《ISO-6336-1-2006.pdf》由会员分享，可在线阅读，更多相关《ISO-6336-1-2006.pdf（118页珍藏版）》请在三一文库上搜索。
1、 Reference number ISO 6336-1:2006(E) ISO 2006 INTERNATIONAL STANDARD ISO 6336-1 Second edition 2006-09-01 Calculation of load capacity of spur and helical gears Part 1: Basic principles, introduction and general influence factors Calcul de la capacit de charge des engrenages cylindriques dentures dr
2、oite et hlicodale Partie 1: Principes de base, introduction et facteurs gnraux dinfluence ISO 6336-1:2006(E) PDF disclaimer This PDF file may contain embedded typefaces. In accordance with Adobes licensing policy, this file may be printed or viewed but shall not be edited unless the typefaces which
3、are embedded are licensed to and installed on the computer performing the editing. In downloading this file, parties accept therein the responsibility of not infringing Adobes licensing policy. The ISO Central Secretariat accepts no liability in this area. Adobe is a trademark of Adobe Systems Incor
4、porated. Details of the software products used to create this PDF file can be found in the General Info relative to the file; the PDF-creation parameters were optimized for printing. Every care has been taken to ensure that the file is suitable for use by ISO member bodies. In the unlikely event tha
5、t a problem relating to it is found, please inform the Central Secretariat at the address given below. ISO 2006 All rights reserved. Unless otherwise specified, no part of this publication may be reproduced or utilized in any form or by any means, electronic or mechanical, including photocopying and
6、 microfilm, without permission in writing from either ISO at the address below or ISOs member body in the country of the requester. ISO copyright office Case postale 56 CH-1211 Geneva 20 Tel. + 41 22 749 01 11 Fax + 41 22 749 09 47 E-mail copyrightiso.org Web www.iso.org Published in Switzerland ii
7、ISO 2006 All rights reserved ISO 6336-1:2006(E) ISO 2006 All rights reserved iii Contents Page Foreword vi Introduction vii 1 Scope . 1 2 Normative references. 2 3 Terms, definitions, symbols and abbreviated terms. 2 4 Basic principles 12 4.1 Application 12 4.1.1 Scuffing 12 4.1.2 Wear . 12 4.1.3 Mi
8、cropitting . 12 4.1.4 Plastic yielding 12 4.1.5 Particular categories 12 4.1.6 Specific applications 12 4.1.7 Safety factors 13 4.1.8 Testing . 15 4.1.9 Manufacturing tolerances 15 4.1.10 Implied accuracy. 15 4.1.11 Other considerations 15 4.1.12 Influence factors. 16 4.1.13 Numerical equations. 18
9、4.1.14 Succession of factors in course of calculation. 18 4.1.15 Determination of allowable values of gear deviations 18 4.2 Tangential load, torque and power. 18 4.2.1 Nominal tangential load, nominal torque and nominal power. 18 4.2.2 Equivalent tangential load, equivalent torque and equivalent po
10、wer. 19 4.2.3 Maximum tangential load, maximum torque and maximum power. 19 5 Application factor KA 19 5.1 Method A Factor KA-A. 20 5.2 Method B Factor KA-B. 20 6 Internal dynamic factor Kv. 20 6.1 Parameters affecting internal dynamic load and calculations. 20 6.1.1 Design 20 6.1.2 Manufacturing. 2
11、1 6.1.3 Transmission perturbance. 21 6.1.4 Dynamic response 21 6.1.5 Resonance. 22 6.2 Principles and assumptions 22 6.3 Methods for determination of dynamic factor . 22 6.3.1 Method A Factor Kv-A. 22 6.3.2 Method B Factor Kv-B. 23 6.3.3 Method C Factor Kv-C. 23 6.4 Determination of dynamic factor u
12、sing Method B: Kv-B. 24 6.4.1 Running speed ranges. 24 6.4.2 Determination of resonance running speed (main resonance) of a gear pair 3) 25 6.4.3 Dynamic factor in subcritical range (N u NS). 27 6.4.4 Dynamic factor in main resonance range (NS 1) The same limitations on gear accuracy grade as in b)
13、apply to gears operating in this speed range. Resonance peaks can occur at N = 2, 3 . in this range. However, in the majority of cases vibration amplitudes are small, since excitation loads with lower frequencies than meshing frequency are usually small. 2) When it is known in advance that gears wil
14、l operate in the supercritical sector, there is no need to evaluate the resonance speed. As a consequence, the dynamic factor can be directly determined in accordance with 6.4.5. 3) For a definition of N, see Equation (9). In practice, the calculated resonance sector is broadened to ensure a safe ma
15、rgin. See Equations (10) and (11) and the preamble thereto. ISO 6336-1:2006(E) ISO 2006 All rights reserved 25 For some gears in this speed range, it is also necessary to consider dynamic loads due to transverse vibration of the gear and shaft assemblies (see 6.3.2). If the critical frequency is nea
16、r the frequency of rotation, the associated effective value of Kv can exceed the value calculated using Equation (21) by up to 100 %. This condition should be avoided. 6.4.2 Determination of resonance running speed (main resonance) of a gear pair 3) This is as follows: E1 1red 30 000 c n zm = (6) wh
