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The geometry factor, which is a fundamental part of the AGMA strength rating of gears, is currently computed using the Lewis parabola which allows computation of the Lewis form factor.(1) The geometry factor is obtained from this Lewis factor and load sharing ratio. This method, which originally required graphical construction methods and more recently has been computerized, works reasonably well for external gears with thick rims.(2-6) However, when thin rims are encountered or when evaluating the strength of internal gears, the AGMA method cannot be used.
Although there is plenty of information and data on the determination of geometry factors and bending strength of external gear teeth, the computation methods regarding internal gear design are less accessible. most of today's designs adopt the formulas for external gears and incorporate some kind of correction factors for internal gears. However, this design method is only an approximation because of the differences between internal gears and external gears. Indeed, the tooth shape of internal gears is different from that of external gears. One has a concave curve, while the other has a convex curve.
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A study of AGMA 218, the draft ISO standard 6336, and BS 436: 1986 methods for rating gear tooth strength and surface durability for metallic spur and helical gears is presented. A comparison of the standards mainly focuses on fundamental formula and influence factors, such as the load distribution factor, geometry factor, and others. No attempt is made to qualify or judge the standards other than to comment on the facilities or lack of them in each standard reviewed. In Part I a comparison of pitting resistance ratings is made, and in the subsequent issue, Part II will deal with bending stress ratings and comparisons of designs.
There are three distinct gear types in angle drives. The most commonly used are bevel and worm drives. Face gear drives are the third alternative.
The paper is not the proof of a discovery, but it is the description of a method: the optimization of the microgeometry for cylindrical gears. The method has been applied and described on some transmissions with helical gears and compound epicyclic, used on different hybrid vehicles. However, the method is also valid for industrial gearboxes.
Gear designs are evolving at an ever accelerating rate, and gear manufacturers need to better understand how the choice of materials and heat treating methods can optimize mechanical properties, balance overall cost and extend service life.
I felt a tap on my shoulder. Turning, I saw the chief draftsman who said, "You're in charge of gears." And he walked away. Dumbfounded, I stared at the back of his head, and sat down at my drafting board. It was November, 1963, shortly after JFK was assassinated, and after I was discharged from the U.S. Army.
A reader asks: We are currently revising our gear standards and tolerances and a few questions with the new standard AGMA 2002-C16 have risen. Firstly, the way to calculate the tooth thickness tolerance seems to need a "manufacturing profile shift coefficient" that isn't specified in the standard; neither is another standard referred to for this coefficient. This tolerance on tooth thickness is needed later to calculate the span width as well as the pin diameter. Furthermore, there seems to be no tolerancing on the major and minor diameters of a gear.
A reader asks: I'd like to know about the different approaches and factors considered while determining the value of Ka in regards to the DIN 3990 and AGMA standards.
Accurate prediction of gear dynamic factors (also known as Kv factors) is necessary to be able to predict the fatigue life of gears. Standards-based calculations of gear dynamic factors have some limitations. In this paper we use a multibody dynamic model, with all 6 degrees of freedom (DOF) of a high-speed gearbox to calculate gear dynamic factors. The findings from this paper will help engineers to understand numerous factors that influence the prediction of dynamic factors and will help them to design more reliable gears.
I have a query (regarding) calculated gear life values. I would like to understand for what % of gear failures the calculated life is valid? Is it 1-in-100 (1% failure, 99% reliability) or 1-in-one-thousand (0.1% failure)?
A discussion of ISO and AGMA standards for gears, shafts and bearings, and the art of designing a gearbox that meets your requirements.
Three experts tackle the question of profile shift in this issue's edition of "Ask the Expert."
One of the best ways to learn the ISO 6336 gear rating system is to recalculate the capacity of a few existing designs and to compare the ISO 6336 calculated capacity to your experience with those designs and to other rating methods. For these articles, I'll assume that you have a copy of ISO 6336, you have chosen a design for which you have manufacturing drawings and an existing gear capacity calculation according to AGMA 2001 or another method. I'll also assume that you have converted dimensions, loads, etc. into the SI system of measurement.
