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BGA Training Notes TN1-1: Introduction to Gear Noise

© British Gear Association 2001 http://www.bga.org.uk

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BGA Training Notes TN1-1

Introduction to Gear Noise

Author: J.D.Smith M.A., Ph.D., C.Eng., M.I.Mech.E. Cambridge University Engineering Department Trumpington Street, Cambridge, CB2 1PZ Email: [email protected] http://www.eng.cam.ac.uk/



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Index

INTRODUCTION TO GEAR NOISE.................................................................................................................3

1.0 CAUSES OF GEAR NOISE .....................................................................................................................3

2.0 THEORETICAL PREDICTION PROBLEMS ......................................................................................6

3.0 PROBLEM SOLVING..............................................................................................................................7

4.0 TEST TECHNIQUES FOR THE GEARS ..............................................................................................9

5.0 TEST TECHNIQUES FOR THE INSTALLATION ...........................................................................11

6.0 TESTING GEARBOXES IN ISOLATION...........................................................................................13

7.0 ECONOMICS OF PROBLEM SOLVING ...........................................................................................14

8.0 LIGHTLY LOADED GEARS ................................................................................................................16

9.0 DEDUCTIONS FROM T.E. ...................................................................................................................17

10.0 FURTHER READING ............................................................................................................................18

APPENDIX...........................................................................................................................................................19

2001 BGA Technical Publications

The Copyright in this Paper rests with BGA and the Paper's authors. The material has been prepared by the BGA to further a wider understanding of gearing and may be freely used and copied in whole or in part for educational purposes. The Paper is not for commercial use.



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INTRODUCTION TO GEAR NOISE

1.0 Causes of Gear Noise When a gear drive is too noisy there is little that is amenable to mathematics and the first essential is to get an understanding of the particular system and to find out the base cause of the problem. There is no standard general solution and the noise may be generated due to poor gears or a poor installation or both. The essential is to identify the relative importance of errors of design and manufacture and of the installation so that a 'cure' is relevant.

Noise arises normally because a force varies, then acts on an elastic system and usually after travelling some distance inside metal, finds a flexible panel which will act as a loudspeaker. In general if we vary a force's amplitude, position or direction we generate a vibration but for standard involute gearing it is variation of amplitude that dominates, unlike Novikov gearing which also has a large position variation. In a few cases noise occurs due to air or oil being trapped between teeth and expelled axially but this usually only occurs at high speed with wide spur gears.

The force variation which causes the trouble arises because a gear drive is not perfect (it never is) and instead of giving a perfectly smooth transmission of rotational velocity from one shaft to another there is a small relative vibration superposed on the steady velocity. This relative vibration is called T.E. (transmission error) and is defined as the error in the position of the output shaft from the correct position if the drive were perfectly smooth. Although it is defined and measured as an angular error we usually turn it into a distance by multiplying by an effective radius of the gear. Fig. 1 shows a typical T.E trace

full scale is 80 microns

Transmission Error for one revolution

Fig. 1

Typically transmission errors may be about 50 microns at once per revolution (eccentricity) and 5 microns at once per tooth frequency. Errors are roughly independent of size and module; this seems highly unlikely but is true especially of the 1/tooth errors that give most noise.



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The route of the vibration is then shown diagrammatically in Fig. 2. It is simplest to regard T.E. as a small but energetic demon which is busily imposing a relative vibration between the gear teeth in contact. Alternatively as we think of involutes as corresponding to a string unwrapping from one base cylinder and wrapping onto the other, then T.E. is a variation in the length of the 'string'.

This vibration drives the 'internals' i.e. the gear masses which are supported on the 'springs' of flexible shafts and bearings in the gearcases. The internal dynamics will normally have several resonances and may increase or decrease vibration levels from the original excitation. The route to the outside world of the gearcase is through the bearings to the bearing housings which are the first point at which we can easily get at the vibration/noise to measure it; by this time the original 10 μm excitation may well have been reduced to 2 μm level.

TE

microphone

Fig. 2.

The rest of the route is through the gearcase, through the antivibration mounts if fitted, into the support structure, through the support structure and into any panel that can be vibrated easily and hence to airborne noise.



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Fig 3 shows the route with the intermediate stages.

Thermal distortions

Pinion distortion ⇒ ⇓ ⇐ Wheel distortion Gearcase deflection ⇒ ⇓ ⇐ Gearcase accuracy

Pinion movement ⇒ ⇓ ⇐ Wheel movement Pinion tooth deflection ⇒ ⇓ ⇐ Wheel tooth deflection Pinion profile accuracy ⇒ ⇓ ⇐ Wheel profile accuracy

Pinion pitch accuracy ⇒ ⇓ ⇐ Wheel pitch accuracy Pinion helix accuracy ⇒ ⇓ ⇐ Wheel helix accuracy

TRANSMISSION ERROR ⇓

Gear Masses Support Stiffness Combined Damping ⇓ Internal damping response ⇓ BEARING FORCES ⇓

Casing Masses Casing Stiffness Casing Damping ⇓ GEARCASE FOOT

VIBRATIONS

⇓ Anti vibration mounts ⇓ TRANSMITTED

STRUCTURE VIBRATION

⇓ Sound radiating panel ⇓ AIRBORNE NOISE

Fig. 3

The whole system is effectively linear provided the loads are high enough to keep gear teeth and bearings in contact and so the final noise level is directly proportional to the T.E. component at that frequency. Hence a 6dB reduction in noise requires the T.E. to be halved.



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2.0 Theoretical Prediction Problems In principle, given large computing effort, we can predict the noise level that we will get from a gearbox. Take the two actual mating gear tooth shapes, calculate loads and deflections of gears and gearcase and by moving through the meshing of a pair of teeth we can calculate the T.E. we will get under 'static' conditions. Put this 'static' T.E. in as the excitation into a model of the dynamics of the 'internals' and we calculate the forces transmitting through the bearings to the gearcase. These bearing forces will vibrate the gearcase and structure and in turn radiate noise so that output noise power level can be calculated. The flow chart is as in Fig. 3.

In practice, the answer is (putting it mildly) not reliable and the effort in terms of metrology time is uneconomic. We need the T.E. accurate to about 1 μm but it is very difficult (and expensive) to make the many measurements needed to better than 2 μm and the finite element deflection calculations are not proven to better than 2 μm so the initial estimate of T.E. may be out by a factor of 3 either way. The only exception to this is when the gear is such a peculiar shape that it is going to be extremely noisy. Estimates of internal and external resonances are unreliable if excitation is near resonances because the system responses are governed by damping which is notoriously difficult to predict and so must be measured. Vibration isolators, usually based upon rubber, have characteristics which are frequency and amplitude dependent and so it is again advisable to measure rather than predict. Prediction can be very helpful in deciding what shape the gears ought to be made to minimise T.E. when at the design stage but for this it is not necessary to use large programs as relatively simple (60 line) programs are sufficient. A typical program is given in the appendix; this type of program can be extended very easily to estimate dynamic responses even when gears come out of contact. Any results are, of course only as good as the assumptions involved.

In the few, fortunately rare, cases where gears are not hermetically sealed inside a gearcase the sound is generated at the gears themselves and can be very, very much louder since it can get out without the 30dB reduction associated with the casing. Sealing the drive is by far the cheapest solution but where this is not possible internal absorption can help or very precise helical gears must be used.



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3.0 Problem Solving Given an unsatisfactory gear installation there is usually the classic confrontation between gear supplier and installation designer. The "badly made rough gears" argue with the "over-sensitive structure design" and a consultant may be called in as referee.

The first move is to listen to the gearbox and to decide exactly what is causing irritation. It can be the steady whine of a typical back axle, the modulated wow-wow-wow type of noise or the traditional grating or rumbling of a low speed gear. The most common case is a steady note at once-per-tooth frequency, easily identified by a frequency analysis but a useful technique is to record the sound (a domestic tape recorder is suitable) and then either match the sound synthetically or look at the waveform and inspect it. A standard PC with output card can easily be programmed to give variants on gear errors.

It is worthwhile remembering that once per revolution is nearly always at too low a frequency to be heard so low frequency noises are in reality modulation of high frequency tones. It is sometimes possible to mask an irritating noise (at once per tooth frequency) by covering it up with deliberately introduced errors of pitch.

Grumbling noises from a drive are more likely to be caused by pitch errors on the gears and these show up well on a transmission error trace as sudden drops or rises in the curve.

If the noise is standard once-per-tooth and harmonics, the attention can then focus on whether the T.E. component at 1/tooth is large, suggesting gear improvement is needed, or whether it is small, in which case it is probably hopeless trying for further improvement and the installation must be made less sensitive to this frequency. The obvious question is 'what is reasonable' and here the trouble starts because, as might be expected, the answer depends on the industry, past history, economics and often politics. Curiously, contradicting common sense, the answer does not vary much with the size of the gear. As a very rough rule > 10 μm suggests a poor gear which definitely should (and could) be improved while < 2 μm is very good. Permissible levels tend to reduce steadily because every design improvement tends to reduce weight which does not help and customers are steadily raising standards. Each industry has to build up its own data bank of what it needs (or can get away with) for each installation. As usual, a quick check of a competitor's T.E. can be very educational.

In the intermediate area where both gear and installation are unsatisfactory the problem becomes an economic problem since if either the installation or the gear can be improved to meet the noise specification, the choice is solely an economic choice. If you are unlucky then both the installation and the gears must be attacked to get the necessary improvement. The debate on whether the gears are of 'reasonable' quality for the job has no nice clearcut answers. On a production line the test experience soon builds up a pattern of permissible T.E. levels. Most firms start off by having to prove to themselves that T.E. and noise are directly connected and in the process learn what levels of T.E. are tolerable.

The techniques are the same regardless of the size of the gear or the materials used. In practice most gears are steel and vibration is unaffected by whether they are hardened or not. The use of plastic gears greatly reduces the Youngs Modulus which reduces noise but the Transmission Errors are usually much higher. This tends to give an improvement at high speeds, above resonances but poorer results at low speeds, below resonances. Sintered gears have much the same vibration characteristics as solid gears but are usually less accurate and so tend to be noisier and cannot take such high loads.



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Multiple shaft drives such as split power drives or epicyclic gears present a problem because each contact should be checked separately and it may not be possible to get access. It is possible to test the drive in its complete state but not only must the T.E. be checked but any lateral movements of the flexibly mounted members must be monitored to detect trouble.

Noise or vibration cancellation is a fashionable approach which is feasible in a limited number of cases. It can help to reduce vibration provided that the frequencies are not too high and that the disturbing vibration is fairly steady in both amplitude and frequency. The snag is that if vibration exciters have to be fitted, they are expensive, heavy and delicate. In practice this tends to be a desperate last resort and it is not an economic solution for the vast majority of problems, though large Diesels have been cancelling second order for most of this century.



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4.0 Test Techniques for the Gears It is very simple to say that "the T.E. at once per tooth must be below 5 µm" but measurement of T.E. is not a standard procedure in the gearing world and so there may be a blank look at the mention of T.E. although DIN now have a (not very satisfactory) standard for specifying it. Some branches of industry such as printing have been using T.E. for a generation whereas others will refuse even to discuss it.

Testing T.E., usually on a 'single flank checker', involves meshing a pair of gears in the correct alignment with highly accurate rotary encoders on each shaft end. The equipment is incredibly accurate, 0.1 μm typically, and the normal production control equipment (Gleason or Klingelnberg) runs at about 10 rpm with just enough torque to keep the teeth in contact. The actual test takes about half a minute and the results can easily be checked by eye or put into a frequency analysis routine which comes as standard with the equipment.

This approach works well for production control purposes. The objective is usually not to get zero T.E. under these circumstances but to get the shape of T.E. which will be exactly cancelled out by the (estimated) elastic deflections of the system under load to give the best possible T.E. under load (zero). This requirement results in the rather curious phraseology of people talking about "the error in the Transmission Error". Fig. 4 is a simplified version showing how two profiles with the correct tip relief will give a shape of T.E. at zero load which under operating torque gives small T.E. The full version of this mapping as load varies is known as a 'Harris' plot following the original work by Gregory, Harris and Munro.

Wheel Tip reliefPerfectinvolute

PinionTip

Combined error at no load

Combined error at full load

Next pair

elastic defl.

Fig. 4

Testing at zero load and estimating deflections under load works well for less critical applications with parallel shaft spur gears which are mounted well and so do not misalign under load. More critical installations need testing of the gears 'in situ' to incorporate gearcase inaccuracies and under correct torque to generate the correct gear, gearshaft, bearing and gearcase deflections and distortions. Unfortunately full torque and low speeds are liable to scuff the gear teeth or wreck plain bearings and the standard equipment is not designed to transmit high torques, to run at speed or to be able to deal with torsional vibrations on the input shaft.

One solution is to bring Mahomet to the mountain and to fit 'portable' encoders onto the equipment in situ. They need a free shaft end and a mounting plate to support the encoder body but they are small, robust and portable. Speeds up to 6000 rpm are allowable,



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accuracy is more than sufficient and valid vibration information is available up to about 1000 Hz so that a typical 600 rpm on a 20 tooth gear presents no problem for up to 5th harmonic of tooth frequency. Ideally the equipment or test rig is run fast enough to protect the teeth but slowly enough to avoid internal component resonances. Since the test setup supplies all the mounting and alignment this option is much cheaper than the production control equipment and is very suitable for development purposes as well as being very portable.

It is sometimes possible to deduce whether corrections to the teeth are working well or not by varying the load torque on the drive and checking whether the vibration is a minimum at the design torque. An alternative method of testing in situ uses tangentially mounted accelerometers to avoid the need for a free shaft end but sliprings (or telemetry) are needed and although once per tooth information is valid the once per revolution output is rubbish for horizontal shafts.



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5.0 Test Techniques for the Installation If the gear pair has a respectable T.E. but there is too much noise the installation must be tackled. The first question is 'is the system sitting on a resonance?' and if there is a resonance, 'how sharp is the resonance?'.

The best technique is to take a resonance curve of the response of the system to excitation at the gear; this excitation consists of the relative movement between the mating flanks and so is the T.E. We use the mesh as the excitation, vary the frequency by varying motor speed, and we then know that we have a constant input of, say, 5 μm at once per tooth frequency and a sweep through the speed range (at constant torque) will give the response plot either using computer order tracking or, much simpler by hand. A.C. motors usually run at fixed speed but either a D.C. motor can be substituted for the testing or, more elegantly, a variable frequency drive can be used with a standard A.C. motor as the price of these drives has dropped greatly. Fig. 5 shows a typical plot.

response

frequency

once per tooth

3rd harmonic

Fig. 5

If it is not possible to use the gear drive itself as the excitation the next best thing is to excite at the first accessible point in the chain, at the bearing housing, using a conventional electromagnetic vibrator. Any resonances of the internals of the gearbox will be missed by this approach but it is usually the supporting structure that is at fault. It is sometimes necessary to reciprocal test by exciting at the gearbox feet and measuring at the bearings.

Once a resonance has been located, development can proceed along relatively standard lines of sitting at the trouble frequency and measuring the mode shape of the vibration. In most cases this will give a clear and easy identification of the cause of the trouble and is a simple, quick and cheap method which does not involve expensive equipment. The mode shape of a panel is a good indicator of whether the panel needs stiffening or not. If centre amplitudes are greater than those at the sides, thickening or ribbing is required. If less, the panel should be left well alone.

If the resonance is too near an excitation frequency there is a choice of altering natural frequencies or of altering tooth numbers if stress ratings allow. Generally high frequencies



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(>500Hz) should be raised and low frequencies (<250Hz) should be lowered but it is safest to look at each case individually.

A rather important link in the 'structure transmission' part of the vibration path is the vibration isolator if there is one fitted. Isolators are very rarely chosen with the idea of stopping tooth frequency since the usual requirement is to stop once-per-revolution. Simple theory suggests that an isolator which is good for once per rev. will be very good for tooth frequency (> 20 per rev) but in practice isolators will often be surging at high frequencies and so ineffective. This may require a two stage isolator but in some cases isolators have not been fitted since they cannot control once per revolution but they could still help at once per tooth. In some cases a non-linear isolator is needed if low noise is required under low torque conditions (motorway cruising) but the system must cope with high loads at times (bottom gear acceleration).

Once a vibration has entered the main structure it is difficult to stop. The best technique may then be to stiffen, damp or break up the elegant 'designer' flat panels which act as loudspeakers. Sound intensity measurements can be useful to identify the offending panels but using an accelerometer by hand is as effective and much cheaper. Occasionally the sound path is via the input or output shaft so it is necessary to isolate in the shaft with a rubber coupling; on a boat a propellor can act as a good loudspeaker.



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6.0 Testing Gearboxes in Isolation It is quite common for gearboxes to be noise tested on a separate rig before being built into equipment. This technique has its uses but should only be employed with care since it can be deceptive.

The gearbox mounting will be different in vibration aspects in the test rig and so may influence the system natural frequencies and the important frequencies radiated from the test rig will not be the same as those from the installation. In general it is better not to use a microphone since they do not give local effects but to use accelerometers mounted at a critical place such as a bearing housing and the results should only be compared at the frequency known to give trouble in practice. Best is to measure T.E. at low speed to check the T.E. and separately check the installation which is likely to be more consistent than the gear errors.



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7.0 Economics of Problem Solving The possibilities are usually to improve the gears, improve the vibration isolation, improve the installation or to select out the better gears and scrap the others. Economics plays a dominating part in the choice.

For one-offs, the capital effort of development of the structure or installation is not worthwhile and so it is usually cheapest to obtain a better set of (ground) gears. The assumption is that ground gears will be better but this is not necessarily so, especially for pitch. Much more important than whether the gear is ground or not is whether it has been properly designed with low noise as a target since it is, as yet, unusual for gears to be designed for quietness. For a long production run, grinding would often be uneconomic and so the effort of improving the design and the vibration response of the installation will pay better dividends. In particular, care should be taken to use the correct 'long' or 'short' tip relief on spur gears and to have the correct (axial) facewidth on helical gears.

As the (customer's) demands become steadily more critical it is necessary to reduce T.E., to improve the installation and to reject those sets of gears with higher T.E. Rather than set an unobtainable low tolerance for all the gears produced it is cheaper to manufacture to reasonable accuracy levels then 100% single flank inspect the gears as matched pairs. This, though possibly giving 5-10% reject rate is often the cheapest approach and avoids the costs of building poor gears into equipment then having to strip out. It is important to test the gears as matched pairs since this reduces scrap rates greatly as well as halving the test times. As production accuracies rise the accuracies of gears can approach those of 'master' test gears and it is then not possible to get 'masters' an order more accurate.

h h hhh

h h h

desired size difference

distr

sizeA B C D A B C D

Fig 6

If the permissible tolerance band in a meshing pair is 2h, unmated gears can only be allowed a production tolerance of +/- h and so only those gears between sizes B and C are permitted whereas testing in pairs will also allow a majority of the gears in ranges A to B and C to D to be used and a few outside this range. Even a gear of size A, (2h below nominal) will mate satisfactorily with all mating gears less than nominal, 50% of the production.



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The idea of 100% single flank testing of gears in pairs on production will often produce hysterics. An economic argument can however be made out that it is a lot cheaper than not testing. Cycle times of about a minute can be achieved either on commercial equipment or a test set up so that even at an initial cost of the order of £100,000 for a commercial machine the cost for testing is of the order of £20/hr and for high throughput a test cost of less than £1 per pair is easily justified in relation to the cost of building gears into a machine and having to strip them out for replacement.

For development purposes the 'portable' equipment costs about £14,000 and is often necessary for critical applications. The limitation is that it cannot be used if free shaft ends are not available though the techniques using tangential accelerometers (and slip rings) can be used provided once per revolution information is not important. The portable encoder equipment works over a very wide range of conditions but correspondingly is less suitable for use by unskilled labour.



