1 30:1 40:1 50:1 60:1 70:1 11.10 7.08 4 .98 3.81 3.07 2. 57 1 .94 1.55 1.30 1.11 5:1 10:1 15:1 20 :1 25 :1 30:1 40:1 50:1 60:1 70:1 17.50 11 .20 7.88 6.03 4.86 4.07 3.06 2. 46 2. 05 1.76 5:1 10:1 15:1 20 :1 25 :1 30:1 40:1 50:1 60:1 70:1 25 .90 16.70 11.80 9. 00 7 .26 6.08 4.58 3.67 3.07 2. 63 300 720 870 5-Inch Center Distance 18 .20 31.00... 10:1 15:1 20 :1 25 :1 30:1 40:1 50:1 0.78 0. 49 0.35 0 .27 0 .21 0.18 0.13 0.11 5:1 10:1 15:1 20 :1 25 :1 30:1 40:1 50:1 60:1 1.38 0.88 0. 62 0.47 0.38 0. 32 0 .24 0. 19 0.16 5:1 10:1 15:1 20 :1 25 :1 30:1 40:1 50:1 60:1 2. 55 1.63 1.14 0.88 0.71 0. 59 0.44 0.36 0.30 5:1 10:1 15:1 20 :1 25 :1 30:1 40:1 50:1 60:1 3.66 2. 35 1.65 1 .26 1. 02 0.85 0.64 0.51 0.43 300 720 870 2. 5-Inch Center Distance 2. 04 4.13 4.68 1.30 2. 67 3.05... 55 .2 70.7 77.8 81.7 84.1 85.6 86.5 86 .9 86 .9 51 .9 67.8 75.4 79. 6 82. 2 83.8 84.7 85 .2 85 .2 48 .9 65 .2 73.1 77.5 80.3 82. 0 83.0 83.5 83.5 46.3 62. 7 70 .9 75.6 78.5 80.3 81.4 81 .9 81.8 Table 6 AGMA Input Mechanical Horsepower Ratings of Cone-Drive Worm Gearinga Worm Speed, RPM Ratio 100 5:1 10:1 15:1 20 :1 25 :1 30:1 40:1 50:1 0.40 0 .25 0.18 0.13 0.11 0. 09 0.07 0.05 300 720 2- Inch Center Distance 1.04 2. 18 . Deg. 20 Deg. 25 Deg. 30 Deg. 35 Deg. 40 Deg. 45 Deg. 0.01 89. 7 94 .5 96 .1 97 .0 97 .4 97 .7 97 .9 98.0 98 .0 0. 02 81.3 89. 5 92 . 6 94 .1 95 .0 95 .5 95 .9 96.0 96 .1 0.03 74.3 85.0 89 .2 91 .4 92 . 6 93 .4 93 .9 94.1. 7.88 19. 60 32. 90 36.50 41 .90 49 .20 20 :1 6.03 15.00 25 .30 28 .00 32. 20 37 .90 25 :1 4.86 12. 20 20 .50 22 .60 26 .00 30.70 30:1 4.07 10 .20 17 .20 19. 00 21 .80 25 .80 40:1 3.06 7.66 12. 90 14.30 16.40 19. 40 50:1. 85.10 98 .70 15:1 11.80 29 .00 48 .20 53 .20 61 .20 71.00 20 :1 9. 00 22 .20 37.00 40 .90 47.00 54.70 25 :1 7 .26 18.00 29 .90 33.10 38.00 44.30 30:1 6.08 15.10 25 .10 27 .80 31 .90 37 .20 40:1 4.58 11.30 18 .90 20 .90