Chicago- Results of recent studies on residual stress in gear hobbing, hobbing without lubricants and heat treating were reported by representatives of INFAC (Instrumented Factory for Gears) at an industry briefing in March of this year.
The use of dimensionless factors to describe gear tooth geometry seems to have a strong appeal to gear engineers. The stress factors I and J, for instance, are well established in AGMA literature. The use of the rack shift coefficient "x" to describe nonstandard gear proportions is common in Europe, but is not as commonly used in the United States. When it is encountered in the European literature or in the operating manuals for imported machine tools, it can be a source of confusion to the American engineer.
Questions: I have heard the terms "safety factor," "service factor," and "application factor" used in discussing gear design. what are these factors an dhow do they differ from one another? Why are they important?
This paper shows a method to calculate the occurring tooth root stress for involute, external gears with any form of fillets very precisely within a few seconds.
In terms of the tooth thickness, should we use the formulation with respect to normal or transverse coordinate system? When normalizing this thickness in order to normalize the backlash (backlash parameter), we should divide by the circular pitch. Thus, when normalizing, should this circular pitch be defined in the normal or traverse coordinate system, depending on which formulation has been used? Is the backlash parameter always defined with respect to the tangential plane or normal plane for helical gears?
This article presents some of the findings of cutting investigations at WZL in which the correlation of cutting parameters, cutting materials, tool geometry and tool life have been determined.
I have outsourced gear macrogeometry due to lack of resources. Now I received the output from them and one of the gears is with β0.8Γ module correction factor for m = 1.8 mm gear. Since bending root stress and specific slide is at par with specification, but negative correction factor β0.8Γ module β is quite high β how will it influence NVH behavior/transmission error? SAP and TIF are very close to 0.05 mm; how will that influence the manufacturing/cost?
Traditionally, gear rating procedures consider manufacturing accuracy in the application of the dynamic factor, but only indirectly through the load distribution are such errors in the calculation of stresses used in the durability and gear strength equations. This paper discusses how accuracy affects the calculation of stresses and then uses both statistical design of experiments and Monte Carlo simulation techniques to quantify the effects of different manufacturing and assembly errors on root and contact stresses.
The geometry of the bevel gear is quite complicated to describe mathematically, and much of the overall surface topology of the tooth flank is dependent on the machine settings and cutting method employed. AGMA 929-A06 β Calculation of Bevel Gear Top Land and Guidance on Cutter Edge Radius β lays out a practical approach for predicting the approximate top-land thicknesses at certain points of interest β regardless of the exact machine settings that will generate the tooth form. The points of interest that AGMA 929-A06 address consist of toe, mean, heel, and point of involute lengthwise curvature. The following method expands upon the concepts described in AGMA 929-A06 to allow the user to calculate not only the top-land thickness, but the more general case as well, i.e. β normal tooth thickness anywhere along the face and profile of the bevel gear tooth. This method does not rely on any additional machine settings; only basic geometry of the cutter, blank, and teeth are required to calculate fairly accurate tooth thicknesses. The tooth thicknesses are then transformed into a point cloud describing both the convex and concave flanks in a global, Cartesian coordinate system. These points can be utilized in any modern computer-aided design software package to assist in the generation of a 3D solid model; all pertinent tooth macrogeometry can be closely simulated using this technique. A case study will be presented evaluating the accuracy of the point cloud data compared to a physical part.
The focus of the following presentation is two-fold: 1) on tests of new geometric variants; and 2) on to-date, non-investigated operating (environmental) conditions. By variation of non-investigated eometric parameters and operation conditions the understanding of micropitting formation is improved. Thereby it is essential to ensure existent calculation methods and match them to results of the comparison between large gearbox tests and standard gearbox test runs to allow a safe forecast of wear due to micropitting in the future.