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8.0 Lightly Loaded Gears The term lightly loaded refers to gears which do not stay in contact all the time. This may be because the driving torques are low or it may be because very large inertias are being driven. An estimate of whether or not this may be happening can be made by multiplying the effective moment of inertia of the lighter side of the system by the acceleration corresponding to the T.E. at the operating speed. Accelerations can be rather high despite the small amplitudes since at 1600 Hz tooth frequency 5 μm corresponds to 50 g acceleration so an effective mass of 20 kg would need 1 tonne force to keep the teeth in contact. If the torque trying to keep the gears together is less than this, then loss of contact will occur followed by an impact which gives high force levels and vibration excitation at a wide range of frequencies. With these systems the noise and vibration are very sensitive to applied load, unlike linear systems. The converse of this is that the vibration levels may be relatively insensitive to T.E. until it drops below a critical level.

Fig. 7 shows a representation of what happens as the T.E. excitation forces the gear teeth out of contact and there is a period, which can be some 90% of the time, with the teeth out of contact followed by a short impact on the rising part of the T.E. curve to propel the pinion into flight again.

response

static transmission errortime

Fig. 7

Peak forces can be estimated as contact stiffnesses are known and come out surprisingly high, up to 15 times the mean force and so can be much greater than "fully loaded" forces. In some cases the gears are driven right across the backlash and an impact on the trailing flank returns the gears at higher velocity. This can give a very powerful vibration with tooth forces an order higher than under full load

In some cases there are torsional vibrations in the system due to Diesel engines or to varying load torques and the angular accelerations due to the torsional vibrations must be added to those of the T.E. to assess whether loss of contact is likely to be occuring.

The standard rules do not apply for this type of problem though some estimates can be done and eventually if the T.E. at once per tooth ( and the torsional vibration) is reduced sufficiently the loss of contact and subsequent impacting will stop. Altering the installation is usually not possible and even if it were, moving resonances may well make the vibration worse. This is one of the problems where outside help should be sought. Loss of contact also may occur when gears in a multispeed drive are meshing but not loaded. The clue to this happening arises from the observed noise frequencies not corresponding to the driving



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mesh frequencies and the observed vibration can appear to be almost 'white noise' i.e. completely random. This is one of the times that frequency analysis is of no help and may be extremely deceptive.

9.0 Deductions from T.E. In theory (forgetting error problems) we can work forward from lots of profile, helix and pitch measurements to deduce T.E. and hence drive smoothness. At first sight it should be possible to reverse the process and deduce the basic errors from T.E. measurements but in practice there is limited scope.

Spur gears are the simplest; we cannot say anything about helix but we can get an idea of what the central 60% of the profile looks like. Pitch information, obtained by sampling the T.E. once per tooth at the pitch point, is directly useful and has effectively eliminated conventional pitch measurements for those who have single flank checkers.

Helical gears give trouble. Pitch information may still be relevant but the once-per-tooth component is a hopeless mixture of profile and helix effects so it is not possible to disentangle the two.This sounds depressing but does not matter much in practice since the important thing is to know that something is wrong to prevent poor gears being built into equipment. The detailed checks on exactly what is wrong then need only be done on the occasional failure and often the T.E. will give a clue to problems, especially if allied to blue contact checking as normally carried out to check strength.

It should be emphasised very strongly that there is not necessarily any connection between noise and strength of gears. In some cases it is possible to make a design quieter by sacrificing some strength or vice versa. A good modern design will however give both strength and quietness.



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10.0 Further Reading Traditionally, gear textbooks have tended to concentrate on the elegant pure geometry of gears and avoid messy topics like stresses and noise.

Most of the current knowledge and ideas on the stressing side are gathered in the standards. In Europe, DIN 3990, its derivative ISO 6336 and the derivative BS436 (1988) and in the U.S.A. the AGMA 2001/2101. The approaches are similar on surface stresses but differ on root stresses with A.G.M.A. possibly too optomistic. Noise of gears and the associated causes are not covered in the standards.

Generally noise and vibration of gears is little discussed. A book such as Drago, 'Fundamentals of Gear Design ', 1988, Butterworths, Boston. gives a general view of the whole subject of gearing. The book "Gear Noise and Vibration", J. D. Smith, Marcel Dekker, 1999 ISBN 0-8247-6005-0 deals in much more detail with the topics in this handout and with some of the analysis methods for vibration and for monitoring. There have been a large number of papers published on gear noise but the majority should be treated with great caution. Massive computer programs, however large, suffer from the "rubbish in = rubbish out" syndrome and often the base assumptions for the programs are unreliable. It is easy to write programs which will predict that a given design of gearbox will be silent but they are not much help when it turns out to be noisy. However, simple programs will help at the design stage to produce designs which are reasonably quiet and reasonably tolerant of errors.

For those who wish to learn about detailed aspects of gears there are good courses at Newcastle lasting several days covering metrology, stressing and vibration - contact The Design Unit, Stephenson Building, Mechanical Engineering Department, Claremont Road, University of Newcastle 0191-2226192.

The British Gear Association has commissioned a series of reports on topics such as gear vibration, profile correction, etc. which are available to members of the organisation.



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Appendix Transmission Error Estimation % Program to estimate static transmission error % First enter known constants or may be entered by input facew=0.125; % arbitrary 25 slices wide gives 5 mm per slice baseload = input('Enter base radius tangential applied load '); bpitch=0.0177; % specify tooth geometry 6mm mod misalig=40e-6; % total across face line 4 bprlf=25e-6; % tip relief at 0.5 base pitch from pitch point strelief = 0.2; % start of linear relief as fraction of bp from pitch pt tanbhelx=0.18; % base helix angle of 10 degrees tthst = 1.4e10; % standard value of tooth stiffness relst=strelief*bpitch; % start of relief line 9 ccp = 10e-6 ; % arbitrary interference at pitch pt in m at start hor = ones(1,25); % 25 slices across facewidth x = (facew/25)*(-12:12); % dist from facewidth centre crown = (x.*x)* 8e-6 /(facew*facew/4); % 8 micron crown at ends te = zeros(1,32); % line 13 for k = 1:32 ; % complete tooth mesh 16 hops ********** for adj = 1:15 % loop to adjust force value >>>> for cl = 1:5 ; % 5 lines of contact possible? $$$$$$$$$$$$ yppt(cl,:)=x*tanbhelx+hor*(k-16)*bpitch/16+hor*(cl-3)*bpitch; rlief(cl,:)=bprlf*(abs(yppt(cl,:))-relst*hor)/((0.5-strelief)*bpitch); posrel = (rlief(cl,:)>zeros(1,25)) ; % finds pos values only actrel(cl,:) = posrel.* rlief(cl,:); % +ve relief only interf(cl,:)=ccp*hor+misalig*x/(facew)-actrel(cl,:)-crown; % local int posint = interf(cl,:)>0 ; % check interference positive totint(cl,:)=interf(cl,:).*posint ; % line 23 end % end cl contact line loop $$$$$$$$$$$$ % disp(round(totint*1e6)) ; pause % only if checking pattern ffst = sum (sum(totint)); % total of interferences ff = ffst * tthst * facew/25 ; % tot contact force is ff residf=ff- baseload ; % excess force over target load % disp(residf) ; pause % only if check force convergence if abs(residf) > 50; % line 28 ccp = ccp - residf/(1.0e9) ; % contact stiffness about 1e9 else break % force near enough end end % end adj force adjust loop >>>>>>> if adj==15; % line 34 disp('Steady force not reached'); pause end te(1,k) = ccp * 1e6; % in microns intmax(1,k) =max(max(totint)); % maximum local interference end % next value of k ********************* xx = 1:32; % steps through mesh line 40 peakint = max(intmax) ; % max during cycle



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contrati = 1.6 ; % typical nominal contact ratio stlddf = peakint*facew*contrati*tthst/baseload ; % peak to nominal disp ('Static load distribution factor') ; disp(stlddf) ; figure;plot(xx,te);xlabel('Steps of 1/16 of one tooth mesh'); ylabel('Transmission error in microns'); end

In the program the first 10 lines (not counting % comment lines) set up the constants and an arbitrary starting position of 10 ?m interference. Line 12 generates the crowning relief proportional to distance x (from face centre) squared. Line 14 starts the main loop to do the 32 steps corresponding to 2 complete tooth meshes. Line 15 starts the force adjustment loop which is set arbitrarily to 15 convergences. Normally the loop will converge within 50N force (roughly 0.05 micron) well before 15 tries and will break out in line 31. If not, a warning is displayed and the program is stopped.

Along each line of contact (line 16) the distance (yppt) of each x slice contact from the pitch line is the sum of the helix effect, the movement due to the 32 steps and the movement due to the change from one contact line to the next. The tip reliefs are calculated in line 18, and those that are positive detected in line 19 and the negative ones put to zero in line 20. Line 21 sums the effects of body interference, misalignment, tip relief and crowning, then in line 23 only the positive interference values are retained. All values of interference are summed and multiplied by the slice stiffness to give the total contact force ff.

This force is compared with the desired contact force and the difference is divided by a guessed overall mesh stiffness to adjust the pitch point interference ccp. The loop repeats until the agreement is within 50N (11 lbf ) in this case. Finally the next step of the 32 steps is selected and convergence is fast because the starting value of ccp will be nearly correct. Plotting the results gives an indication of whether curious contact line length variations are occurring. If so the display instruction after line 24 can be activated to look at the interference pattern. Also the peak loading per unit length of tooth is compared with the nominal average loading to give the load intensification factor.

BGA Training Notes TN2-1: Gearbox Systems: Miniature Instrumentation for Troubleshooting in Rotating Machinery Applications


Page 1


Gearbox Systems:

Miniature Instrumentation for Troubleshooting in Rotating Machinery Applications Authors: J. Rosinski, D.A. Hofmann, Design Unit, Gear Technology Centre, University of Newcastle upon Tyne, Stephenson Building, Claremont Road, Newcastle upon Tyne, NE1 7RU [email protected] http://www.newcastle.ac.uk/~nmecheng/



Page 2

Index

1. MESH ALIGNMENT DETECTION AND CORRECTION TECHNIQUES.........................................4 2. FATIGUE LIFE EVALUATION USING RAINFLOW COUNTING DATA LOGGER ......................6 3. MEASURING TORQUE AND TORSIONAL VIBRATION INSIDE A DIESEL ENGINE ................7 4. RAIL TRACTION FAILURE INVESTIGATION..................................................................................9 5. DIESEL ENGINE OIL PUMP NOISE INVESTIGATION ..................................................................11 6. WIND GENERATOR NOISE TROUBLESHOOTING .......................................................................13 7. CONCLUSION......................................................................................................................................14

REFERENCES...............................................................................................................................................14





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ABSTRACT The paper presents several case studies, which demonstrate the use of miniature instrumentation for troubleshooting power transmission systems in rotating machinery. The following case studies are discussed:

• telemetry system for mesh alignment correction in large, high performance gearboxes · marine diesel engine cam shaft stress and torsional acceleration measurement · rail traction failure investigation

• diesel engine oil pump noise investigation

• wind turbine noise troubleshooting

• roller coaster chain failure troubleshooting

The cases presented demonstrate that by using modern instrumentation techniques data can be acquired which enables difficult rotating machinery problems to be solved.

Keywords: Rotating Machinery, Diagnostics, Instrumentation, Troubleshooting, Rainflow Count, Telemetry

INTRODUCTION Power transmission system problems can result in high vibration or noise levels but also in catastrophic failures of critical transmission components.

Any failure of critical power transmission components is a cause of concern and requires corrective action. Often failures of uncritical components may be an early warning of a more serious problem with the transmission system.

Sometimes failure results from problems with materials, heat treatment or incorrect sizing of components. These problems are usually relatively simple to detect and rectify.

Much more difficult to overcome and understand are failures that result from higher than anticipated loads in the transmission. Such problems often arise from dynamic overloads that are generated due to the complex dynamic response of the power transmission system.

In many practical applications, the use of relatively simple computing facilities, combined with a reliable dynamic model, may provide sufficient understanding and consequently a solution to the problem. In other cases the model becomes so complex that its accuracy can not be verified, and then the only way to understand the problem is to carry out practical in-service measurements.

In spite of significant advances in condition monitoring techniques, measurements of acceleration on the outside of machine casings, often do not provide sufficient evidence to identify the ultimate reason for the failure. This is due to the unknown relation between the amplitude of linear acceleration measured on the outer machine casing and the amplitude of loads transmitted in the rotating machinery.

Ideally dynamic torque, dynamic bearing loads, torsional vibration or stressing should be measured directly on the rotating components to identify the cause of the problem. However, such measurements are often difficult and may call for more advanced instrumentation techniques.

In this paper several case studies are presented, which use different measuring and modelling techniques to solve problems in various rotating machinery applications.



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1. MESH ALIGNMENT DETECTION AND CORRECTION TECHNIQUES Large industrial, turbo generator and marine gears often suffer from the inherent problem of misalignment, which causes increased noise, vibration and in many cases gear failures. It has been shown [1] that gear misalignment errors of 100μm per 200mm face width can cause more than a 3-fold increase in gear stress and up to a 10-fold increase in gear noise. During the design stage the misalignment can be minimised by designing bearing offset corrections to compensate for torsional wind-up, bearing clearances, and theoretical shaft deflection. Additional corrections can be made after visual inspection of the mesh contact patterns. However, none of the above methods allows for complex gearbox distortion during normal gearbox operation.

A practical method to investigate in-service gear misalignment has been developed and successfully used by the Design Unit, at the Gear Technology Centre (Newcastle, UK). This method involves strain gauging the root of the teeth at a number of transverse positions across the facewidth and measuring dynamic loads across the facewidth during normal gear operation. Special strain gauging techniques have been developed which allow permanent placement of strain gauges at the root of the teeth, with total thickness of the complete installation not exceeding 150Iμm, and capable of continuously running at speeds of up to 6000 rpm.

In this application all critical electronic components, including signal conditioning and multi-channel telemetry systems are mounted on the gear elements themselves and non-contact power and signal transmission provides an extremely compact and rugged system.

A detail of strain gauge installation on a large single helical pinion is shown in Photo 1. Typical examples of load intensity traces measured before (a) and after (b) correcting mesh misalignment are shown in Fig.1 below.

Photo 1 Instrumented pinion for in service gear

Figure 1 Typical load intensity plots before (a) alignment measurements and after (b) correcting mesh alignment



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The technique described in the above case study allowed significant improvements in gear noise and vibration to be achieved on large high speed gearboxes, and also permitted an increase of power density with a significantly reduced risk of gear failures.



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2. FATIGUE LIFE EVALUATION USING RAINFLOW COUNTING DATA LOGGER Miniature unattended data loggers with on-board rainfall count algorithm have been successfully used in environments with very high shock and high EMR. An embedded microcomputer evaluating the fatigue life of the roller coaster chain drive is shown in Photo 2. Another interesting application of the rainflow count data loggers is shown in Photo 3 where they were mounted on two motorised and two trailer car axles for long-term monitoring of torsional and bending axle stresses.

Typical rainfall count data acquired from long term in-service monitoring is shown in Fig. 2. The data acquired in the form of rainflow count (mean-range classification matrix) allows easy prediction of fatigue life.

Photo 2 - An embedded microcontroller evaluating chain fatigue life on a roller coaster application.

Photo 3 - Data logger fitted on train axle for fatigue life evaluation (simultaneous measurements of torsional and bending stresses

Fig. 2 - A typical Rainflow Count result acquired on a chain driving a roller coaster



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3. MEASURING TORQUE AND TORSIONAL VIBRATION INSIDE A DIESEL ENGINE Governor drive failures were experienced during the prototype development of a V16 marine diesel engine. These were typically fatigue failures of bevel gears, which drive the governor drive shaft.

A typical pattern of bevel gear failures emerged, which showed that tooth damage always occurs at the same angular positions corresponding to fuel injection timing. Closer analysis of the bevel gears showed that substantial dynamic overloads must have been taking place in the drive system. The nominal transmitted torque required to drive the governor is approximately 7 Nm. From gear strength calculation it was clear that the transmitted dynamic torque must exceed 100 Nm to damage the gears.

To investigate the cause of these very high torques, the dynamic torque in the governor drive shaft was measured simultaneously with the torsional acceleration of the camshaft. Measurements of dynamic torque were difficult due to the very restricted space inside the diesel engine. A detail showing the end of the camshaft and governor drive is shown in Fig. 3.

To measure the dynamic torque transmitted during normal engine operation, strain gauges, arranged in a full bridge configuration, were fitted at the relieved section of the drive shaft. An axial hole was provided in the shaft centre, which allowed three coin-cell batteries, with an isolating nylon sleeve, to be mounted inside the shaft. The existing cavity, between the upper part of the shaft, and the splined adapter plate, was used to house all the electronic components. These included a stable DC bridge supply voltage regulator, a precision strain gauge amplifier, filter, and a voltage controlled oscillator (VCO). An aerial was wound on the lightweight carrier, which was clamped between the shaft and the splined adapter, as shown in Fig. 3. A miniature pick-up probe was bolted inside the engine. Thin, coaxial cable, routed through the case, was used to feed the frequency-modulated signal to a discriminator unit (base station).

By instrumenting the shaft in the above way, the system dynamics was not modified in any significant way.

Fig. 3 - Detail showing engine end, with miniature instrumentation for torque andtorsional acceleration measurements on the cam shaft and governor drive shaft.



Page 8

The torsional acceleration measuring system did not impose severe problems in terms of space limitation. Two miniature capacitive accelerometers, encapsulated inside two aluminium half-rings, were rigidly clamped on the camshaft as shown in Fig. 3. Accelerometers were mounted in-phase on the opposite sides of the shaft. The signal from the two accelerometers was subtracted, and then amplified in the subsequent electronics to measure torsional acceleration. Fig. 3 shows a general arrangement of the modified end-cap, with the instrumentation/battery compartment, an aerial, and a pickup probe.

The bandwidth of the oscillator-discriminator telemetry system used for torque and acceleration measurement covered frequencies from DC to 2kHz.

Dynamic torque and torsional acceleration signals were simultaneously recorded on a dual channel signal analyser. Typical time traces, measured at 1000 rpm, for 1I2 of the maximum engine load are shown in Fig. 4b and waterfall order tracked run-up torque measured for'/� maximum load in Fig. 4a.

A detailed investigation of the torque and torsional acceleration showed that at low loads, 'bursts' of high torque in the drive shaft roughly correspond to 'bursts' of torsional acceleration. For higher engine loads, dynamic torque transmitted in the drive shaft exceeded nominal torque by more than 10-fold. With such a high dynamic torque component, fatigue failure of the gears is plausible, especially when account is taken of bevel pinion inertia effects.

Closer analysis of the recorded signals confirmed a wide band character of frequency spectra with a dominant frequency of 1450 Hz for torsional acceleration, and approx.1250 Hz for torque transmitted in the governor drive shaft. '

The waterfall plot presented in Fig. 4a clearly shows that high dynamic loads in the drive shaft are predominantly caused by a resonance at approx.1250 Hz. This can be seen as a constant frequency curve shown in the right top corner of the graph. Four straight lines at the left side of the diagram indicate response to kinematic excitation occurring at 8,16, 24 and 32 times per camshaft revolution. These components correspond to fuel injection frequencies, but they appear to be much less important than oscillation due the resonant response of the timing gear system.

The above case study describes a very useful application of two miniature telemetry systems, which were successfully installed and operated inside a diesel engine. The examples presented show that measurements of the torque and torsional acceleration provide invaluable quantitative data on the dynamic behaviour of geared systems. Similar data can not be derived from measurements of acceleration on the engine structure.

Fig. 4 - Order tracked waterfall torque run-up test (a) and (b) torsional acceleration (upper: trace) with torque (lower trace) acquired with inside engine instrumentation



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4. RAIL TRACTION FAILURE INVESTIGATION Unexplained failures of rubber couplings used in a train traction system have occurred shortly after putting new trains into service. The failures were due to breakage of the load carrying nylon bundles protected by an external rubber moulding.

An extensive investigation was initiated on a specially built test rig to understand the reasons for the unusual coupling failures. As none of the tests carried out on the test rig explained the reason for multiple coupling failures, the decision was made to carry out in service load measurements during normal train operation. For these tests a flexible coupling was instrumented to measure the transmitted torque and temperature of the rubber elements.