17、ere mred is the relative mass of a gear pair, i.e. of the mass per unit face width of each gear, referred to its base radius or to the line of action * 1212 red *2*2 121b22b1 m mJJ m mmJrJr = + (7) where * 1,2* 1,2 2 b1,2 J m r = (8) and rb is the base radius See 6.4.8 for the method of calculating
18、an approximate value of mred. See Clause 9 for the stiffness c. Method A is to be preferred for less common transmission designs. A method for deriving approximate values is specified for the following cases: a) pinion on large diameter shaft; b) two neighbouring gears rigidly joined together; c) on
19、e big wheel driven by two pinions; d) simple planetary gears; e) idler gears. The ratio of pinion speed to resonance speed is termed the “resonance ratio”, N (n1 is in revolutions per minute): red111 E1 30 000 mnnz N nc = (9) The resonance running speed nE1 may be above or below the running speed ca
20、lculated from Equation (6) because of stiffnesses which have not been included (the stiffness of shafts, bearings, housings, etc.) and as a result of damping. For reasons of safety, the resonance ratio in the main resonance range, NS, is defined by the following upper limit. S 1,15NN 2 Cv1 0,32 0,32
21、 Cv2 0,34 0,57 0,3 Cv3 0,23 0,096 1,56 Cv4 0,90 0,57 0,05 1,44 Cv5 0,47 0,47 Cv6 0,47 0,12 1,74 1 2,5 Cv7 0,75 0,125 sin ( 2) + 0,875 1,0 2 Hlim ay 1 18,45 1,5 1897 C =+ NOTE When the material of the pinion (1) is different from that of the wheel (2), Cay1 and Cay2 are calculated separately, then Ca
22、y = 0,5 (Cay1 + Cay2). ISO 6336-1:2006(E) 28 ISO 2006 All rights reserved Key X contact ratio, Y factor, Cv NOTE For the equations used for calculation, see Table 4. Figure 2 Values of Cv1 to Cv7 for determination of Kv-B (Method B) ISO 6336-1:2006(E) ISO 2006 All rights reserved 29 Key X allowable
23、stress number, Hlim, N/mm2 Y tip relief, Cay, m NOTE When the pinion material (1) is different from the wheel (2) then Cay = 0,5 (Cay1 + Cay2). Figure 3 Tip relief Cay produced by running-in (see Table 3 for calculation) Bp, Bf and Bk are non-dimensional parameters to take into account the effect of
24、 tooth deviations and profile modifications on the dynamic load 4). pb eff p At (/ ) c f B KF b = (15) f eff f At ( / ) c f B KFb = (16) a k At 1 ( / ) c C B KFb = (17) The effective single pitch and profile deviations are those of the “run-in” pinion and wheel. Initial deviations are generally modi
25、fied during early service (running-in). The values of fpb eff and ff eff are determined by deducting estimated running-in allowances (yp and yf) as follows: pb effpbp ffy= (18) effff ffy = (19) 4) Equation (17) is not suitable for the determination of an “optimum” tip relief Ca. The amount Ca of tip
26、 relief may only be used in Equation (17) for gears of quality grades in the range 0 to 5 as specified in ISO 1328-1. For gears in the range 6 to 12, Bk = 1,0. Also see 4.1.1.2. ISO 6336-1:2006(E) 30 ISO 2006 All rights reserved Considerations of probability suggest that, in general, magnitudes of t
27、ransmission deviation will not be greater than the allowable values of fpb and ff for the wheel, which are the larger. They are therefore used in Equations (18) and (19) respectively; these are usually the values for the largest wheel. In the event that neither experimental nor service data on relev
28、ant material running-in characteristics are available (Method A), it can be assumed that yp = y, with y from Figure 17 or 18 or 8.3.5.1. yf can be determined in the same way as y when the profile deviation ff is used instead of base pitch deviation fpb. Ca is the design amount for profile modificati
29、on (tip relief at the beginning and end of tooth engagement). A value Cay resulting from running-in is to be substituted for Ca in Equation (17) in the case of gears without a specified profile modification. The value of Cay can be obtained from Figure 3 or calculated according to Table 4. See Claus
30、e 9 for single tooth stiffness c. 6.4.4 Dynamic factor in main resonance range (NS 30 6). Kv can be read from graphs (see 6.5.1) or computed (see 6.5.2). The method gives similar values. 6.5.1 Graphical values of dynamic factor using Method C vF350 () 1KfKN=+ (34) fF takes into account the influence
31、 of the load on the dynamic factor, K350, the influence of the gear accuracy grade at the specific loading of 350 N/mm and N is the resonance ratio see Equation (9). The curves for gear accuracy grade in Figures 5 and 6 extend only to the value 22 1 ( / 100) / (1 )v zuu+ = 3 m/s, which is not genera
32、lly exceeded for this accuracy grade. a) For helical gears with overlap ratio W 1 (also approximately for 0,9), the correction factor fF shall be in accordance with Table 5 and (K350 N) shall be in accordance with Figure 5. b) For spur gears, the correction factor fF shall be in accordance with Tabl
33、e 6 and (K350 N) shall be in accordance with Figure 6. c) For helical gears with overlap ratio 0,9) 2 11 v23 2 t A 1 100 1 Kv zu KKK F u K b = + + (36) where numerical values for K1 and K2 shall be as specified in Table 7, and K3 shall be in accordance with Equation (37). If (Ft KA) /b is less than