I must admit that after thumbing through the pages of this relatively compact volume (113 pages, 8.5 x 11 format), I read its three chapters(theory of gearing, geometry and technology, and biographical history) from rear to front. It will become obvious later in this discussion why I encourage most gear engineers to adopt this same reading sequence!
In 1961 I presented a paper, "Calculating Conjugate Helical Forms," at the semi-annual meeting of the American Gear Manufacturers Association (AGMA). Since that time, thousands of hobs, shaper cutters and other meshing parts have been designed on the basis of the equations presented in that paper. This article presents the math of that paper without the formality of its development and goes on to discuss its practical application.
Optimization is applied to the design of a spiral bevel gear reduction for maximum life at a given size. A modified feasible directions search algorithm permits a wide variety of inequality constraints and exact design requirements to be met with low sensitivity to initial values. Gear tooth bending strength and minimum contact ration under load are included in the active constraints. The optimal design of the spiral bevel gear reduction includes the selection of bearing and shaft proportions in addition to gear mesh parameters. System life is maximized subject to a fixed back-cone distance of the spiral bevel gear set for a specified speed ratio, shaft angle, input torque and power. Significant parameters in the design are the spiral angle, the pressure angle, the numbers of teeth on the pinion and gear and the location and size of the four support bearings. Interpolated polynomials expand the discrete bearing properties and proportions into continuous variables for gradient optimization. After finding the continuous optimum, a designer can analyze near-optimal designs for comparison and selection. Design examples show the influence of the bearing lives on the gear parameters in the optimal configurations. For a fixed back-cone distance, optimal designs with larger shaft angles have larger service lives.
In the past gear manufacturers have had to rely on hob manufacturers' inspection of individual elements of a hob, such as lead, involute, spacing, and runout. These did not always guarantee correct gears, as contained elements may cause a hob to produce gears beyond tolerance limits.
There are different types of spiral bevel gears, based on the methods of generation of gear-tooth surfaces. A few notable ones are the Gleason's gearing, the Klingelnberg's Palloid System, and the Klingelnberg's and Oerlikon's Cyclo Palliod System. The design of each type of spiral bevel gear depends on the method of generation used. It is based on specified and detailed directions which have been worked out by the mentioned companies. However, there are some general aspects, such as the concepts of pitch cones, generating gear, and conditions of force transmissions that are common for all types of spiral bevel gears.
A universal gear is one generated by a common rack on a cylindrical, conical, or planar surface, and whose teeth can be oriented parallel or skewed, centered, or offset, with respect to its axes. Mating gear axes can be parallel or crossed, non-intersecting or intersecting, skewed or parallel, and can have any angular orientation (See Fig.1) The taper gear is a universal gear. It provides unique geometric properties and a range of applications unmatched by any other motion transmission element. (See Fig.2) The taper gear can be produced by any rack-type tool generator or hobbing machine which has a means of tilting the cutter or work axis and/or coordinating simultaneous traverse and infeed motions.
An accurate and fast calculation method is developed to determine the value of a trigonometric function if the value of another trigonometric function is given. Some examples of conversion procedures for well-known functions in gear geometry are presented, with data for accuracy and computing time. For the development of such procedures the complete text of a computer program is included.
The paper describes a procedure for the design of internal gear pairs, which is a generalized form of the long and short addendum system. The procedure includes checks for interference, tip interference, undercutting, tip interference during cutting, and rubbing during cutting.
This article investigates fillet features consequent to tooth grinding by generating methods. Fillets resulting from tooth cutting and tooth grinding at different pressure angles and with different positions of grinding wheel are compared. Ways to improve the final fillet of the ground teeth with regard to tooth strength and noise, as well as the grinding conditions, are shown. "Undergrinding" is defined and special designs for noiseless gears are described.
Beginning with our June Issue, Gear Technology is pleased to present a series of full-length chapters excerpted from Dr. Hermann J. Stadtfeldβs latest scholarly β yet practical β contribution to the gear industry β Gleason Bevel Gear Technology. Released in March, 2014 the book boasts 365 figures intended to add graphic support of a better understanding and easier recollection of the covered material.