A schematic diagram showing the instrumented coupling is shown in Fig. 5. To reduce the risk of losing important information, both temperature and torque sensors were duplicated, with one set of data transmitted via a slip ring, and a second via miniature telemetry systems. The instrumented coupling shown in Fig. 5 was fitted on the train, to investigate _in-service loads and temperatures occurring during normal train operation. The results from the first measuring session carried out during torrential rain revealed that a significant oscillating torque component was transmitted at 103 Hz during wheel slip. Typical results showing dynamic torque recorded during the transition from accelerating to braking are shown in Fig. 6.

Fig. 5 - Schematic diagram showing a modified flexible coupling to measure dynamic loads (telemetry and slip ring channel) and intemal temperature of the coupling (telemetry and slip ring channel).



Page 10

An explanation of a mysterious 103 Hz frequency component was found within the wheel set itself. Torsional resonance of the wheel set measured with impact excitation was found at exactly 103 Hz. With an axle mounted gearbox, asymmetrically positioned close to one wheel, high torsional oscillations of the wheel set during the slip, caused by self-excited torsional vibrations, which were amplified by the gearbox and resulted in an unacceptably high dynamic torque component transmitted in the rubber coupling. After fitting a strain gauge telemetry system on the axle and taking simultaneous measurements of axle stresses and dynamic torque transmitted in the coupling, this dynamic model was fully verified. Modifying the microprocessor traction control system, to minimise wheel slip, finally solved the problem.

Fig. 6 - Dynamic torque transmitted through the coupling. Transition from train acceleration to braking during torrential rain (bursts of the signal correspond to high amplitude toque oscillations at 103 Hz).



Page 11

Fig. 8 - Schematic diagram showing original and modified gerotor pump sealing land design.

5. DIESEL ENGINE OIL PUMP NOISE INVESTIGATION An oil pump noise problem was identified in the new design of a 1.8L diesel engine. To understand the reasons for excessive noise generated by the pump, it was necessary to measure torque and torsional acceleration of the pump drive shaft. The oil pump was of a very compact gerotor design, which made fitting the required instrumentation a challenging task.

Special instrumentation was designed and fitted on the pump drive shaft as shown schematically in Fig. 7. This measured torque transmitted through the pump drive shaft, and torsional acceleration of the gerotor pump element.

This instrumentation allowed qualitative and quantitative information on the origin of pump noise and vibration to be acquired very effectively.

It was found that the main kinematic excitation features were due to skew gears and gerotor pump lobe passing order components. However, closer inspection of torsional acceleration traces showed, that the most significant order components were caused predominantly by lobe passing orders, with negligible signal corresponding to skew gear drive excitation orders.

Measurement of dynamic pump pressure with a miniature piezoelectric transducer confirmed that pressure pulsation, was the major excitation source in the torsional �acceleration and noise spectra. A

Fig. 7 - Miniature instrumentation for torque and torsional acceleration measurement fitted on the gerotor pump drive shaft.



Page 12

simple modification of the sealing land design, as shown in Fig. 8 resulted in substantially reduced hydraulic pressure pulsation.

Pump tests after introducing optimised sealing land modifications reduced overall noise generated by the pump by an average 7 dB (A) measured at 1 m distance from the pump.



Page 13

6. WIND GENERATOR NOISE TROUBLESHOOTING Wind energy, is widely accepted as being environmentally friendly. However, wind generators utilising gearbox speed multipliers, can suffer from high noise levels. This problem is more significant in cases where several wind generators are placed within a relatively small area. In some reported cases, noise levels measured around wind farms, exceed 80 dB (A), which causing noise nuisance and frequent complaints.

To better understand the relative noise contribution from the gearbox, tower and blades, it was necessary to measure acceleration directly at all of these sources. Whilst measurements of acceleration on the gearbox and tower structure did not represent any technical difficulties, measurements of acceleration on the blades during normal operation of the generator required special techniques.

To measure blade acceleration during normal wind generator operation, a voltage mode accelerometer, followed by an ICP amplifier was interfaced with a miniature slip ring capsule to power the accelerometer and return the measured signal to a stationary recording instrument. The operator, suspended on a crane at 30m height carried out the complete installation and actual measurements. A detail showing a wind generator rotor with a miniature slip ring capsule is shown in Photo 3.

The results of noise and vibration measurements provided important information not only on the relative contribution of noise sources, but also about the primary reason for excessive noise excited by the speed-up gearbox. From the frequency spectra measured on the rotating blades it was found that the predominant noise frequencies were due to ghost order components (71st and 73rd orders), which were `imprinted' in the output stage pinion and

Photo 3 - Detail showing accelerometer and slip ring mounted on the rotor nose of a 330kW wind generator and an overall view of the generator.



Page 14

the output stage wheel. This was a result of excessive kinematic error in the gear grinding machines used in the gear manufacturing process. The tooth contact frequencies, although smaller than ghost components, were also outstanding in the spectra. A typical spectrum of vibration measured on the blade during normal operation of the generator is shown in Fig. 9.

7. CONCLUSION The case studies presented in the paper emphasise the importance of practical measurements. In many cases, direct measurements of torque, stresses or torsional acceleration on the rotating components can not be substituted by simpler indirect measurements on the machine case or enclosure. Constant progress in microelectronics and micromechanics allows use of advanced measurement techniques in very inaccessible and demanding environments, where conventional measurement techniques can not be easily used. Problem solving in rotating machinery can greatly benefit from these new measurement techniques.

REFERENCES 1. Noise and vibration of high-speed marine gears, Internal Report, Design Unit, DU-1272, MoD (Navy),1995.

Fig.9 - A typical vibration spectrum measured on the rotating blade showing high peaks or corresponding to ghost frequency components (71st. and 73rd.order) and other characteristic gearbox Tooth Contact Orders (22,27,34,68). Wind speed 17 m/s.

BGA Training Notes TN-2: Gearbox Systems - Problems and Solutions


Page 1


Gearbox Systems: Problems and Solutions

Author: D.A. Hofmann, Design Unit, Gear Technology Centre, University of Newcastle upon Tyne, Stephenson Building, Claremont Road, Newcastle upon Tyne, NE1 7RU [email protected] http://www.newcastle.ac.uk/~nmecheng/



Page 2

Index

1. INTRODUCTION .....................................................................................................................................3

1.1 DESIGN LOAD BASED ON APPLICATION FACTOR KA ................................................................................4 1.2 DESIGN LOAD BASED ON LOAD HISTOGRAM............................................................................................5

2.0 ON-LINE STARTING...............................................................................................................................8

2A STEEL FURNACE DRIVE ..........................................................................................................................10 2B COMPRESSOR DRIVE...............................................................................................................................11

3. THE EFFECT OF ENGINE CHARACTERISTICS ...........................................................................13

4. SIMILAR SYSTEMS ..............................................................................................................................15

5. EFFECT OF COUPLING CHARACTERISTICS ...............................................................................16

6. GEAR OVERLOADS DUE TO GEARCASE MOUNTING...............................................................20

A) OUT-OF-BALANCE FORCES.....................................................................................................................20 B) MOVEMENT OF AN AXLE-MOUNTED GEARBOX......................................................................................22





Page 3

1. INTRODUCTION The design of reliable, cost-effective gearing depends critically on :

• Accurate Gear Fatigue Strength Data

• Accurate Gear Load Data

• Accurate Gear Stress Analysis

In these notes the problem of determining accurate gear loads is considered.

In any gear system (fig.1 ) consisting of a motor, shafts, couplings, a gearbox and a driven machine, the loads to which the gearing is subjected are a function of the characteristics of both the driving and the driven machine, and of the couplings and shafts. In all but the simplest cases, the transmitted torques should be accurately calculated or measured, and an equivalent torque calculated. When this is not possible, an application factor KA, may be used as an approximate way of accounting for dynamic overloads in a particular drive system.

Figure 1.0 - Transmission System



Page 4

1.1 Design Load Based on Application Factor KA The torque being transmitted by a gearbox may be greater or smaller than the nominal power of the driving motor or engine. If it can be assumed that the motor has been correctly sized to supply the mean power required by the driven machine, the effect of dynamic overload can be allowed for by an application factor KA appropriate to the type of prime mover and the class of driven machine. KA for typical prime movers and a range of driven machines is shown in Table 1.0 below. Using an appropriate value of KA, the design torque TD is:

TD = TN . KA

Where TN is the nominal motor torque.

It should be noted that the application factor KA also makes allowance for the typical operation of the machine, that is the number of starts per hour and the utilisation of the machine.

Examples of Driven Machines with Different Working Characteristics

Application Factor KA for Different Prime Movers

Driven Machines

Cha

ract

er

Elec

tric

M

otor

Tu

rbin

e

Mul

ti C

ylin

der

Engi

ne

Sing

le

Cyl

inde

r En

gine

Generators, uniformly loaded belt or platform conveyors, worm conveyors, light elevators, packaging machines, feed gears for machine tools, ventilators,1ight centrifuges, centrifugal pumps, mixers for light fluids or constant density materials.

Uniform 1.05 1.25 1.5

Non-uniformly loaded belts or platform conveyors main drives of machine tools heavy elevators turning gears of cranes, industrial and mine ventilators, heavy centrifuges, centrifugal pumps, mixers for high viscosity or variable density materials, multi-cy1inder piston pumps, feed pumps, extruders (general), calenders, rotary furnaces, rolling mills (continuous zinc strip, aluminium strip as well as wire and bar rolling mills).

Moderate Shock

1.3 1.5 1.75

Extruders for rubber, mixers with interrupted operation for rubber and plastics, ball mills (light), wood working (mills, saws,1athes), billet rolling mills, lifting gear, single cylinder piston pumps.

Medium Shock

1.5 1.75 2

Excavators (bucket wheel gears, mufti-bucket gears, sieve gears, power shovels), ball mills (heavy), rubber dough mills, breaker (stone,0re) metallurgical machines, heavy feed pumps, rotary drilling apparatus, brick moulding presses, braking drums, peeling machines, cold strip rolling mills, briquette presses.

Heavy Shock

1.8 2.00 2.25

Table 1.0: Application Factor KA



Page 5

1.2 Design Load Based on Load Histogram The equivalent design load is best based on a measured or calculated torque-time characteristic, which takes account of all operating conditions, acceleration, variable speed running, braking and the system inertias. When the dynamic load characteristic, as shown in Fig. 2.a, has been determined, this can be reduced to a load histogram by calculating the time spent at each torque level T1 T2 ..Ti and running speed, n, and summing this for the life of the gearbox.

Figure 2.0 - a) Load-Time - Characteristic b) Load Histogram

To allow for different running speeds, the load histogram is best based on the total number of pinion revolutions ΔN1, ΔN2..ΔNi at each load level T1 T2 ..Ti for the design life of the gearbox, as shown in Fig. 2b.

The equivalent design torque, To, for a variable torque drive as in Fig. 2 can be calculated using the linear cumulative damage hypothesis of Palmgren/Miner. For a series of torque levels T1 T2 ..Ti occurring for load cycles ΔN1, ΔN2..ΔNi,the design torque TD for a design life ND <N ∞ is given by:



Page 6

TD =

q

D

qii

qq

NTNTNTN

1

2211 ...

∆+∆+∆

Equation 1.

where q is the slope of the Load-Life curve, [Note 1]. Cumulative damage theory assumes that there is no fatigue damage at stresses below the endurance limit (occurring at N ∞ ) so

that in the calculation of TD, applied torques at a level lower than T ∞ must not be considered in the calculation. Similarly, for a total design life greater than N ∞ , N ∞ must be in place of ND in Equation 1.

Failure load (% of load at endurance limit)

(a) Case Carburising Steel, Bending Failure (b) Direct Hardening Steels (inc. Induction Hrd.) Bending Failure (c) Direct and Case Hardening Steel, Contact Failure - No pitting (d) Direct and Case Hardening Steel, Contact Failure - Some pitting * N0 and N ∞ for (a)

Figure 3: - Idealised Load-Life Curves

Fig. 3 shows the typical Load-Life curves for bending failure and surface failure for wrought gear steels, through-, surface- or case-hardened. It should be noted that the nominal fatigue limit for surface strength (pitting) is 5.107 to 109 load cycles, while that for bending strength is only 3.105 to 3.106. The values of N0, N ∞ and q for different gear materials are given in Table 2 below.



Page 7

Note 1: For root bending stress, the slope of the S-N curve q* is identical to that of the Load-Life curve q. For contact stress, however, q* = 2.q

Gear Steel and and Heat

Treatment Bending Failure Surface Failure

No Pitting Some Pitting N0 N ∞ q N0' N ∞ ' q' N0'' N ∞ " q''

Thro-Hard Wrought Steel, CI Perlitic & Bainitic, Ductile Iron

104 3.106 6.22 105 5.107 6.6 6.105 109 7.9

Induction & Flame Hardened Steel

104 3.106 6.22 105 5.107 6.6 6.105 109 7.9

Case Carburised Steel 103 3.106 8.7 105 5.107 6.6 6.105 109 7.9 Gas and Plasma Nitrided, Nitriding - Steel, CI & Ductile Iron, Ferritic

103 3.106 17.0 105 2.106 5.7

Alloy Steels Bath and Nitro- Carburised

103 3.106 84.0 105 2.106 15.7

Table 2 : lndex "q" for Load



Page 8

2.0 On-Line Starting The torques generated in a drive system with conventional direct 'on-line' start AC motors are often misunderstood, resulting in both over - and under specification of gearboxes, and can result in very rapid gearbox failures.

The steady state starting characteristics of AC motors are well known, Fig. 4.

Figure 4:- Static Speed-torque Characteristics

At zero speed, a typical 3 phase squirrel cage motor will provide 70 - 80% of full load torque (FLT), with a peak pull-in torque of between 240% and 260% of FLT. At rated torque, the motor will operate at 2 to 4% slip.

What is less well known and understood is the transient dynamic behaviour of a squirrel cage motor in the first 1/10th of a second of an on-line start.

Figure 5 : - Ac Motor Transient Torque



Page 9

On start up, an AC motor generates a very high dynamic torque at 50 Hz, the transient component of which decays rapidly. The start-up torque is given by the expression

Mt = M1 + M2 e Tt−

sin(ω t +α ) (2)

Typical values for this expression, with the torque constants expressed in terms of FLT, are:

M1 = 0.7 ... 0.8 FLT M2 = 4 ... 6 FLT T = 0.03 sec. α = 0.17 ω = 314 (at 50 Hz)

The typical air-gap torque characteristics, with exponential decay of the transient torque M2, are shown in Fig. 5.

What is the significance of the transient and run-up torque characteristic for the specification and design of a mechanical transmission system, consisting of motor, couplings, gearbox, and driven machine or load?

Figure 6 : Motor, Gearbox, Driven Machine Arrangement

In a typical drive system as shown in Fig. 6, during acceleration from stand-still to normal motor running speed, typically 980, 1470 or 2950 rpm, the couplings and gearbox will transmit a load torque plus an acceleration torque, the acceleration torque being proportional to the inertia of the driven machine in relation to the total system inertia.

With:

Motor Inertia = mI

Load Inertia = I

Gear ratio = i Instantaneous motor 'air-gap' torque = mT

Load Torque = T

the acceleration torque transmitted by the gearbox is

IiIITTTT

mm +∗

∗−+= 2) )(

(3)



Page 10

The effect of gear ratio on maximum acceleration (and transient) torques is very significant, and often far more important than the actual inertia of the driven machine. Two typical drives are considered, both with a 100 KW motor with an armature inertia of lm = 2 Kg m2

2a Steel Furnace Drive Considering the gear drive which tilts a small steel furnace for pouring (Fig. 7). The whole vessel weighs about 500 Tonnes, with an inertia about the pouring axis of 2 x 106 Kg m2, and is designed to rotate 1800 in 1 min, i.e. at a maximum speed of' 0.5rpm. The overall gear ratio between 4 pole motor running at 1500 rpm and the furnace bull-ring is thus 3000:1

Figure 7: - Steel Furnace Pouring Drive

The maximum torque due to maximum operational out-of-balance of the furnace, referred to the motor, is 0.4 FLT. If no compliant couplings are used, and the torsional compliance of shaft and gears are neglected, the maximum gearbox torque at peak motor air gap torques, i.e. at

peak transient torque = 4.5 (FLT) and maximum pull-in torque = 2.5 (FLT)

can be checked using equation (3)

The peak transient torque is

( )FLTT

FLTT

81.0

)(1023000.2

102)4.05.4(4.0

1

62

6

1

=

∗+

∗∗−+=

The peak gearbox pull-in torque:

( ) ( )FLTT

FLTT

61.0

)(1023000.3

1024.05.24.0

2

62

6

2

=

∗+

∗∗−+=



Page 11

In this case, the peak transient torque is only 0.81 FLT and the peak pull-in torque is only 0.61 FLT. The nominal design torque of the gearbox therefore does NOT need to be based on the stall torque of the motor.

2b Compressor Drive Considering the simple speed-up drive for a centrifugal compressor as shown in Fig. 8. In this case, motor speed 1500 rpm, compressor speed 6000 rpm, and the drive characteristics are as follows:

Impeller inertia : 6 Kg m2

Motor inertia : 2 Kg m2

Gear Ratio : 1:4 Peak Transient AirgapTorque: 4.5 FLT Max. Airgap Pull-in Torque : 2.5 FLT Load torque at zero speed : 0.05 FLT Load torque at pull-in speed : 0.8 FLT

Fig 8: - Centrifugal Compressor Drive

Again assuming there are no compliant couplings in the drive, peak transient and pull-in torques can be calculated from (3)

Maximum transient torque:

( )

FLTT

FLTT

4.4

)(6

412

605.05.405.0

1

21

=

+

∗

∗−+=



Page 12

Maximum pull-in torque:

( )

FLTT

FLTT

46.2

)(6

412

608.05.208.0

2

22

=

+

∗

∗−+=

In this case, the full transient and pull-in torque would be imposed on gearbox and couplings, and steps should be taken to reduce the severity of gearbox loading, by either:

• fitting a suitable compliant coupling to absorb the transient torque peak, in which case the gearbox should still be designed to transmit a peak starting torque of 2.45 x FLT, or

• fitting some soft start device or a fluid coupling or a speed controller which would limit acceleration torque to about 1.5 x FLT.

Potential problems with coupling selection are considered in Section 5 below



Page 13

3. The Effect of Engine Characteristics In some cases IC engines can subject a geared drive to very large dynamic torques. In a micro-light engine, to minimise weight is not fitted and a flywheel, the engine relies on the inertia of the propeller to provide sufficient rotary inertia for satisfactory performance. The propeller load characteristic is similar to a 'fan' characteristic, e.g. at constant speed, torque is constant with time. The torque output of the single-cylinder two-stroke however, varies as shown in Fig. 9, ranging from -2 Tm (where Tm is the mean torque) during induction/compression to + 5 Tm during the power stroke. In this particular instance the propeller inertia lp is very large compared to the engine crankshaft inertia lc, even when the gear ratio i is taken into account.

Referred to the engine crankshaft,

pc Ii

I ∗< 2

1

Figure 9 : Schematic of Geared Propeller and Engine and Propeller Torque Characteristic

Although the propeller aerodynamic load torque is constant at Tm, because of the high propeller inertia the gearbox torque will be substantially the same as the dynamic engine torque and vary from + 5Tm to -2Tm. Some pinion (crankshaft) gear teeth will thus always see the peak torque + 5Tm, some the minimum torque -2Tm. In this rather exceptional case, by using an integer gear ratio, no teeth will be subjected to a reversing load of + 5Tm and -2Tm, and the gearing too can be designed for a design torque, TD = + 5Tm, .

With an engine speed of 5000 rpm, and a design life of 3000 hours, the crankshaft gear will be subjected to 1.1 x 109 load cycles, so that the gear must be designed for an infinite fatigue life (NL > N ) and a design torque TD = + 5Tm.



Page 14

This example illustrates a common phenomena, that complex duty cycles in apparently difficult dynamic systems, have a dominant feature which determines the design rating for the gears.