34、100 N/mm, this value is assumed to be equal to 100 N/mm. See 6.4.1 a). 3 3 2 1 2 22 11 22 If 0,2 2,0 1001 If 0,2 0,357 2,071 1001001 1 v z u K u v zv z uu K uu = + = + + u (37) b) For helical gears with overlap ratio maxcalm max cal 2 FbF W b bb b = Figure 7 Distribution of load along face width wit
35、h linear equivalent misalignment (illustration of principle) Figure 8 Calculation of load per the unit face width (F/b)max with linear distribution of the load on the face width Fm = (Ft KA Kv) -,-,- ISO 6336-1:2006(E) 44 ISO 2006 All rights reserved 7.4.3.1 Gear mesh The stiffness of each increment
36、 results from the method of calculating the deformations. The stiffness of each spring is the mean value of mesh stiffness c according to Clause 9. The load is assumed to be in the zone of single tooth contact without load sharing. The load sharing between helical teeth is not considered. In certain
37、 gears such as thin rimmed gears, the stiffness can vary. Similarily, at ends of the face width the stiffness values can be less than in the centre face. These effects are ignored in Method B. 7.4.3.2 Gear body The deformations of the gear body due to bending and torsion can be considered by regardi
38、ng the gear body as a part of the shaft. Different diameters are used for calculating the bending and torsional deformation in the area of the gear mesh, which should be between the root and the tip diameter of the pinion/gear. The value for bending is the (tip diameter root diameter)/2 plus root di
39、ameter. For torsion it is the root diameter plus 0,4 modules. The load is in the base tangent plane for bending. 7.4.3.3 Shaft / hub connection For normally shrink-fitted gears or connections, the shaft is stiffened to a diameter midway (dmid) between hub diameter and bore (dbore), see Figure 9. Fig
40、ure 9 Definitions of various diameters 7.4.3.4 Pinion and gear shaft The bending deflections of the pinion and gear shaft (with variable inside and outside diameter) have to be calculated according to the linear bending theory. The bending deflections can be caused by all gear meshes and by all othe
41、r external loads (belts, chains, couplings, etc.). The diameter in the tooth area to be used for bending is (tip diameter root diameter)/2 plus root diameter. The load is in the base tangent plane for bending. The torsional deformation of the pinion and gear have only to be calculated in the area of
42、 the gear mesh. It has to be considered that the torque is decreasing along the face width. A diameter of root diameter plus 0,4 modules should be used. 7.4.3.5 Bearings The elastic deformations of the bearings may be calculated by the input of stiffness values for the applied load. If exact stiffne
43、ss values are not known, minimum and maximum values have to be chosen to verify the influence of the bearing deformations. -,-,- ISO 6336-1:2006(E) ISO 2006 All rights reserved 45 7.4.3.6 Housing The elastic deformations of the housing may be determined by calculation or empirical means. In the case
44、 of tapered bearings, the axial deflection of the housing due to gear loads and external thrust loads must be considered in determing the bearing clearances and resulting shaft positions at the bearings. 7.4.3.7 Foundation The elastic deformations of the foundation may be determined by calculation o
45、r empirical means. 7.4.4 Static displacements 7.4.4.1 Shaft working position in bearings The operating bearing clearances must be considered, including the effects of manufacturing variations, thermal expansion, interference fits, axial clearance in tapered bearings and oil film thickness in plain b
46、earings. 7.4.4.2 Manufacturing errors Manufacturing variations (permissible variations in gears, housings, etc.) may be estimated from drawing tolerances or established manufacturing standards. The use of ISO 1328 for tooth alignment fH is permitted for the estimate of total manufacturing variation,
47、 provided that contact checking at assembly is used to verify that its use is appropriate. 7.4.5 Assumptions The methods used for determining the values of bearing deformations, bearing clearances, housing deformation and the values for manufacturing errors must be stated. If influence factors are n
48、eglected, it has to be justified that they are low enough in magnitude. 7.4.6 Computer program output In order to verify the computer calculations, the output of the program must include the full list of input values and all relevant intermediate results. To understand the input assumptions and the output value of KH, the following data is required in tabular graphic form: deflections of the shafts (bending and twist); bearing forces; gear data; load distribution; load distribution factor. 7.5 Determination of face load factor using Method C: KH-C The formulae for the calcul