Page 15

4. Similar Systems In many cases apparently similar drive systems can subject a gearbox and shafts to very different dynamic overloads. Neglecting the effect of transient start-up torques, a simple geared fan drive as shown in Fig.10 subjects the gearbox to virtually no overload, and both motor and gear torque are constant.

Figure 10: - Fan Drive Figure 11: - Wind Turbine

A wind-turbine or aero generator would at first sight appear to be just an inversion of the simple fan, with the propeller driving an alternator. A typical Wind Turbine of 500 kW to 1 MU, as shown in Fig.11, will typically run at 50 rpm, with a 30:1 speed up box. With this high ratio, the alternator inertia referred to the motor is very high, and dynamic over torques of 2 to 3 fold can occur in gusting winds.



Page 16

5. Effect of Coupling Characteristics In section 2 it was shown that high transient torques are generated when a large rigidly coupled inertia is driven by an asynchronous motor with direct 'on-line' starting. The transient torques can be reduced if an appropriate compliant coupling is used. However, an inappropriate selection of coupling can actually increase the peak transient dynamic torque to which the gearbox is subjected. An example of this, on a large vacuum pump drive, was investigated in detail both experimentally and by simulation, with the results discussed below.

Drive specifcation:

Motor Power P = 450 KW,1500 rpm Inertia of Motor mI = 12.2 Kg m2

Transient Torques 1M = 0.7 FLT

2M = 4.3 FLT

Pump inertia I = 912 Kg m2

Gear Ratio i = 5.7:1 Load torque at 0 rpm T = 0

Without dynamic analysis the peak transient air gap torque will be

FLTMMTm 0.521 <+<

and, with this peak torque, the maximum transient gearbox torque which could occur with rigid couplings throughout the drive will be, from (3)

( )

FLTT

FLTT

48.39127.52.2

912050 2

=

+∗∗−+=



Page 17

Figure 12 : Coupling Characteristics in Pump Drive

Measurements on the actual drive fitted with couplings with the progressive stiffness characteristic'a' plotted in Fig.l2 showed that transient starting torques peaked at over 8 FLT.

To understand why the peak torque with a compliant coupling was more than two times greater than the peak torque without a coupling, the system run-up characteristics were simulated. Fig.13 shows the transient air-gap torque



Page 18

Figure 13 :Transient air gap torque

for the motor using (2), which peaks at 4.5 FLT. With this air-gap torque characteristic, the simulated gearbox torque is plotted in Fig.l4, curve 'a', which shows a peak torque of 7.9 FLT, corresponding well with the measurements.

Figure 14 : Transient Gearbox Torques

Substituting a stiff, linearly compliant coupling of 270 .103 Mn/rad (characteristic 'b' in Fig.12) for the very progressive coupling actually used would result in the start-up characteristics shown in curve 'b' in Fig.14. With this alternative 'stiffer coupling the peak transient torque is only 3.0 FLT, compared to 7.9 FLT for the highly progressive coupling, although the actual



Page 19

'wind-up' in the linear couplings is only .033 rad (1.80) compared to .06 rad (3.40) for the very progressive coupling.

This example clearly shows that the coupling stiffness characteristic can significantly affect the dynamic torques to which a gearbox is subjected. In high inertia drives, very progressive couplings with low initial stiffness and very high terminal stiffness should be avoided, since these allow the motor to initially accelerate rapidly, but then generate high 'impact' torques when the coupling suddenly stiffens up.



Page 20

6. Gear Overloads Due to Gearcase Mounting The majority of industrial gearboxes are bolted down rigidly to a chassis or foundation. The problems associated with foundation movement, and the resulting misalignment of gear shafts and gears are well known, and can lead to gear breakages. The best way of avoiding these problems is to use gearbox mountings which do not over-constrain the gearbox - for example simple three point mounts.

Shaft mounted gearboxes offer another, elegant way of avoiding foundation shortcomings. Unfortunately, these are not immune from other problems, and two examples of these are given to illustrate potential pitfalls.

a) Out-of-Balance Forces In a shaft-mounted gearbox, shown schematically in Fig.15, a high speed 160 KW electric motor running at 4000 rpm drives the gearbox through a shaft with two flexible couplings. To protect the gearbox from shock-loads, the torque reaction link uses soft bushes, giving a very low link stiffness.

Figure 15 : Schematic of Shaft Mounted Gearbox

Gear and bearing problems which occurred in service were investigated by straingauging the pinion shaft in the gearbox, which showed that high dynamic loads were being transmitted to the gearbox at pinion rotational frequency, as shown in Fig.16.



Page 21

Figure 16 : Dynamic Pinion Torque

In this particular case, the dynamic pinion torque, +/- 800 Nm peaking at 1600 Nm exceeded the nominal drive torque of the motor, 380 Nm, by more than a factor of 2 to 4 and severely overloaded the gears and bearings. Further investigation of the cause of this high dynamic torque at pinion rotational frequency revealed that it was caused by significant drive line out-of-balance, causing a pitching mode vibration of the gearcase as shown in Fig.l7.

Figure 17 : Pitching Mode Vibration of Gearcase

The driven machine in this case was of very high inertia, so that the pitching mode vibration of the gearbox around the shaft resulted in the electric traction motor being accelerated and decelerated at 66 Hz, leading to high dynamic torques being generated in the motor-pinion line. This particular problem, pitching vibration of a shaft mounted gearbox due to out-of balance, can be resolved by:

• reducing the out-of balance and hence the pitching torque about the axle.

• increasing the stiffness of the torque reaction link, to reduce pitching amplitude.



Page 22

• reducing the stiffness of the couplings between motor and gearbox.

b) Movement of an Axle-Mounted Gearbox A common application of 'shaft-mounted' gearboxes is in rail-traction in Electrical Multiple Units (EMU's). The arrangement is shown schematically in Fig.18.

Figure 18 : Arrangement of Axle Mounted Gearbox

The gearwheel is mounted directly on the axle, which also supports the gearcase including the pinion. The torque reaction from the gearbox is taken by a reaction link between gearbox and bogie. When the wheels run over a rail joint or similar feature, it moves on the primary suspension, that is there is relative vertical movement between bogie and axle which also results in a pitching movement of the gearcase about the axle. Measurements of the torque resulting from gearcase movement are shown in Fig.19, which show high dynamic torques, equal to motor stall torque, resulting from the wheel primary suspension movement.



Page 23

Figure 19 : Gearbox Torque Induced by Bogie Movement

In this case, the dynamic torque amplitude can be reduced by either:

• improving the de-coupling between gearbox and bogie by reducing the torque reaction link stiffness, or by

• better de-coupling between pinion and electric motor by reducing the coupling stiffness.

BGA Training Notes TN3: Worm Gears


Page 1

BGA Training Notes TN3

Worm Gears

Authors: Cedric Barber Dr. Mike Fish Holroyd Gear Works, Milnrow Rochdale, Lancashire, OL16 3LS [email protected] http://www.holroyd.renold.com



Page 2

Index

INTRODUCTION .................................................................................................................................................3

1.0 COMPILATION OF THE DESIGNATION...........................................................................................5

2.0 POWER/TORQUE RATING ...................................................................................................................9

THERMAL RATING ................................................................................................................................................9 MECHANICAL RATING ........................................................................................................................................10

3.0 EFFICIENCY...........................................................................................................................................16

4.0 BACKLASH .............................................................................................................................................18

5.0 LUBRICATION.......................................................................................................................................21

OIL BATH ...........................................................................................................................................................21 FORCE FEED .......................................................................................................................................................22 TYPES OF LUBRICANT.........................................................................................................................................22 VISCOSITY SELECTION........................................................................................................................................23 GREASE LUBRICATION .......................................................................................................................................24

6.0 DEFLECTION .........................................................................................................................................26

7.0 MESH CONTACT...................................................................................................................................29

8.0 FAULT TRACIING IN WORM GEAR APPLICATIONS.................................................................32

9.0 ALTERNATIVE WORMGEAR FORMS.............................................................................................37

10.0 CONTACT INSPECTION AND ANALYSIS .......................................................................................39





Page 3

INTRODUCTION During the 20th. Century there have been several methods available for the calculation of the power capacities and efficiencies of worm gears. These mainly emanating from engineers employed in gear manufacture and various universities.

In the United Kingdom a British Standard covering the subject was introduced which was the result of collaboration between interested parties, and had in part been outlined by Dr A E Merritt, at the time Assistant Chief Engineer David Brown and Sons, in his paper 'Worm Gear Performance' presented to the Institution of Mechanical Engineers in February 1935.

This Standard, B5721: 1937 Specification for Worm Gearing, became widely used in this country and was revised in 1963. This was replaced by a further edition in 1983 which is in effect a metricated version of the 1963 publication.

The BSS 721: 1963 and 1983 have been recognised internationally as being reasonably accurate in providing methods of assessing the wear, strength and efficiency, values of wormgears having involute helicoid form, although it is now recognised that the surface stress factors for phosphor bronze wormwheels are somewhat conservative and the Standard is restricted in the provision of these factors for the range of materials used for wormwheels.

In the United States the American Gear Manufacturers Association, AGMA, published their own standards, culminating in the current ANSI / AGMA 6034 - B92. Practice for Enclosed Cylindrical Wormgear Speed Reducers and Gearmotors. In company with this Standard AGMA also produce the ANSI/AGMA 6022-C93 Design Manual for Cylindrical Wormgearing, which contains a great deal of useful information relating to design procedures, tooth modification, recess and approach action, contact patterns, lubrication and worm bending, the last being necessary where relatively short working lines with high tooth loadings are required.

In Germany, another country with a relatively large number of wormgear manufacturers, there has not been a national standard to date, the majority of those procedures referring to the works of Niemann and Winter, who co-wrote the work Maschinenelemente in which can be found a method of rating wormgears and worm reduction gear units. This publication, with revisions and additions, formed the basis of a proposed standard DIN 3996, which has so far been produced only in draft form but has been adopted as the foundation of a new standard ISO 14521, currently under discussion and formulation.

It would be rather premature to consider in detail this proposed ISO 14521 at present due to the conflicting results which are being obtained from the procedures contained, these being somewhat at variance with the BS and AGMA experience. Hopefully these matters will be resolved in the next few months.

In this presentation therefore we will consider primarily the design and rating procedures specified in the BSS 721: 1983 Specification for Worm Gearing. This will be supplemented by comments concerning materials not covered by the Specification and the effect of various lubricants.



Page 4

We will also consider the gear mesh contact requirements of various applications and look at the conditions which can occur at the mesh in a worm and wormwheel in use and the assessment which can be made with these.

There will be a demonstration of software which, from the inputting of the design and manufacturing parameters, will produce a diagram for the resulting mesh contact.



Page 5

1.0 COMPILATION OF THE DESIGNATION Worm gearing forms a convenient means of connecting two non-intersecting shafts, usually at right angles to each other, and permits the adoption of a high ratio without undue difference in diameter between the worm and its mating wormwheel. It is the intention in this presentation to outline methods of design and capacities in accordance with the British Standard 721: Part 2: 1983 Worm Gearing with additional information with regard to materials, lubrication, and application.

We will be concerned with wormgears having an involute helicoid form, that is where the worm thread flanks are in the shape of an involute on a section at right angles to the axis of the gear, and in following the practice of defining the main gear features in the form of a designation. This information, together with the pressure angle and centre distance, is sufficient to enable all detail dimensions to be calculated from formulae or tables.

The basis of the involute helicoid as applied to a worm is illustrated in figure 1 where the involute curve originates at the base circle ABC in the transverse section, i.e. at 90º to the worm axis. The form is generated from the base lead angle and base circle.

Fig. 1

In the designation we use the idea of a limited series of basic forms of worm thread of unit module which would be applied to different pitches by using the principle of geometrical similarity. This is done by adopting for each basic form, and unit module, a nominal worm pitch diameter represented where possible by a whole number. This is called the 'diameter quotient', and denoted by q. Thus if a q = 10 for example, the nominal pitch diameter is 10 inches, or 10 millimetres, for unit module and is the same as that of a spur gear having 10 teeth of unit module. In general therefore, taking the worm and wheel together, and denoting



Page 6

the number of teeth in the wormwheel by z2, the pitch diameters and centre distances are the same as those of a pair of spur gears having q and z2 teeth, and the same pitch.

Fig. 2 - Determination of q for maximum efficiency for any worm speed

A method of determining values of q is contained in BS721 and reproduced here in the figures 2 and 3. Care has to be taken when using the lower values specified, since if the gear loads are relatively high and a close ratio is involved the bending stress and deflection in the worm may prove to be unacceptable. Also the base diameter of the worm must not impinge on the throat radius of the wormwheel otherwise undercutting of the worm thread will result. If either should be problematical the value of q can be increased until the situation is resolved.

Each value of q is, however, associated with the various possible values for the number of threads in the worm, denoted by z1. The shape of any worm, given a standardised system of detail dimensions, is then completely expressed by the designation written z1/q/m, where m is the module.

The designation for the worm and wheel together can then be written as z1/z2/q/m.

The number of threads in the worm, with the exception of traction gears, can be obtained

from the formula gR

z1a4.27 += taking the nearest whole number to the result. The

symbol a is the centre distance, Rg the ratio: z2 is then the next whole number below Rg z1



Page 7

except in the case of a single thread worm when z2 = Rg. z2 shall not be less than 17. This is done to avoid a situation where too small a proportion of the wheel teeth are in contact at any one time resulting in rough running. Conversely if more teeth than necessary are included relative to the centre distance a weak gear can result.

Fig. 3 - Determination of q for worm speeds up to 300r/min

One point to note is that as far as possible the value of T should not be equally divisible by z1 For example, a nominal ratio of 10/1 should not be interpreted as an actual 30/3 since if the wheel has to be cut with a single point cutter this will follow the path of one thread of the worm and so will cut 10 tooth spaces at 3 pitch divisions for each revolution of the wheel. It will then be necessary to index the wheel one pitch to cut a further 10 spaces and once again to cut the remaining 10. This results in a long machining time with the risk of indexing

errors and so should be avoided. The Module can be obtained from qz

m+

=2

x 2 a

To obtain the highest mechanical efficiency the diameter of the worm should be as small as possible consistent with the avoidance of excessive bending stress in the wormshaft, but the

lead angle must not exceed 45 degrees. Since the lead angle is obtained from tan qz1

you

will see that, with a relatively close ratio, once the ratio qz1

is equal to a value greater than 1

an increase in diameter is necessary. Therefore q must then be made equal to z1.



Page 8

There are however other factors which may ultimately dictate the selection of q. An example is where the worm is bored and so this factor may dictate the eventual diameter of the root of the worm threads and a q value will have to be selected which ensures there is sufficient material beneath the threads to the bore or keyway. Once the designation is established it is then possible to carry out the calculations necessary to check the torque rating.



Page 9

2.0 POWER/TORQUE RATING There are generally two factors that may limit the power capacity of an enclosed gear drive. The first is concerned with the tooth pressure and the resistance to the various types of wear or failure associated with direct tooth load. The second is governed by the temperature of the lubricant reaching the engaging teeth. If the temperature is too high, the oil film may fail at moderate tooth pressures with the result that the tooth surfaces are rapidly destroyed.

Thermal Rating

As the ratio of a worm gear increases so does the value of the efficiency reduce. Since the heat generated is proportional to the power loss the ultimate power transmission capability of the drive may be limited by temperature rise than by any other factor with a gear in continuous operation, although the development of various lubricants in recent years has much reduced this difference. The power lost is taken partly by overcoming friction between the mating surfaces of the worm threads and wormwheel teeth, partly through friction within the bearings and oilseals, and oil churning losses. The whole of this loss is converted into heat and the heat so generated is transmitted to the gear casing.

Whilst the frictional loss is approximately constant in a gear transmitting uniform power the oil churning losses will reduce as the heat build up takes place and the viscosity of the lubricant becomes lower. The graph shown in figure 4 illustrates the results of tests conducted on a 7" centre gear set in this connection.

Fig. 4



Page 10

The rate of heat dissipation will be dependent upon the design of the gearcase and the surrounding structure and it is due to this and the variables involved that it is not currently practicable to introduce a definitive method of calculating the thermal rating for a wormgear drive. There is no guidance with regard to the thermal aspects in the British Standard.

Mechanical Rating

Wear Rating is defined as the resistance of the working surfaces of the worm and wheel teeth to physical wear due to the working loads imposed by the drive. Obviously the wear rating will be dependent upon the combination of materials used for the gears and the conditions under which they operate, particularly with regard to lubrication and surface finish. In addition to these the wear rating also depends upon the operational speed, wheel pitch diameter, tooth pitch, and the relation of the number of wheel teeth to worm threads.

The Strength Rating can be defined as the resistance to physical failure by fatigue of the gear teeth due to the loading conditions imposed by the drive. Generally the Strength Rating is well in excess of the Wear Rating, the difference reducing as the ratio increases.

The Wear and Strength Ratings are frequently referred to as the Mechanical Ratings to differentiate from the Thermal Rating.

Unlike the Thermal Rating the Mechanical Ratings are based on the working lives of the gears which is usually specified by the machine designer. The British Standard Specification 721:1983 assumes a basic working life of 26,000 hours, but the application of life factors to the selection power or torque allows a variation from this figure to virtual zero up to 100,000 hours.

Whilst in calculation the facility is given for a transmitted power to zero life it is dangerous to assume figures approaching this can be relied upon. It is therefore suggested that working lives should not be admitted below 1,000 hours.

In order to asses the Wear and Strength Ratings it is necessary to calculate the values relating to both the worm and the wormwheel, the lowest of the four then being taken as the ruling capacity. The following formulae estimate these values assuming a working life requirement under constant load and speed of 26,000 hours.

Permissible load for wear.

For the normal rating the permissible torque, M, (in N•m) on the wormwheel shall be limited by wear to the lower of the following values:

Worm - 0.00191Xc,1 σcm,1 Zd 21.8m

Wormwheel - 0.00191Xc,2 σcm,2 Zd 21.8m

where

d2 is the reference circle diameter of the wormwheel (in mm);

m is the axial module (in mm);

Xc,1 and Xc,2 are the speed factors for wear on the worm and wormwheel respectively;



Page 11

σcm,1 and σcm,2 are the surface stress factors on the worm and wormwheel respectively;

Z is the zone factor as given in the BS. This value is based on a wormwheel effective face width of 12 +qm but this can be increased proportionately to 13.2 +qm .

The speed factors correspond to the combination of rotational speed and rubbing or sliding speed. The rubbing speed is given by:

Vs = 0.0000524d1n1 sec γ

or Vs = 0.0000524mn1 √(z12 + q2)

in metres/second

The surface stress and zone factors are taken from charts contained in the BS 721, figures 5 and 6 respectively.

Permissible load for strength.

For the normal rating the permissible torque, M, (in N•m) on the wormwheel shall be limited by strength to the lower of the following values:

Worm - 0.0018Xb,1 σbm,1 mIf,2 d2 cos γ

Wormwheel - 0.0018Xb,2 σbm,2 mIf,2 d2 cos γ

where

Xb,1 and Xb,2 are the speed factors for strength on the worm and wormwheel respectively;

σbm,1 and σbm,2 are the bending stress factors on the worm and wormwheel respectively;

If,2 is the length of root of the wormwheel teeth (in mm);

d2 is the wormwheel reference circle diameter (in mm);

m is the axial module (in mm);

γ is the lead angle of worm thread.

The speed factors correspond to the rotational speed only and can be obtained from a chart contained in the BS. The bending stress factors also are given in that Standard.

The length of the root in the wormwheel tooth If2 = (da1 + 2c) sin-1

+ cdb

1a

e

2 where the

angle is in radians.

The ruling capacity is the lowest value obtained in the calculations relating to wear and strength.



Page 12

Fig. 5

The range of Surface and Bending stress factors contained in BSS 721, Table 7 figure 7 is rather limited in the range of materials covered failing to include worms produced in oil hardened or surface flame hardened materials. This is unfortunate since there are many applications where materials having superior surface stress qualities over normalised materials can be utilised which do not require the higher cost involved in gas carburised/case hardening.

The specification also lists only phosphor/tin bronze to BS 1400 PB2 and cast iron, whereas nickel bronze to DIN 1705 GZ-CuSn12Ni is widely used in Germany in many applications where it offers advantages over phosphor bronze, and aluminium bronze to BS 1400 AB-2 can be used in certain cases where relatively high torques have to be transmitted but where the rubbing speed does not exceed 1.3 metres/sec.

The draft standard ISO 14521, based on draft DIN 3996, proposes values for nickel bronze to DIN GZ-CuSn12Ni, 1½% nickel, and aluminium bronze to DIN GZ-CuA110Ni, equivalent to BS 1400 AB-2, which would be equivalent to σcm - 18.6 and σcm - 23.6 respectively when used in conjunction with a casehardened worm. The equivalent bending stress factors would be 76.7 and 95.8 respectively.



Page 13

Fig. 6

The surface stress values in both the BS and proposed ISO standards are based on lubrication with a mineral oil. If it is intended that the drive will be permanently lubricated with a synthesised hydrocarbon oil (SHC) or polyalphaolefin (PAO), the σcm values can be increased by 10% or polyalkyleneglycol (PAG) 25%. There are continuing improvements taking place in the field of lubricants which can be used for wormgearing and the designer should maintain an interest in this area of development.

If it is required to select a gear with an operating life other than 26,000 hours the selection power or torque can be amended by reference to the chart 'Total equivalent running time and related life factors for wear and strength' figure 8. Against the required life in hours can be obtained, either for wear or strength, the required life factor and the selection power or torque should be divided by this factor, or the calculated capacity multiplied by it if it is required to establish the resultant life.

Where the speed or load requirement, or both, can vary during the life of the gears and the cycle can be anticipated at the design stage the effect of this can be determined by calculation in terms of the equivalent time on full load.



Page 14

Fig. 7

The calculation is shown where a maximum torque M1 is acting for a period H1, at a mean speed n2.1, with smaller torques M2, M3, acting for periods H2, H3, at mean speeds n2.2, n2.3 etc.

Hc = H1 + H2 )()(

2.1

2.2

nn

1

2

MM 3 + H3

)()(

2.1

2.3

nn

1

3

MM 3 + etc

The total equivalent running time for wear at torque M1 and wormwheel speed n2.1 is then given by:

Hec = Hc x number of complete cycles expected during the life of the gears.



Page 15

Similarly the equivalent running time per cycle for strength can be calculated. In both cases the gears should then have a capacity equal to or greater than the required M1 at speeds n2.1 for the equivalent time.

In the event of a check having to be made against a stall condition, or a similar situation where a short period overload can occur, this can be made against the 'Momentary overload capacity'.

This states:

Wear - 0.00382 σcm,2 Zd21.8m

Strength - 0.004 σbm,2 If,2 d2m cos γ

This assumes the torque equal to the lower of the above will not exceed 15 seconds in duration.

It should be stated in this situation that the wear value must not be exceeded. There are occasions where it may be suggested that since a load can apply for only a very short period the strength value only needs to be considered but it should be noted that if the strength rating is higher than the wear and the wear value is exceeded irreparable surface damage can occur such that the operational life is severely reduced.



Page 16

3.0 EFFICIENCY The efficiency of a worm drive depends on the lead angle of the worm threads, and on the coefficient of friction. The coefficient of friction between a steel worm and a bronze wheel falls as the surface sliding speed increases. At very slow speed or under starting conditions, the coefficient may be as high as 0.135, falling quickly as the speed increases, and eventually reaching a minimum value of about 0.02 at worm pitch line speeds of 12 metres/second or over. The efficiency of a worm gear will therefore increase as the speed increases.

Values relating to the coefficient of friction relative to the rubbing speed, assuming a casehardened and ground worm meshing with a phosphor bronze wormwheel and lubricated with a mineral oil having a viscosity between 60 cSt and 130 cSt at 60ºC, can be obtained from the BS, reproduced here as figure 8.

Fig. 8



Page 17

The values are then substituted in the formulae:

Worm driving tan γ tan (γ + φ) Wormwheel driving tan (γ - φ) tan γ where tan φ is the coefficient of friction and Tan γ + tan φ

tan (γ + φ) = 1 - tan γ.tan φ

tan γ - tan φ

tan (γ - φ) = 1 + tan γ.tanφ

If the ‘wormwheel driving’ efficiency is less than zero at 0 rpm the gear will theoretically not backdrive from the static situation. This feature can be influenced by other factors however and if it is required that the gear should not backdrive from the static then a brake or holdback should be incorporated.

At some point in the speed range any ratio of wormgear will eventually attain a theoretically positive backdriving efficiency and at this point it is possible that in some applications the load could attempt to drive or accelerate the prime mover. Again it would be necessary to anticipate this condition.

The coefficient of friction is influenced by the lubricant and, as earlier stated, the values published in the BS721 are based on mineral oil. If a polyalphaolefin lubricant is to be utilised for an application a value equivalent to 83% of the value obtained can be used, or with a polyglycol 75%. These values are approximate and for guidance only.



Page 18

4.0 BACKLASH The backlash of a set of gears is the minimum distance between the non-working flanks of a set of gears when the working flanks are in contact. It is usually measured at the pitch line of a gear when this component is mounted at correct centres to its mating component, which is held rigid, and an attempt made to rotate in each direction the gear against which the measuring instrument is located.

With a wormgear set it is most convenient to hold the worm whilst 'rocking' the wormwheel with a dial gauge indicator located normal to a tooth.

Depending upon the bearing arrangement it is sometimes necessary to incorporate a small amount of axial clearance in the wormline bearing assembly and so when checking the backlash alone the worm should be restrained axially. Until this clearance has been taken up by an increase in the temperature of the components when the drive is in use it will add to the initial clearance which includes the backlash in the gears.

The bronze wormwheel will also expand a little as the temperature increases and this can lead to reduction in backlash. When the amount of backlash required in a gear set is specified this factor should be taken into account.

For a worm meshing with a bronze wheel in a cast iron or steel case, the backlash (in mm) between worm and wormwheel is reduced by approximately

3.42m (z2 + 7) (T – 20)

106

T is the working temperature of the oil in ºC.

Where bronze or aluminium cases are used, the reduction in backlash will be less than the above.

With certain exceptions, the amount of backlash allowed in a pair of gears is normally of very little importance in its effect on their performance and satisfactory running, since most gear drives are usually of such a nature that the driving faces of the teeth are continually in contact due to the driving pressure.

In spite of conditions of very high speed running it is seldom found that the driving faces of the teeth separate even momentarily, and provided that there is sufficient backlash for the oil film the actual amount of backlash does not matter. It is therefore wrong to assume that by cutting down the amount of backlash the gears will run more smoothly.

If, from an applicational point of view, reduced backlash tolerances are required, it is preferable that these gears are manufactured to produce the required amount at the correct centre distance.

As mentioned earlier, the amount of backlash introduced is usually irrelevant, but in instances where there are torque reversals, such as in a cam or crank drive, this must be controlled. Similar conditions apply in a passenger lift where the load is nearly balanced and the effects of acceleration and braking may be to change the direction of applied torque on the gear.



Page 19

In order to maintain the reduced backlash condition the rate of wear in the wormgears must also be anticipated, and so in those cases where it applies, the initial selection must be a generous one.

There are instances when even this method is not adequate and the actual backlash must be closely controlled throughout the life of the machine in which the wormgears are used. Applications where this is necessary include the table drives to rotary and indexing tables, and the table drives in hobbing machines. On these occasions wormgears incorporating the 'dual lead' or 'duplex' feature are a much favoured option.

Fig. 9

Fig. 10



Page 20

Dual lead or duplex wormgears are, as their title suggest, constructed with two leads. One flank of the worm thread and its mating sides of the wheel teeth are manufactured with one lead, and the other side and its mating teeth to a slightly different lead. So far as the worm is concerned this has the effect of producing a worm thread which increases in thickness from one end of the worm to the other. If the worm is therefore moved in an axial direction when in mesh with the wormwheel backlash in the set can be adjusted to the required amount. Diagrammatically this is shown in figure 9.

The assembly must allow axial adjustment of the wormshaft which can usually be achieved without a great deal of dismantling to achieve access. The figure 10 shows an example of how this can be achieved.

The advantages of this system are:

a) The centre distance is fixed and the accuracies of pitch and form produced and checked in the position in which the components eventually operate.

b) The torque capacity of a dual lead gear set is equal to that of a single lead wormgear although, as would frequently be the case with a reduced backlash gear, the service factor is usually a generous one in order to keep the rate of wear to a minimum.

c) When the wheel is initially bedded to the worm any further adjustment does not bring any change to the contact condition and so neither the accuracy nor the efficiency deteriorates as a result of this operation.



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5.0 LUBRICATION In contrast with spur, helical and bevel gearing, in which the frictional properties of the lubricant are not in general of first importance, worm gears present conditions in which such properties virtually govern the performance of the gears. The losses caused by friction between the worm and its mating wheel are often of greater magnitude than those due to churning and bearing friction, and will account for a high proportion of the temperature rise. The tooth losses, which increase as the ratio of the gears increase, are proportional under specific conditions to the coefficient of friction.

As with other types of gear tooth friction falls but churning losses rise with an increase in oil viscosity or operating speeds. In some cases the net effect of an increase in viscosity grade would be a net increase in operating temperature, and where this is so the viscosity grade has to be limited on that account. In other cases the net effect of an increase in viscosity grade would be a reduction in operating temperature, but a limit to the viscosity grade that can be used may be imposed by thickening of the oil when cold. If this goes too far the demand on the starting torque may be excessive and there is a danger of oil starvation to the teeth.

Oil Bath

With wormgears relying on oil bath lubrication it is usually the case that having stood for some time in a static environment the oil will attain a temperature approximating to the ambient with an appropriate viscosity. As the drive is taken up and power transmitted the resulting friction will cause the oil to increase in temperature and the viscosity consequently to reduce.

There are various methods available which provide a means of selecting the correct viscosity for an application, of which more later, and in all cases this is done at the assumed working temperature. Lubrication of the gears and possibly the associated bearings would then be assured through agitation with the resultant circulation also aiding cooling.

This however necessitates either the worm, wormwheel, or both, being partly submerged sufficiently to ensure that the lubricant is directed to the mesh where, due to the rotating action and dependent upon the speeds, it will be thrown outwards to the associated components. The static level therefore should be at such a position that these requirements are met but should be such that the actual mesh is not submerged otherwise when driving cavitation may result causing overheating and consequential damage to the gears.

There is a limitation on the capability of an oil bath or static sump system since in certain applications the rotational speeds can be too high to facilitate collecting the lubricant in sufficient quantity to guarantee an adequate supply to the mesh. This limit is usually imposed at a rubbing speed of 12.8 metres/second for involute form wormgears of good quality.

At the higher sliding speeds or where the gears are positioned such that lubrication from an oil bath is not practicable it may be necessary to supply the oil by pumping through piping.



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Force Feed

Whilst adding to the cost the provision of a forced feed system should ensure a controlled viscosity in all situations and the avoidance of drag during start-up.

Ideally piping should be provided which directs the oil to the mesh, the pipe being 'fishtailed' at the outset to provide a wide jet of fluid aimed at the mesh, primarily to the side of the worm at which it enters the wheel and to the wheel as it enters the worm, see figure 11.

Fig. 11

With a system of this type it is likely the lubricant will feed other components but will eventually return to a main reservoir where it will be cooled. The extent of cooling taking place will have to be known since, as with the static sump arrangement, the required viscosity at the mesh will be that at the point of supply but with the cooling induced it is likely that the initial viscosity will be lower than is necessary for the oil bath arrangement.

The amount of lubricant required for the wormgears only can vary dependent upon the speeds and transmitted power but as a general guide the following may be useful.

In litres/minute = Centre distance in millimeters 22

Types of Lubricant

Straight mineral oils are frequently used in industrial worm gear drives and although inferior to some fatty oils in coefficient of friction lubricants of this type are preferred because of their superior stability towards oxidation and associated chemical changes. Generally these oils have an operational ceiling temperature in the region of 100ºC but consistent use near this leads to accelerated deterioration in performance.

However many of the power ratings published by the manufacturers of wormgear units today assume an operating sump temperature of 100ºC, where the temperatures at the mesh can



Page 23

be 110 -120ºC. In these instances alternative lubricants are proposed which may be of the type containing polyalphaolefin (PAO) base fluids, or those based upon polyglycol's (PG).

Both of these alternatives are frequently termed 'synthetics' but it should be noted that the PAO lubricants can generally be used in machines where the oils seals, sealants, and paints are compatible with mineral oils, but a number of PG lubricants are not compatible with these and may require Viton seals, silicone sealants, and epoxy based paints.

These lubricants usually have a lower pour point than mineral oils of an equivalent viscosity and a much higher operating range. Also the change in viscosity relative to temperature is less and so it is sometimes possible to select a lower viscosity PAO or PG lubricant in certain applications which reduces the effect of oil churning losses at start-up.

The reduction in the coefficient of friction, referred to in the section of this presentation dealing with efficiency, results in a lowering of the heat generated at the mesh and facilitates an increase in power throughput. When taking advantage of this it should be recognised that whilst the performance of the wormgears is enhanced by the adoption, that of the bearings is not to the same extent. This point is of importance particularly where a wormgear drive may have been specified initially for lubrication by a mineral oil but a change is being made for some reason.

It is advisable not to use lubricants with sulphur phosphorus or sulphur chlorine additives in wormgear drives since at elevated temperatures these can have a detrimental effect on phosphor bronze.

Aluminium bronze is more resistant to these components but care has to be used when selecting PG type lubricants where the wormwheel is produced from this material.

Viscosity Selection

There are several procedural methods published for the selection of the viscosity for wormgear drives, many of which are issued by the producers of the various lubricants, and one of the less complex of these, published by the Mobil Oil Company, is shown in figure 12 which can be used knowing the transmitted power, speed, centre distance and nominal ratio. The method used in arriving at the required value is shown on the figure.

If the working temperature is not known the viscosity value can be taken as equating to that published by the respective producer at 100º C. Where the ambient temperature is above 30ºC move up one grade.

Lubricants are usually graded by their viscosity at 40º C and usually the producers of these lubricants use in their designation of the product the nominal viscosity at this temperature. In their publications dealing with specific lubricants they show the viscosities at 40 and 100º C, the pour points, the viscosity index, and several other factors dealing with the chemical and physical attributes. There will also be information in the text indicating the types of application for which the product is suitable.

It is not advisable to assume the oil is sufficiently fluid at the pour point for use in a gear drive, it is safer to assume the limit is 5º C higher.

If the variations in ambient temperature are such that, for example, these may periodically be below the pour point plus 5º C it will be necessary to consider:

a) A heater in the sump or reservoir.



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b) Changing the oil with the season, i.e. summer/winter.

c) Select an PAO or PG which has a power point 5º C below the lowest ambient but will yet have an adequate viscosity at the higher.

Grease Lubrication

In many wormgear drives, the preference is for grease rather than oil. This is usually due to the sealing being considered inadequate to keep a fluid in the gearcase. Those greases intended specifically for the lubrication of bearings are not suitable for wormgear use, being too stiff in consistency.

Bearing greases are usually to Grade 1, 2 or 3 of the NLGI classification. Grade 0 greases can be used in certain spur and helical gear drives but for worm gearing a grease to Grade 00 will be necessary. Even then care has to be taken in the selection since the thermal conductivity of a grease will be inferior to that of a fluid, it retains wear debris, churning losses tend to be high, and the limit on sliding speed is much lower.

Some of these disadvantages can be reduced by using a grease which has been specifically developed for use in wormgearing comprising a synthetic fluid incorporated in a semi-fluid grease formation. As a general guide mineral oil based greases should not be selected for sliding speeds in excess of 2.5 metres/second (500 feet/minute) and synthetic based greases 7 metres/second (1400 feet/minute).



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Fig 12



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6.0 DEFLECTION When a torque is applied to the worm, the tendency is for a displacement of the wheel in a combined sideways and tilting movement. Figure 13A illustrates the case where the bearings are not preloaded, a worm torque in the direction of the arrow causing the wheel to move its central plane from AA to A'A'. If the bearings are preloaded the amount of deflection is reduced and the effect in 13B is obtained, this now being the result of distortion in the housing and possibly in the wormshaft.

If the factors referred to in A and B are not fully resolved the net result can be as illustrated in figure 13C.

Fig. 13

It is due to this tendency to deflect under load, inherent in wormgearing, which necessitates a close control of the bearing settings in the wheeline and preferably for a preload to be incorporated where possible, so reducing deflection to a minimum.

Its magnitude depends on the general design of the assembly and on the amount of the side thrust exerted by the worm on the wheel. For a given wormwheel torque the side thrust depends on the ratio and on the diameter of the worm relative to the centre distance. Worms having a high lead angle will obviously cause more deflection than those of a low lead angle.

In figure 14A is shown the contact zone on a wormwheel tooth and in figure 14B the possible result of deflection δ. This has caused a smaller area of contact concentrated at the entry side.

It will be noted that the direction of wheel deflection relative to the worm is always in the direction in which the contacting part of the worm threads is moving, and the resulting concentration of contact is therefore always at the entering side of the wheel tooth. This applies to both right and left hand thread worms irrespective of direction of rotation. Apart from the concentration of load this deflection condition prevents the entry of lubricant



Page 27

between the surfaces, there is heavy initial wear at the entry side with the risk of cracking in the surface of a case hardened worm, and there is a tendency to be noisy at high speeds.

Fig. 14

A further result of deflection from true gearing positions is that the uniformity of angular velocity transmission is affected. Worm gearing may be considered as an infinite number of rack sections through the worm in planes normal to the wheel axis, each rack being odontically conjugate to its corresponding section of the worm wheel. If the wheel is displaced axially relative to the worm, the action is no longer conjugate, because the profile of the rack changes progressively and is different at each section. On the entering side of the tooth, the rack section is of a high obliquity (it may be as much as 45 deg.) and on the leaving side it is of low obliquity (often approaching zero). The deviation from uniform angular velocity can therefore be appreciable, particularly on worms of high lead angle when the change in tooth section is greater.

If the existence of some deflection is taken as inevitable the problem of allowing for deflection is mainly one of cutting the wheel in such a manner that this slight difference in curvature of the wheel teeth is produced in the first place so that heavy initial wear is avoided, the oil freely enters the tooth faces, and uniformity of angular velocity is maintained as near as possible in any of the deflected positions of the wheel.

There are occasions when, in an existing application, the amount and direction of deflection is known, or it may be practicable to remove the gear set from the drive and assemble this into a wormgear inspection machine where the on-load contact condition can be simulated. The adjustments necessary to achieve this from the design centre and positioning can be measured. With this information it is then possible to roughly estimate the amount of modification necessary to accommodate the misalignment and wormgearing having involute helicoid thread form lend themselves particularly to a contact modification of this nature.



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The Overdriving Situation

In cases where the wheel has at times to drive the worm the direction of relative deflection with the worm driving is opposite to that with the wheel driving for a given direction of rotation. For example, it can be seen in figure 15 that for clockwise direction of rotation of the worm the wheel deflects in direction A when the worm is driving, and direction B when the wheel is driving.

Fig. 15

It is clear therefore that the more usual leaving side contact condition which is so essential to the majority of wormgear application can be detrimental in a case such as this where, in the overdriving situation, that area of contact is now at the entry side.

This being so it is necessary to specify a modification to the tooth flanks of the wheel which provides for the resulting deflection and still ensures that lubricant can enter the mesh in both the worm driving and overdriving conditions.



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7.0 MESH CONTACT An instantaneous contact between a position on a worm thread and that corresponding on the wormwheel tooth flank is in the form of a line. As rotation takes place and that same point on the worm traverses the flank of the wormwheel lines of contact are developed as shown in figure 16. In this diagram we see that the positions 1, 2, and 3 on the worm traverse the lines of the same numbers on the wheel tooth flank.

Fig. 16

The total area of contact is the result of other points adjacent to those numbered and cover the projected area over the effective face width and the working depth when the gear is fully bedded-in. In the 'as manufactured' condition however it would be undesirable to achieve this situation since it is not practicable to accurately anticipate the distortions resulting from deflection occurring under load in the gearcase, bearings and support shafts.

It is therefore necessary to create a situation which makes allowance for a certain amount of deflection whilst providing the facility for the ingress of lubricant to the mesh in the deflected condition. This is done by designing and setting the cutter in the wormwheel gear generation machining operation such that an area of contact is achieved towards the bearing side of the tooth so providing a gap or allowance at the entry face, the amount of which is predetermined. The cross-hatched area in figure 17A indicates this initial contact zone with the on load condition figure 17B.

There are occasions however when the wheel drives the worm for periods other than what might be expected during a relatively infrequent machine stopping operation, for example, in a lift or elevator where the gear will spend a considerable amount of its working life overdriving, or in a speed increasing situation where the wheel perpetually drives the worm.

In applications such as this it is advisable to alter the gear cutting procedures so that the contact zone on the wormwheel is now centrally disposed, as illustrated in figure 18 which provides for the ingress of lubricant when either component is the driver.



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Fig. 17

Transferring the contact zone from the bearing side to a more central position effectively reduces the amount of useable flank on the wheel and also shortens the length of contact on the worm. When selecting wormgears incorporating this feature therefore it is sensible to de-rate from the more usual capacity relative to the 26,000 hour rating procedure earlier described.

Fig. 18



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The incorporation of worm gears into the screw actuation to the rollers of a rolling mill in a steel works calls for different techniques to those previously described. In this situation the rotational speeds can be quite low, frequently of the order of 1 rpm of the wormwheel with extremely high tooth loads when rolling is taking place.

Off load the tooth pressures are a fraction of those on load and since it is not practicable to arrange for a graduated escalation in the transmitted torque to facilitate bedding, it is frequently necessary to cut the teeth with an 'as machined' contact pattern covering about 60-75% of the face yet achieving an adequate allowance at the entry for the ingress of lubricant.

Since in most instances the gearcase is exceptionally robust, the wheeline bearings of the sleeve type, and the worm of a relatively low lead angle, with a substantial 'q' value, deflection is relatively small and so provided close controls are exercised in design and manufacture the gears will last for several years.



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8.0 FAULT TRACIING IN WORM GEAR APPLICATIONS The majority of worm gear sets used in industrial gear drives comprise a case hardened steel worm, ground on the thread profiles, and a wormwheel having the teeth generated in a centrifugally cast phosphor bronze rim fitted to a cast iron centre. Since the wormwheel is manufactured from the softer substance it is usually this component which first discloses various conditions that may lead to a shortening in life. .

Under certain conditions deterioration of the working surfaces of the wheel teeth, worm flanks, or breakage of a tooth, might occur and the following information will provide useful basic knowledge if you are called upon to identify the possible cause of any trouble which may be attributed to a worm gear set.

The Wormwheel

A condition which may become apparent during the early life of the gears is that of pitting. Initial pitting can appear on the faces of the wheel teeth during the running-in period and is usually attributed to local over stressing caused by high spots on the teeth. The effect is a form of fatigue failure of the bronze surface, starting with minute cracks which gradually enlarge until a small piece drops out. The cavities so caused may be quite deep

This early or initial pitting is to be expected concentrated towards the leaving side of the wheel tooth flank. The cross hatching in Figure 19 illustrates the initial contact pattern on the wormwheel teeth when correctly assembled.

Fig. 19

The pressure on the tooth flank is not uniform, the greater being in the initial contact area during the bedding in process and it is in this zone in which the initial pitting to be expected, as illustrated in figure 20. This also shows the spread of contact which can be expected when load is applied, this now spreading towards the entry side due to deflection of the teeth, wormshaft, bearings, and housing.



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Fig. 20

Eventually the surface of the gear will become bedded to the worm and the pitting will stop.

Sometimes gears which have been in use for a considerable-period of time develop pitting that similarly comes to a halt. For example, bearing wear may cause slight changes of alignment accompanied by local over stressing of the teeth, which is in due course reduced by corrective wear. Pitting beginning gradually and showing no tendency to rapid increase is likely to be of this temporary kind.

In relatively few cases the pitting does not halt but continues to increase, the pits growing in depth or total area, and this is the kind of pitting usually described as progressive. Unfortunately in the initial stages of pitting it is difficult to foretell whether it will develop into the arrested or progressive type.

Under these circumstances and until the situation has stabilised a regular inspection procedure has to be followed. The relative age of the pits can be deduced by observation, the initial variety by this time being blackened and perhaps partially smoothed over, the new being bright coloured and unstained. If, from these new pits, there are light surface cracks or discolouration alongside, then further pitting is likely. The spread of pitting towards the entry side is shown in figure 21.

A combination of wear and pitting is the most serious form of tooth failure and a . careful watch should be kept if this is suspected. The pitting is not necessarily serious in itself, but if associated with a high rate of wear the tooth of the wheel may be very quickly destroyed.

As with other types of gearing the correct selection, and proper maintenance in good condition, of the lubricant is of considerable importance. If this has been kept clean, and the tooth pressures have not been excessive, the mating surface of both worm and wheel should be quite smooth.



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Fig. 21

Unclean lubricant or deterioration in the condition of the oil due to excessive heat or aging can cause the formation of ridges across the face of the teeth resulting in localised contact on the peak of these ridges which, when high tooth pressures are possible, can bring about a breakdown of the oil film at these points and rapid wear or abrasion then occurs.

Fig. 22

Damage of this type is shown in figure 22 where the score marks can be seen. Where this feature is apparent and to the degree illustrated there will also be damage to the thread flanks of the worm. If this is discovered in time the process may be arrested somewhat by draining the existing lubricant, flushing with a light oil under no load conditions, and re-filling with clean oil having a higher viscosity than the original. Before using a heavier grade of



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lubricant however the suitability of the resulting viscosity must be checked against the requirements of the assembly and associated components, and the rubbing speed.

A further alternative lubricant could be a synthetically based type with a higher film strength and lower friction properties than the original.

If a wormwheel has been incorrectly positioned to the worm such that the contact under load moves towards the entry side so reducing the deflection or oil entry allowance the lubricant is unable to reach the area under pressure in sufficient quantity and so again metal to metal contact is likely with resultant overheating, scoring of both worm and wheel, and rapid wear.

Fracture of teeth can occur either through bending fatigue, impact, or overload. The usual form of fracture starts with small cracks at the roots of the teeth, extending down into the rim section.

The Worm

A casehardened and ground worm will usually outlast the life of a wormwheel by a comfortable margin, but there are occasions when failure occurs, and this may happen in one of the following ways.

Dirt held in suspension in the lubricant can become embedded in the surface of the wheel tooth flanks, and due to the action of the gears will act as a lapping medium, progressively wearing away the driving face of the worm thread. Once the case has been worn away the process will be accelerated.

Overloading or excessive shock loads, particularly if the mounting of the gear is not rigid, can result in lifting of the case on the thread flanks. Minute cracks appear on the flanks which spread due to surface stress, and flaking then occurs. When these particles break away from the thread, small crevices are left which rapidly cut into the softer bronze of the wheel, and may indeed wear or cut away the teeth very quickly.

Fig. 23

The cracks caused by shock loads are random in the contact area as illustrated in figure 23 Cracks can also occur during prolonged overload which is of sufficient duration to cause overheating, examples are shown in figure 24 These are radial in pattern and tend to occur chiefly in the area towards the crest of the thread flanks where the wheel contact is



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concentrated at this point. Cracks of this form can also be found when overheating is due to incorrect or insufficient lubricant.

A further result of overheating is that the worm threads will be discoloured around the area of contact with, in some instances, a coating of carbonised lubricant adhering to the surface.

Fig. 24

Fractures of the worm can occur at the root, which is often the smallest diameter, and also at a mid position between the bearings. On the occasions when this does happen it is usually the result by considerable overload or shock loading accompanied by excessive deflection of the shaft.

The gearbox and its mountings should always be rigid otherwise excessive deflection may take place, this usually being in the form of sideways tilting of the wheel, which causes heavy contact on the entering edge. The rate of wear would increase as a result since the leading edge of the teeth then acts as a scraper, removing lubricant from the worm thread flanks before it reaches the point of maximum pressure.



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9.0 ALTERNATIVE WORMGEAR FORMS

We have so far considered the parallel type worm with involute helicoid thread form, a configuration which is more widely manufactured than any other. This, to a great extent, is due to the fact that the involute thread is a ruled surface built up from an infinite number of straight lines so that the form can be easily checked by a simple straight line method, and the involute helicoid surface can be generated in a variety of ways. It can be ground by a plane grinding wheel set at predetermined angles without resort to any trial and error methods, the set-up of the machine being relatively simple, the resulting accuracy being dependent on only two machine settings, mainly the lead and the inclination of the grinding wheel.

The hob can be similarly generated to pre-determined sizes with the gashes inserted in the normal way

The form in the worm is therefore convex in shape, the degree of curvature being dependent upon the lead angle and pressure angle.

Fig. 25

This type of wormgear is designated as the ZI form in DIN 3975 and another form of profile which involves a parallel worm in the ZC in which the worm flank is concave in the axial section as shown in figure 25. It is claimed that this form of gear, where a concave surface engages with one convex in shape, results in lower Hertzian pressure than is the case with the ZI and improved lubricant retention through the mesh, the direction of sliding being basically parallel to the lines of contact. The conclusion is that these features make possible a greater load transmission capability than is possible with the involute helicoid.

A further variation in the gear sometimes referred to as the ‘all encircling’ type of worm has threads formed on the surface which follows the pitch circle of the wormwheel, that is the worm partly encircles the wheel as illustrated in figure 26.

A higher power transmission capability than is practicable with the parallel involute worm is claimed for this gear due to the increased number of engagements.



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Fig.26



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10.0 Contact inspection and analysis The geometry of a worm wheel tooth surface is notoriously complex making direct inspection difficult. Flank contact must be sufficient to cope with the surface stress of the operating load but allow relief between the flanks which avoids tooth edge contacts and permits adequate lubrication during initial bedding-in of the gear set. Coating one component in marking blue ink and meshing the components together is a crude but quick check that the appropriate contact conditions have been achieved. A typical example of the contact marking on an involute helicoid wheel tooth after the test is illustrated in Figure 27

.

“CENTRAL” CONTACT POSITION

Fig. 27

The exact relief can be determined by inserting feeler gauges during the meshing action to determine the minimum clearance. Changes in the relationship between relief, area, and position can be made to subsequent gear set manufacture by a qualified inspector through adjustments of worm grinding or wheel hobbing machine settings. The position and size of the contact area for a given application is largely defined by experience. Historically the marking blue test has proved worthwhile as it is practical to carry out on a production line basis and often gives sufficient information for assessing the current machine settings. However, the method is subject to slight variations in technique potentially giving inconsistent results. Further, compensation for the conditions of tooling and machinery at the time of production is only possible after component manufacture.

Improvement in computing speed has enabled the development of highly detailed mathematical models replicating component geometry and surface interaction during mesh. New co-ordinate measuring machine (CMM) technology has permitted the development of more precise and repeatable measurement. Changes in tool geometry over a working life span can be modelled and compensated for in machine settings prior to component manufacture. Calculated tooth surface co-ordinates using digital measurements of actual tool forms can confirm accurate surface generation. The increasing use of CMM inspection will lead to far greater consistency in future production.

The relief on the tooth surface created by the tool design permits small additional wheel rotations relative to the nominal angular rotation. This represents an error in positioning and is referred to as “transmission error”. The error value (in microns or arc seconds of wheel

ENTRY RELIEF EXIT RELIEF

CONTACT AREA

WORM ROTATION



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displacement) can be plotted against actual wheel rotation over a full revolution to give a graph of total potential accuracy error. Figure 28A is a single flank transmission error test taken from a manufactured worm gear set.

Fig. 28A

Fig. 28B

Figure 28B shows 4-5 tooth engagements in more detail. Form the curve it can be seen that there are several frequency contributions to the wave form. If recorded digitally, the wave form can be analysed by mathematically reducing the signal using a Fast Fourier Transform (FFT). This represents the total wave form as a sum of pure sine waves of increasing frequency. Plotting each sine wave amplitude against frequency creates a characteristic spectrum as in Figure 29 for the complete wave.

Fig. 29



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By examining intervals of cycles per wheel rotation frequency this graph can be used to diagnose and correct error sources caused by the design or manufacturing process. The second derivative of the wave form gives acceleration which can be associated with a transmitted force and operating speed. This can be used to determine whether the gear set will contribute to vibration when connected to a system with known fundamental frequencies.

The transmission error wave form will change as the transmitted load changes due to tooth bending and elastic deformation of the surface. As a consequence of this, a gear system may display a differing frequency spectrum through a range of loads. Designers of spur and helical gearing use this fact to reduce noise and vibration levels by applying a specific profile to flanks at no load which will minimise transmission error due to deformation at the design load. The same principle can be applied to worm gearing to improve positioning accuracy and reduce vibration.

BGA Training Notes TN4-1: Gear Cutting Tools - Part 1


Page 1

BGA Training Notes TN4 -1

Gear Cutting Tools - Part 1

Author: W Clark Stocksmoor Engineering 10 Shepley Road Stocksmoor Huddersfield HD4 6XW Email: [email protected] Photographs and CAD data courtesy of: David Brown Textron Power Transmission http://www.davidbrown.com



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Index

GEAR CUTTING TOOLS - PART 1 ..................................................................................................................3

GEARS. ..................................................................................................................................................................3

GEAR PRODUCTION..........................................................................................................................................6 HOBS. ...................................................................................................................................................................6

Monobloc Ground Form..................................................................................................................................6 Monobloc Unground Form..............................................................................................................................6 Inserted Blade..................................................................................................................................................6 Hob Design Features. ......................................................................................................................................7

SHAPER CUTTER DESIGN....................................................................................................................................12 External Gears...............................................................................................................................................13 Internal Gears................................................................................................................................................13 Basic Rack. ....................................................................................................................................................15 Quality ...........................................................................................................................................................16

SHAVING CUTTERS. ............................................................................................................................................17 Types. .............................................................................................................................................................17 Helix Angle. ...................................................................................................................................................18 Diameter / Number of Teeth. .........................................................................................................................18 Width..............................................................................................................................................................19 Serrations.......................................................................................................................................................19 Clearance grooves. ........................................................................................................................................19 Design Life.....................................................................................................................................................20 Quality ...........................................................................................................................................................21





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GEAR CUTTING TOOLS - PART 1 1st Slide ~ Introduction.

Gears. 2nd Slide ~ Selection of Power Transmission Units.

In General the purpose of a gear train is to transmit power through rotary motion from one shaft to another. In most cases it is also a requirement that the motion be uniform hence the aim to provide mating gear tooth profiles which are conjugate when the shafts operate at varying centre distances and under varying loads. The goal when designing Gear Tools is to become ever closer to achieving this condition in conjunction with producing an ever more efficient cutting tool.

There are an almost infinite variety of forms which can be used as gear tooth profiles, however the Involute profile is the one most commonly used due to its unique qualities and the simplicity in its theory and production capabilities.

3rd Slide ~ The Involute Curve.

Although the Involute profile is the one most commonly used for gear tooth forms there are occasions when others can be used to advantage, i.e. Cycloidal, Conformal etc.

4th Slide ~ Conformal Gear Pair.

5th Slide ~ Interesting Quotations.

However for the purpose of this presentation we shall consider only gear teeth with Involute profiles. The gear teeth can be either Spur or Helical and designed to operate on either parallel or crossed axes. Worm and Bevel Gears have been omitted from this presentation, each being a complete subject in their own right and time does not permit a useful coverage of these subjects

It is common practice to identify gears by reference data, namely Pitch and Pressure Angle. These elements are referenced to a particular diameter normally termed the Pitch Diameter.

6th Slide ~ Representation of Pitch.

The Pitch is normally defined in one or more of three ways listed as follows: -

1) Diametrical Pitch ~ number of teeth per inch/mm of pitch diameter. (usually applied when using the Imperial system of measurement)

2) Circular Pitch ~ the arc distance between similar and adjacent tooth profiles measured on the pitch diameter.

2) Module ~ the amount of pitch diameter per tooth (usually applied when using the Metric system of measurement)



Page 4

7th Slide ~ Representation of Pressure Angle.

The Pressure Angle is usually defined as follows: -

A line tangent to the cylinder from which the Involute curve is generated (termed the base cylinder) cuts the pitch circle at a point termed the pitch point. The angle formed by a radial line to the point of tangency and a radial line passing through the pitch point is termed the Pressure Angle.

Whilst this method of categorising gears is extremely useful the use of the terms pitch, pitch diameter and pressure angle can be very misleading particularly to not so well informed Gear Engineers. It is worth noting that a gear as an infinite number of pitches, pitch diameters and pressure angles within clearly defined limits of the base and blank diameters, the pitch and pressure angle having a definite relationship at any chosen diameter.

8th Slide ~ Representation of Base Pitch.

The important element is Base Pitch which may be defined as the arc distance between similar and adjacent tooth profiles measured on the base cylinder circumference.

The pitch and pressure angle at any diameter can then be calculated from this base pitch.

Base Diameter = Transverse Base Pitch * Number of Teeth / Pi

9th Slide ~ Representation of Diameter Relationships.

Then at any diameter ‘d’

a) Transverse Pressure Angle = Arc Cosine [ Base Diameter / Diameter(d) ] b) Transverse Circular Pitch = Transverse Base Pitch / Cosine Transverse Pressure

Angle

10th Slide ~ Development of Gear Pitch Cylinder.

In the case of Helical Gears we must consider three planes namely: -

1) Axial 2) Normal 3) Transverse

Then:

Pressure Angle: The pressure angle in each plane varies with diameter.

Pitch: Axial Plane - is constant, does not vary with diameter and is fundamental.

Normal Plane – varies with diameter and the Normal Base Pitch is fundamental.

Transverse Plane – varies with diameter.

11th Slide ~ Representation of Axial Pitch.



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For any Involute Gear the Axial Pitch and the Normal Base Pitch are fundamental in defining the Gear

Normally the drawing of the gear will contain detailed dimensions including outside, tip chamfer (where applicable), root, and base diameters together with tooth thickness etc. In addition it is common practice to define a number of diameters on the gear pertaining to mating gear(s) and tools.

The most common of these can be listed as follows: -

12th Slide ~ Table of Gear Data.

13th Slide ~ Definition of Important Diameters on Gear Profile.

a) End of Active Profile Diameter (EAPD) ~ this would normally be the Outside Diameter or where applicable the start of Chamfer Diameter.

b) End of Tip Modification Diameter (ETMD) ~ this would normally be the EAPD but not always. This is the diameter where the tip modification is a maximum.

c) Start of Tip Modification Diameter (STMD) ~ this is the diameter where the theoretical tooth profile starts to depart from the true involute form of the main profile when moving in a direction towards the tip of the tooth.

d) Operating Pitch Diameter (OPD) ~ this is calculated with reference to a mating gear or gears operating at a defined centre distance between shafts.

e) Start of Root Modification Diameter (SRMD) ~ this is the diameter where the theoretical tooth profile starts to depart from the true involute form of the main profile when moving in a direction towards the root of the tooth.

f) Start of Active Profile Diameter (SAPD) ~ this is normally the diameter which meshes with the EAPD of the mating gear.

g) End of Root Modification Diameter (ERMD) ~ this would normally be the SAPD, but not always, and is the diameter where the root modification is a maximum.

h) Crossover Diameter (COD) ~ this is the diameter where the undercut trochoid, intentionally formed by a pre-finishing tool, cuts the theoretical finished tooth profile.

i) Shaved to Diameter (SHTD) ~ this is the diameter where the tip of the Shaving Cutter tooth would mesh on the theoretical finished tooth profile. This is often used to define lowest point of theoretical contact with a Gear Roll, Grinding Wheel or other finishing tools.

The position of these diameters together with the amounts of profile modification need to be measured so they are normally detailed on the working drawing in the form of a Profile Control Chart.

14th Slide ~ Specimen Profile Control Chart.



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Gear Production. 15th Slide ~ Methods of Gear Manufacture.

We will all be aware there are many ways of producing gears using a variety of tools both singularly or in conjunction with each other. (i.e. Hobbing, Hobbing / Shaving, Hobbing / Rolling, Hobbing / Grinding etc. and the Hob can be replaced by Shaper Cutter, Rack Cutter. Rotary Cutter etc.)

However taking cognisance of time available and the relative usage of particular types of tools this presentation will be restricted to Hob, Shaping Cutter and Shaving Cutter design.

16th Slide ~ Picture of Hob.

Hobs.

17th Slide ~ Hob Types.

Over the years hobs have been manufactured in a number of forms, the most popular being

a) Monobloc Ground Form b) Monobloc Unground Form c) Inserted Blade

MONOBLOC GROUND FORM

This was the original concept for hobs used in volume production. These were used successfully for many years, however in many cases they failed to meet the machine production capabilities, the hobs failing when run at high speeds and feeds. This was mainly due to the difficulty in achieving ideal relief angles when using a grinding wheel to finish grind the tooth profiles.

MONOBLOC UNGROUND FORM

Ground form hobs were used in large quantities in automobile plants in conjunction with a shaving operation. It was realised that some quality in the hob could be sacrificed in the interests of improving production. This led to the development of the Accurate Unground Form or Black Hob where ideal Relief Angles could be achieved albeit with some reduction in quality. This type of hob dominated the market in the USA for many years but failed to achieve the same success in Europe.

INSERTED BLADE

The Inserted Blade hob was first developed in Germany during WW II due to scarcity of High Speed Steel. The body of the tool being manufactured from standard steel with high speed steel inserts to form the cutting edges. The blades were manufactured in special bodies enabling ideal relief angles to be produced before assembly in the tool body. These hobs took a long time to gain popularity but from the late sixties on dominated the market place.

With the advent of PVD Ceramic Coatings (TiN, TiCN etc.) and high quality High Speed Steel Substrates (both conventional and powder metallurgy), optimum relief angles are



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achievable so the whole picture has changed and in recent years the Monobloc Ground Form Hob once again dominates the market.

Development of machine tools has enabled the hobs to be produced to high levels of accuracy with optimised relief angles. By using relatively small diameter grinding wheels optimum life is also ensured.

HOB DESIGN FEATURES.

18th Slide ~ Hob Design Features.

19th Slide ~ Hob Design Features ~ Length, O. Dia. Collars/ Hubs.

Outside Diameter ~ the diameter will normally have been decided at the quotation stage where a number of factors is taken into account namely: - - national standards. - company Standards. - machine tool capacity - machine tool fixturing limitations and fouling points on the gear

blank. - production requirements. - available substrate blank sizes. - pitch (depth of tooth). - bore size ~ where specified.

Length~ the length will normally have been decided at the quotation stage where a number of factors are taken into account. - national standards. - company standards. - minimum lengths for complete generation of product gear. - machine tool capacity (amount of available shift etc.) - machine tool fixturing limitations (fouling etc.). - production requirements. - collar / hub dimensions.

Collars (Hubs) ~ a number of factors are taken into account when determining the collar / hub diameter and length. - national standards. - company standards. - type of driving arrangement. - diameter at bottom of gashes

Bore ~ the bore will normally have been decided at the quotation stage where a number of factors are taken into account. - national standards. - company standards. - hob diameter - machine tool (available arbors – type and sizes) - type of drive (conventional keyway in bore or end driving slots)

Note: - Hobs may be manufactured solid on arbor particularly in the case of small diameter tools

20th Slide ~ Hob Design Features ~ Bores & Means of Driving.



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Means of Driving ~ the means of driving will normally have been decided at the quotation stage where a number of factors are taken into account. - national standards. - company standards. - machine tool (availability and type of arbors) - difference in bore diameter and diameter at root of gashes

Hobs may be driven in a number of ways including keyway in the bore, enddriving slots, grippers etc.

Most hobs are designed to be driven through an axial keyway in the bore unless the outside diameter is restricted, resulting in insufficient material between the root of the gashes and the bore to accommodate a keyway without seriously weakening the tool.

Where this occurs the hob would be provided with end driving slots however in some cases the hob outside diameter is to small to accommodate a bore and is then made integral with a driving arbor, this type of hob is often referred to as a shank hob.

Many different methods have been used to locate and drive a hob particularly to facilitate automatic loading of the hob. However the ones noted cover the majority of cases.

Number of Threads

21st, 22nd and 23rd Slides ~ Effect of Changing Number of Threads ~ Hobs.

In the case of hobs used to finish cut gears it is normal practice to use single thread hobs in the interest of accuracy. (smoothness of tooth profile etc.).

In the case of hobs used for Pre finishing operations (Preshave, Pregrind, Preskive etc.) then high production rates become very important and considerable advantage is to be gained by using multi-thread hobs. However it is generally recommended that their use should be restricted in line with the following parameters: -

a) the number of teeth in gear should not be less than about 15 per thread of hob. in the case of a two thread hob the teeth in gear should not be less than 31 in the case of a three-thread hob the teeth in gear should not be less than 47

b) the number of teeth in the gear should not be divisible by the number of threads in the hob. If a 2-thread hob, produced to normal standards of accuracy is used to cut a 40 toothed gear there is a good chance that adjacent pitch errors will occur (the same thread cuts the same alternate teeth through out the cutting cycle). However in the interests of production efficiency it is sometimes considered essential to use a multi-thread hob where the number of teeth in gear is divisible by the number of threads in the hob. In this case the tolerance on the thread divide error has to be reduced and this may increase the cost of the tool.

Note: -

Whilst Tool Designers and Manufacturers have to pay cognisance to the fact that machine tools used both to manufacture hobs, and use them will vary in age, and state of technology incorporated in them. Using the latest technology and best quality machine tools divide problems are not an issue.



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Hand of Threads.

The hand of threads would normally be decided at the quotation stage. A right hand hob has a thread helix in the form of a right hand screw, to avoid ambiguity it is common practice to define the direction of rotation.

It is normal practice to manufacture the hob the same hand as the gear, ensuring the cutting forces are directed against the backlash in the machine drive. However the development of advanced technology machine tools as enabled hobs of the opposite hand to be used successfully. There is good evidence to suggest that using hobs of the opposite hand to the gear increase productivity.

24th Slide ~ Gashes ~ effect of Number Variation ~ Hobs.

Gashes. ~ a number of factors are taken into account • national standards. • company standards. • availability of Index Plates used for sharpening the hob

(particularly at the customers sharpening facility). • type of hob ~ Standard Gash or Multi-Gash (Optimised Gash). • customers preference ~

usually influenced by customers requirement for either a high production tool or one where life is the overriding requirement.

The number of gashes is normally decided at the design stage.

25th Slide ~ Relief Produced Using Abrasive Wheel / Tool ~ Hobs.

Cam Drop ~ this controls the infeed of the form tool and the grinding wheel during the relieving operations which in turn control the top and side relief angles and is a feature of the design programs.

26th Slide ~ Relief Produced Using Tool also Illustration of End of Useful Life ~ Hobs.

27th Slide ~ Substrate Materials ~ Composition.

28th Slide ~ Substrate Materials ~ Performance.

29th Slide ~ Substrate Materials ~ Constituent Effects.

30th Slide ~ Substrate Materials ~ Carbide.

Substrate Material ~ the substrate material would normally be decided at the quotation stage where a number of factors are taken into account. It will normally be Conventional or Powder Metallurgy High Speed Steel, or Tungsten Carbide.

Whilst Powder Metallurgy H.S.S. is used extensively for Hobs the development of P.M.H.S.S. containing Ceramic is producing some interesting results.

- customer preference.



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- gear material (type, machineability etc.) - production requirements (speeds, feeds etc.)

Extensive information on material properties have been compiled in databases and these or used to optimise Tool Substrate Material selection.

31st Slide ~ Coating Effects.

Coatings.

Ceramic coatings are now applied to almost all cutting tools, in the case of hardened and ground tools the PVD system is used. TiN coatings are maybe the most widely used although many different coatings have, and are being developed to suit particular applications.

The type of coating would normally be decided at the quotation stage where a number of factors are taken into account. Customer Preference. Gear Material (Type, Machineability etc.) Production requirements (Speeds, Feeds etc.)

Extensive information on Coating Properties have been compiled in databases and these or used to optimise Tool Coating selection.

32nd Slide ~ Hob Basic Rack.

Basic Rack.

When establishing the form of the Basic Rack the designer uses the gear data including any required profile modifications. He must also consider, where applicable, the Shaving Cutter or any other means of finishing the gear form.

He is free to vary the pressure angle, usually within wide limits providing he adjusts the pitch accordingly. The Base Pitch of the Basic Rack must be identical to the Normal Base Pitch of the Gear

Unless a deliberate error is introduced to produce an out of pressure angle condition, which is sometimes used to vary finishing stock, particularly on ground gears.

The choice of Pressure Angle will be influenced by the need to produce a particular form in the root of the gear. He will normally maintain the largest tip radius (Basic Rack) possible with a ‘full’ radius as maximum. He will also aim to maintain sufficient tip protuberance on the basic rack to produce the required undercut on shaved gears or gears finished by other means. These ideals must be the target without destroying active profile. At all times the designer will be seeking to keep the pressure angle as high as possible for a more efficient cutting tool.

In the cases where protuberance is used the tip area of the basic rack provides a fertile ground for the creative tool designer in the quest to achieve the ideal gear root form. However the form consisting of a constantly varying protuberance running into a constant protuberance which blends with the tip radius which in turn blends with the tip is used in the majority of cases.



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33rd Slide ~ CAD Output (Hob 1)

34th Slide CAD Output (Hob 2)

35th Slide CAD Output (TRR2)

36h Slide CAD Output (TRR2)

Computer programs are in use to carry out the design optimisation where the output consists of complete design and manufacturing data together with accurate layouts of the tooth form. It is worth noting that in the case of gears which are subject to a finishing operation a decision has to be made as to what root form is required at the hobbing stage.

1) should the undercut be below the finished form leaving the root trochoid untouched by the finishing tool. It is common practice in the USA and England to provide this type of undercut whilst in Continental Europe the tendency is to provide little or no undercut a step in the root trochoid being acceptable.

2) in the case of gears which are to be full form ground the root trochoid will need to closely follow the finished form to minimise localised burning caused by unequal finishing stock in this area.

37th, 38 th , 39th, 40th and 41st Slides ~ Sample Hob Quality Standard.

Quality

Hobs are normally manufactured to International, National, or Company Standards, the relevant quality being agreed at the quotation stage.

Modern computer programs enable correlation between the desired gear quality and the hob quality. However it is important to recognise that hobbing machine and fixturing quality together with blank accuracy have a major influence on the quality of gear produced.

42nd Slide ~ The Modern Hob Features.

Modern Hob

Considerable evidence is being amassed to confirm that the hob of the future will have the following features: -

1) Diameter ~ the diameter to be made as small as possible even to the extent of using end driving slots in place of a normal keyway in the bore or in many cases solid on arbor.

- improved cutting action. - reduced approach and leaving times. - increased cutting efficiency i.e. hob can run faster for same

peripheral speed as a larger diameter hob. - reduced hob run out errors.

2) High length to diameter ratio.

- makes maximum use of automatic hob shift (i.e. less downtime through tool changes.)

- more cost effective tool.



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3) Multi-Thread.

- increased cutting efficiency.

4) Optimised number of Gashes

- reduces tooth loading allowing increased speed / feed.

5) Superior Substrate Material.

- powder metallurgy high speed steel has superior metallurgical properties to conventionally produced high speed steels. However in many applications Tungsten Carbide is now being used to advantage and in addition the use of Powder Metallurgy high speed steel containing ceramic is being developed with varying, but encouraging results. It is important that the correct balance between speed and feed is observed at all times.

6) Ceramic Coated surface. - improves wear resistance - increases cutting efficiency

7) Hand of hob.

- there is some evidence to suggest that when using the latest hobbing machines the use of opposite hand hobs together with conventional direction hobbing (not climb) leads to increased production rates.

43rd, 44th, 45th and 46th Slide ~ Hob Optimisation Printout.

Shaper Cutter Design

47th Slide ~ Picture of Shaping Cutter.

48th Slide ~ GSC Design Features (List).

49th and 50th Slides ~ GSC Design Features.

Number of Teeth ~ the number of teeth in cutter will be decided at the quotation stage where a number factors are considered.

1) various standards 2) capacity of Shaping Machine. 3) fouling points on the Gear Blank. 4) maximum diameter before interference (trimming etc.) takes place

when cutting Internal Gears. 5) in the case of Helical Gears care has to be taken to ensure a

suitable guide is available. This restriction may not apply on all machine tools as some have variable guides and on the latest machines the helix is controlled electronically. Where the helix produced is controlled by a conventional guide the lead of cutter



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must equal the lead of guide this in turn will decide the number of teeth in cutter and hence the nominal diameter.

51st Slide ~ Helical Guide.

52nd Slide ~ GSC Design Features (Forms)

Diameter ~ the nominal diameter is a function of the number of teeth whilst the final outside diameter is decided at the design stage and is a feature of computer program output. In the case of internal gears the number of teeth in cutter may have to be adjusted to ensure interference does not occur.

Width/Length ~ Two widths/lengths must be considered: -

a) the effective width ( the toothed portion) b) the overall width/length.

EXTERNAL GEARS.

In addition to the width/length the overall form must be established. Shaper Cutters are normally used to cut external gears where due to adjacent fouling point’s etc. a hob cannot be used. The effective width is normally decided from standards and based on pitch but subject to amendment to ensure clearance with adjacent fouling points when the cutter is in operation. In general cutters are made in the form of a disc with a counter bore to facilitate sharpening. However in many instances, particularly where there is only a small clearance between the end of the gear tooth and an adjacent face it is necessary to provide a deep counter bore to accommodate the arbor nut.

INTERNAL GEARS.

Shaper Cutters are a very popular means of cutting Internal Gears. When used to cut relatively large gears the above conditions apply, however for use in cutting relatively small diameter gears the cutters are normally produced with a shank or some form of hub. The overall length is then dependant on the facewidth of the component and any fouling points, which decide the proximity of the cutting head to the face of the component.

Bore ~ he bore will normally have been decided at the quotation stage where a number of factors are taken into account.

- national standards. - company Standards. - cutter diameter - machine tool (available arbors – type and sizes)

Shank/Hub ~ the type of shank/hub will normally have been decided at the quotation stage where a number of factors is taken into account.

- national standards. - company Standards. - cutter diameter - machine tool (available arbors – type and sizes) - limitations due to fouling points etc.



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53rd and 54th Slides ~ Sharpening ~ Shapers.

Type of Sharpening ~ in the case of spur shaper cutters these are normally ‘dish’ ground giving a front rake angle in the order of 5 degrees. In the case of helical gears there are two generally accepted ways of sharpening.

1) Normal ~ this is the usual method where the cutting face is ground normal to the nominal helix angle with a front rake angle in the order of 5 degrees. However in some cases, to improve cutting efficiency the cutting face is ground not normal to the nominal helix angle but substantially so.

2) Nicked and Chamfered ~ this type of sharpening is normally used where the run out of the cutting face is impeded (double helical gears without gap etc.). With this type of sharpening the front face is first ground flat then following the contour of the cutting edges the acute side is chamfered and the obtuse side nicked.

Relief Angles. Side Relief ~ values for side relief angles have been established and

normally lie between two and three degrees.

Top Relief ~ the value for top relief angle is dependent on the value of the side relief angle and the pressure angle of the basic rack.

In many instances compromises have to be made to obtain the best balance between optimum values for top relief and side relief angles for a given pressure angle of basic rack. This is a feature of the computer program output.

55th Slide Substrate Materials and Coatings ~ Shapers.

Substrate Material ~ the substrate material would normally be decided at the quotation stage where a number of factors are taken into account. It will normally be Conventional or Powder Metallurgy High Speed Steel. However whilst Powder Metallurgy H.S.S. is used extensively for Gear Shaper Cutters the development of P.M.H.S.S. containing Ceramic is producing some interesting results.

- customer preference. - gear material (type, machineability etc.) - production requirements (speeds, feeds etc.)

Extensive information on Material Properties have been compiled in databases and these or used to optimise Tool Substrate Material selection.

Coatings ~ Ceramic coatings are now applied to almost all cutting tools, in the case of hardened and ground tools the PVD system is used. TiN coatings are maybe the most widely used although many different coatings have, and are being developed to suit particular applications.

The type of coating would normally be decided at the quotation stage where a number of factors are taken into account.



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- customer preference. - gear material (type, machineability etc.) - production requirements (speeds, feeds etc.)

Extensive information on Coating Properties have been compiled in databases and these or used to optimise Tool Coating selection.

BASIC RACK.

56th and 57th Slides ~ Through Grinding.~ Shapers.

It is assumed that the Cutters will be finish ground on a ‘through ground’ type of machine where the form of the abrasive wheel takes the form of the basic rack space. The centre distance between the abrasive wheel and the cutter is progressively reduced and increased as the abrasive wheel reciprocates backwards and forwards across the cutter face as the cutter is rolled in strict relation to the wheel form.

The centre distance is at a minimum at the back of the cutter and the action produces the desired top and side relief angles.

58th Slide ~ GSC Basic Rack ~ Shapers.

When establishing the form of the Basic Rack the designer uses the gear data including any required profile modifications. He must also consider, where applicable, the Shaving Cutter or any other means of finishing the gear form.

He is free to vary the pressure angle, usually within wide limits providing he adjusts the pitch accordingly. The Base Pitch of the Basic Rack must be identical to the Normal Base Pitch of the Gear

Unless a deliberate error is introduced to produce an out of pressure angle condition, which is sometimes used to vary finishing stock, particularly on ground gears.

The choice of Pressure Angle will be influenced by the need to produce a particular form in the root of the gear. He will normally maintain the largest tip radius (Basic Rack) possible with a ‘full’ radius as maximum. He will also aim to maintain sufficient tip protuberance on the basic rack to produce the required undercut on shaved gears or gears finished by other means.

These ideals must be the target without destroying active profile. At all times the designer will be seeking to keep the pressure angle as high as possible for a more efficient tool.

Having decided the Basic Rack form this is the used to generate the cutter profiles using the process known generally as the Through Ground process. Whilst the form produced is accurate and predictable

There is an inherent problem similar to the problem of the varying backlash condition met when running a pair of gears at non-standard centres. In the case of Shaper Cutters it results in a varying depth of tooth being cut as the tool is sharpened, this of course also affects chamfer ramp and modification diameters.

59th, 60th ,61st , 62nd ,63rd, 64th, 65th, and 66th Slides ~ CAD Output ~ Shapers.



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Computer programs are in use to carry out the design optimisation including positioning the nominal design section so that the varying depth of tooth cut is the same at the beginning and end of tool life. The output consists of complete design and manufacturing data together with accurate layouts of the tooth form produced by three cutter sections, front, nominal and life.

It is worth noting that in the case of gears which are subject to a finishing operation a decision has to be made as to what root form is required at the shaping stage.

a) should the undercut be below the finished form leaving the root trochoid untouched by the finishing tool. It is common practice in the USA and England to provide this type of undercut whilst in Continental Europe the tendency is to provide little or no undercut a step in the root trochoid being acceptable.

b) in the case of gears which are to be full form ground the root trochoid will need to closely follow the finished form to minimise localised burning caused by unequal finishing stock in this area.

QUALITY

67th and 68th Slides ~ Sample Standards ~ Shapers.

Shaper Cutters are normally manufactured to International, National, or Company Standards, the relevant quality being agreed at the quotation stage.

Modern computer programs enable correlation between the desired gear quality and the cutter quality. However it is important to recognise that shaping machine and fixturing quality together with blank quality have a major influence on the quality of gear produced.

69th and 70th Slides ~ Optimisation Output ~ Shapers

71st Slide ~ Picture of Shaving Cutter.



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Shaving Cutters.

A shaving cutter is very similar to a gear but with undercuts at the root of the teeth and serrations across the face of the teeth and is designed to operate at a crossed axis angle when meshing with the product gear. It removes fine slivers of material from the faces of the gear teeth by the relative sliding motion between the cutter and gear tooth faces due to the crossed axes angle.

72nd Slide ~ Design Features ~ Shavers.

TYPES.

There are essentially 4 methods of close mesh shaving where the gear and shaving cutter are run in close mesh. Component tooth thickness is obtained by controlling the final centre distance between cutter and gear axes.

73rd and 74th Slides ~ Operation Types ~ Shavers.

1. Conventional.

In this method the gear and shaving cutter normally operate at between 10 and 15 degrees crossed axis angle. This angle may be reduced to as little as 3 degrees where adjacent fouling points restrict ideal shaving conditions.

The shaving cutter is reciprocated in an axial direction with respect to the gear this progressively moves the contact area across the face of the gear tooth until the desired size is obtained.

The serrations on a conventional shaving cutter are annular. (I.e. in the same position on all teeth.)

The shaving cutter is normally ground with a true helix. Helix modification of the gear tooth is achieved by a mechanism in the shaving machine, which varies the centre distance as the contact area progresses across the tooth. The contact area on the shaving cutter teeth is concentrated around the crossed axis point as it tracks across the gear tooth face.

2. Diagonal.

In this method the shaving cutter is caused to traverse at an angle to the gear axis. This angle is normally called the diagonal angle.

The area of contact moves across the face of the gear tooth as the shaving cutter passes across the gear axis the contact area on the shaving cutter also tracks across the face of the shaving cutter tooth.

Conventional shaving cutters can be used in the diagonal mode, however if the diagonal angle exceeds about 60 degrees then the shaving cutter must be produced with differential serrations to prevent tracking. Differential serrations is the name given to a pattern of serrations where their relative position changes progressively from one tooth to the next producing a lead of serrations. The pattern of serrations is normally designed to suit particular operating conditions.



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The use of diagonal shaving speeds up the operation and allows both profile and helix modification capabilities to be built into shaving cutter profiles.

3. Underpass.

Under pass is a limiting condition of diagonal shaving with a diagonal angle of 90 degrees.

4. Plunge.

Plunge shaving is a further development in the quest for faster production times.

In this case the shaving cutter is fed radially to the required centre distance. The area of contact extends across the face of the shaving cutter and gear as the centre distance reduces. It is necessary to modify the shaving cutter tooth helix in the form of a hollow to maintain constant gear tooth thickness across the facewidth.

There are a number of other methods of shaving but these only used in special cases so they will not be covered in this presentation.

Having chosen the method of shaving the type of Shaving Cutter can be specified.

HELIX ANGLE.

The approximate Helix Angle of the Shaving Cutter is derived from the Gear Helix Angle and the desired Crossed Axis Angle of the Shaving Cutter. Whilst the design engineer is normally free to choose the desired crossed axis angle, in some instances his choice is limited by the possibility of collision with adjacent gears flanges etc. In these cases calculations are made to establish the maximum crossed axis angle possible.

DIAMETER / NUMBER OF TEETH.

The Shaving Cutter final outside diameter will usually only be fixed at the design stage, however

Shaving Cutters are normally referred to by number of teeth and a nominal diameter. The Pitch and the Shaving Machine capacity normally control the nominal diameter of the Shaving Cutter from which the number of teeth is chosen to suit this diameter whilst bearing in mind two important factors.

a) Use a prime number of teeth where possible to ensure hunting action. b) Check available Index Plates used on serrating and profile grinding machines, this

should include those held at the customers sharpening facility where applicable.

Having fixed the number of teeth in Shaving Cutter the approximate outside diameter can be confirmed to sufficient accuracy for quotation purposes.

The outside diameter (new cutter) will be determined at the design stage when the operating pressure angle range as been established.



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WIDTH

Conventional shaving cutters are normally made to standard widths usually about ¾ inch (20mm).

The width of a Cutter used for Diagonal, Underpass or Plunge Shaving is determined by the facewidth of the gear and the diagonal angle. The area of contact must always lie within the effective facewidth.

SERRATIONS.

75th Slide ~ Serration Types ~Shavers.

In the USA and UK serrations are normally formed in the transverse plane whilst in Europe they are formed normal to the tooth helix. There is no real advantage in either system, however in recent times the European system has enjoyed increasing popularity.

European type serrations cannot be formed on USA/UK system serrating machine.

CLEARANCE GROOVES.

These are clearances / undercuts machined in the shaving cutter tooth root spaces to facilitate swarf clearance and in particular to provide clearance for the tool tip when machining the serrations.

There are two systems in general use: -

a) drilled holes. (mainly USA / UK) b) milled dovetail grooves. (mainly Europe)

There is no real advantage in either system.

76th Slide Substrate Materials and Surface Treatments ~ Shapers

Substrate Material ~ most shaving cutters are manufactured from conventional High Speed Steel (M2) although there is a trend to use more Powder Metallurgy H.S.S. with its superior construction and resistance to edge breakdown. There are a number of different steels under test in the search for an optimum substrate. There is no case to warrant the use of sophisticated high alloyed H.S.S. as the cutting conditions are not arduous and little heat is generated.

Extensive information on Material Properties have been compiled in databases and these or used to optimise Tool Substrate Material selection.

Surface Treatments ~ it is normal practice to subject the shaving cutter to an additional surface treatment after the main heat treatment operation. to remove any de-carburised layer in the serration grooves. This is necessary because these are not normally machined after heat treatment.



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Shaving cutters are not normally ceramic coated. The surfaces in the serration grooves have a relatively poor finish and are difficult to clean effectively and this can cause problems during the coating process.

DESIGN LIFE.

The design life of the shaving cutter must be carefully selected by taking into account a number of factors.

One method is to calculate values for all relevant elements of the shaving cutter over a series of operating pressure angles meshing with the product gear. In addition the operating pressure angles where the balanced condition applies are calculated. The balanced conditions occur at the operating pressure angles where the same numbers of flanks are in contact on each side of the gear and cutter teeth.

77th,78th and 79th Slides ~ CAD Output (Shav1).

80th Slide ~ CAD Output ~ Design Graph.

The design life can then be chosen and positioned relative to a balanced condition to ensure satisfactory meshing conditions throughout the life of the shaver.

Almost all shavers are designed to operate with a fixed crossover / shaved to diameter on the gear hence a single balance condition is attained at some point in the cutter life. However it is possible to keep the shaving conditions in constant balance by varying the shaved to diameter through the design life. This is made possible by providing either an extended undercut (equal to or less than the shaving stock) at the root of the gear teeth or no undercut.

81st Slide ~ Undercuts & Serrations Shavers

A further consideration is that the cutter teeth must have adequate beam strength to prevent distortion and possible fracture under load. This is normally controlled by the size and form of the undercuts relative to tooth thickness at the base of the teeth.

It is normal practice to ensure that the serrations on each flank do not crossover below the outside diameter. However on very small pitch cutters where it is extremely difficult to form normal serrations they are formed by machining circular grooves across the cutter facewidth.

Computer programs are available, which automatically optimise the cutter design by taking into account all relevant factors. These programs output complete design and manufacturing data including regrind information. It is normally expected that the shaving cutter design will produce a satisfactory component as designed, however on occasions the form needs to be developed to suit the operating conditions (i.e. shaving machine condition, preshave form etc.).

82nd,83rd and 84th Slides ~ CAD Output (Shav2).

85th Slide ~ Regrind Information ~ Shavers The regrind information is in the form of a graph where for any tooth thickness the position of the relevant diameters are defined in the form of roll angles or graph lengths measured from the base diameter. The tooth thickness is normally specified as a dimension over a ball / roller located in a tooth space, whilst the relevant outside radius is quantified to obtain the correct shaved-to diameter.



Page 21

QUALITY

86th and 87th Slides ~ Sample Standards ~ Shavers.

Shaving Cutters are normally manufactured to International, National, or Company Standards, the relevant quality being agreed at the quotation stage.

Modern computer programs enable correlation between the desired gear quality and the cutter quality. However it is important to recognise that shaving machine and fixturing quality together with blank quality have a major influence on the quality of gear produced.

88th, 89th and 90th Slides ~ Optimisation Output ~ Shavers.

91st Slide ~ Picture of Large Gear.

BGA Training Notes

TN4-2: Gear Cutting Tools - Part 2


1

BGA Training Notes TN4 -2

Gear Cutting Tools - Part 2

Author:

Anthony Hadwick C.Eng. MIEE Consultant Spring Cottage Snelsmore Common Newbury Berkshire RG14 3BN

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Index

GEAR CUTTING TOOLS - PART 2 .....................................................................................................................4

1.0 CHOICE ......................................................................................................................................................5

FORMING PROCESSES ......................................................................................................................................6 Milling. ..............................................................................................................................................................6 Broaching ..........................................................................................................................................................6 CBN Grinding....................................................................................................................................................6

GENERATING PROCESSES...............................................................................................................................7 Planing ..............................................................................................................................................................7 Shaping..............................................................................................................................................................7 Hobbing.............................................................................................................................................................7

FINISHING PROCESSES.....................................................................................................................................8 Shaving ..............................................................................................................................................................8 Grinding ............................................................................................................................................................8 Honing/Deburring .............................................................................................................................................8

2.0 SELECTION ...............................................................................................................................................9

GEAR SHAPING ..................................................................................................................................................9 GEAR HOBBING .................................................................................................................................................9 SELECTION FACTORS.....................................................................................................................................10

What Type of Hob Should Be Chosen? ............................................................................................................10 Options ............................................................................................................................................................11 Finishing Operations.......................................................................................................................................12

3.0 TOOL MATERIALS ................................................................................................................................13

RELATIONSHIP BETWEEN WEAR RESISTANCE AND HOT HARDNESS ...............................................15 TOUGHNESS COMPARISON OF POPULAR HSS..........................................................................................16 MATERIAL REQUIREMENTS FOR MODERN GEAR TOOLS .....................................................................17 INFLUENCE OF ELEMENTS IN HSS ..............................................................................................................18 COMPARATIVE HSS CHEMICAL ANALYSIS ..............................................................................................19

4.0 TOOL COATINGS...................................................................................................................................20

COATING COMPARISON.................................................................................................................................20

5.0 SPEEDS & FEEDS ...................................................................................................................................21

SUPER HIGH SPEED DRY CUT HOBBING....................................................................................................................22 Comparison Test Data on Liebherr LC153 Hobbing Machine........................................................................24

6.0 QUALITY..................................................................................................................................................25

GEAR QUALITY FROM VARIOUS GEAR PRODUCTION METHODS.................................................................................26

7.0 GEAR TOOL MAINTENANCE .............................................................................................................28

GENERAL MAINTENANCE.............................................................................................................................28 SHARPENING OF GEAR TOOLS.....................................................................................................................28 RE-SHARPENING..............................................................................................................................................29 GASH RADIALITY. ...........................................................................................................................................30 GASH TO GASH SPACING...............................................................................................................................30 GASH LEAD ERRORS.......................................................................................................................................30 THE EFFECTS OF RUN-OUT ...........................................................................................................................30 MINIMISING RUN-OUT PROBLEMS..............................................................................................................31 QUALITY OF GRINDING .................................................................................................................................31 IN CONCLUSION...............................................................................................................................................32

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GEAR CUTTING TOOLS - PART 2 Part 2 covers the following items relating to gear tooling.

1.0 CHOICE.

The various methods we might expect to use in gear production will be discussed.

2.0 SELECTION.

How we go about deciding which method we should use.

3.0 TOOL MATERIALS.

What should our gear tool be made of?

4.0 TOOL COATINGS.

Should we consider surface treatments?

5.0 SPEEDS AND FEEDS.

It is important to use the tool efficiently.

6.0 QUALITY.

How to achieve the accuracy required.

7.0 GEAR TOOL MAINTAINANCE.

Ensure you always get the best from your tool.

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1.0 CHOICE

There are a number of ways to produce gears, and the final choice of tool will rest on the equipment available to you and the design of the part. Other influential factors will be the number of gears to be produced and the quality to be attained.

MOST COMMON PRODUCTION METHODS

• HOBBING

• SHAPING

• PLANING

• MILLING

• CBN GRINDING

• FORM ROLLING

FINISHING APPLICATIONS

• SHAVING

• GRINDING

• HONING

The production methods mentioned can be further split into two groups, as follows

• Gear Generation

• Gear Forming.

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FORMING PROCESSES

The forming process is where the cutting tool profile is the same shape as the gear tooth space it will cut.

Examples are:

MILLING.

Most commonly used in tool rooms for jobbing work where quality and accuracy are not important. Gear milling cutters are supplied in sets of 8 cutters to cover a range of teeth from 12- RACK. The tooth forms as you will appreciate are only approximate.

Cutters can also be manufactured to order to suit specific profiles.

When producing ‘power take off’ shafts or similar parts, it is fairly common to use sets or pairs of cutters on special multi-head machines. These cutters are matched on diameter, width and with their forms central. Special keyway formation gives a staggered cutting effect enabling optimum cutting conditions. High quality combined with special grades of HSS and coatings ensure very high productivity from these tools.

BROACHING

Many high volume internal gears, serrations and splines are produced in this way.

This is a very fast application but the high cost tooling and the requirement of very special machine tools limit its use to specialist producers mostly in the automotive and aerospace industries.

CBN GRINDING

Precision formed CBN wheels and the creep feed method of grinding are used. Due to its cost it is usually used in conjunction with large volume production or very large gears.

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GENERATING PROCESSES

The generating process occurs when a tool of one profile, produces a gear tooth form with a different shape.

PLANING

The ‘SUNDERLAND’ & ‘MAAG’ Machines are the most common of this type. It is a process suited to low volume, because of the slow laborious cutting strokes. However it is very versatile and can be used for spur and helical gears.

SHAPING

Generally used for the production of internal gears or external gears where there is some form of restriction close to the teeth.

HOBBING

Where ever possible this is the preferred method for production, and with modern grades of tool materials, surface coatings and optimum hob design it is the most productive.

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FINISHING PROCESSES

SHAVING

For this application you require a special machine and shaving tools. The gears are pre-cut using one of the previous processes to leave a few microns on the flanks this stock is then removed by the shaving process and at the same time the surface finish and accuracy are improved. The shaving cutter is very much like a high precision gear with serrations cut into the teeth to give a shaving action when the tool rotates the teeth slide over the gears tooth surface using the cross axis method. There are four basic types of shaving tool.

• Conventional; • Diagonal; • Underpass; • Plunge.

GRINDING

This again requires special machinery and is used mainly to improve the accuracy and surface finish on hardened gears. There are various types of machines used for gear grinding, some use the generating process whilst others use form grinding method.

HONING/DEBURRING

Mostly used to remove heat treatment scale or nicks and burrs caused by poor handling between gear cutting and heat-treatment. The honing machines operate using a high-pressure contact on sprung centres and the cross axis method. Similar to shaving but the tools do not have serrations only a hard abrasive surface that does the work.

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2.0 SELECTION In selecting the tooling best suited to your application we must consider a number of factors.

In this instance we will consider that subcontract, milling, planing, and broaching are all ruled out and we are to utilise the two most common production techniques namely HOBBING & SHAPING. Both of these methods can produce finished cut gears or gears with a stock allowance for later finishing operations.

GEAR SHAPING

This process is normally used where gears have run out restrictions or for internal gears.

There are various designs available to suit different situations.

• The most common is the standard Disc Pattern. • When an adjacent restriction occurs it will be necessary to use an Extended Back Boss

Pattern ’EBB’, With this design the retaining nut is concealed within the cutter body. • When similar restrictions apply but size is also critical then the alternative designs of

Screwed Hub Pattern or Shank Pattern have to be used.

When cutting internal gears there will be a maximum number of teeth in the cutter for any given number of teeth in the gear, and the Pressure Angle And tooth depth of the gear influences this. Where there is any doubt, the tool manufacturers will be able to use their computer programs to advise you the precise situation.

Gear Shaper Cutters can be designed to cut straight spur gears, helical gears, racks, worms, coarse threads, face gears and many special profiles.

GEAR HOBBING

This is the process we recommend you use were ever possible and for this reason we will cover it in more detail.

Hobbing is fast, accurate and versatile; one hob can be used for cutting straight spur gears or helical gears of the same normal pitch and normal pressure angle.

However there are considerations to be made in the choice of hob design as a number of alternatives can now be offered. Using the hob correctly is very important. If the hob is running at an inappropriate surface speed for the material being cut excessive wear occur. (We will cover surface speeds in a later section today.) It is also critical to the accuracy of the finished gear that the hobs are correctly mounted, to run true both ends within 10 Microns and any run out should be in phase both ends. To achieve the best productivity it is important to discuss with the hob manufacturer your needs and machinery available to you, they can then propose the ideal specification for you.

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SELECTION FACTORS

• Volume of gears to be produced

• Final quality of gear

• Final hardness/hardness when cut

• Machines available

• Basic method to be used

• Is sub-contract required or necessary

WHAT TYPE OF HOB SHOULD BE CHOSEN?

There is the basic hob design where the diameter and length are similar and standard to suit its pitch. These hobs have conventional gash design and are often available from stock. These are ideal for jobbing work or small trial batches.

If you have modern CNC hobbing machines which can operate at speeds in excess of 1100 RPM then your choice is wide open and for high production we recommend small diameter long hobs of solid design, in some cases on solid shanks as opposed to bore type. These hobs when produced in special high grade high speed steel with a PVD coating will operate at very high speeds an still give good tool life (usually around135mtrs/min). However there are currently under trial various new grades of HSS and multi layer coatings which enable surface speeds of around 300/350 meters per min to be used and a few trials being carried out at around 450 metres/min. If high quality gears are required grade ‘AA’ multi gash hobs can in a number of cases overcome the need for shaving or grinding.

For companies who still do not have the luxury of CNC hobbing machines but need to compete to the best of their ability then we recommend they have hobs designed to suit the available criteria. What is done in these circumstances is to first consider the material to be cut and its machineability, then knowing the grade of HSS and type of coating, the optimum cutting speed can be determined. From here given the range of available RPM of the hob spindle we can calculate back to the optimum hob diameter. With the slower RPM available on the conventional machine your hob will be of a larger diameter than the modern hob but will give you the better economy, extra starts can also be introduced to further increase productivity. You can, if higher quality gears are required still add extra gashes to improve the generated profile and another bonus is the reduction in tooth loading.

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OPTIONS Conventional/Standard hob.

Often available from stock

Small diameter long length.

These are usually multi gash design in special PM HSS and coated, many are designed as multi start type. (for use on high speed CNC hobbing machines)

Larger diameter longer length.

In this case the hob diameter is calculated to operate at the optimum surface speed to suit the material being cut and close to the top RPM of the hobbing machine. These can also be multi start for further improvements in productivity (for use with slower convential hobbing machines)

Other types of hobs which are used far less often in modern factories and are as follows:

Inserted Blade.

There are very few manufacturers still producing these as the ‘Modern Hob’ out performs them in most volume applications.

Solid Carbide.

Recent results have been encouraging but very careful consideration must be taken when choosing these. Firstly a specially designed machine is required; most CNC machines are not suitable, as the high temperature swarf is required to be removed directly from the machine. Carbide hobs are brittle and good housekeeping is essential. Even hob sharpening can be a problem as quite often two separate angles are required on the cutting face.

Carbide Skiving Hobs.

Special purpose tools for hard hobbing of larger type hardened gears.

Throw away Hobs.

These are similar to the modern hob but have so many gashes that there is no opportunity to re-sharpen them. They appear to be an expensive option.

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FINISHING OPERATIONS

If it should be necessary to select a finishing operation then the choice rest on various factors.

Gear Shaving.

This will improve the surface finish and general quality of the gear but it is not an exact science and often the involute profiles leave something to be desired. Using a tool designed for the specific gear where a very good design balance is obtained will help. It is also critical to keep the tools sharp. Shaving takes place prior to any heat treatment process and therefore gear distortion will more likely occur in the finisher product.

Gear Grinding.

This is the best way to ensure accuracy and surface finish on hardened gears, care must be taken to ensure the grinding wheel is well dressed to minimise the risk of creating grinding burns. Grinding is an expensive operation requiring special gear grinding machines and is one of the most common operations in gear production to be put with specialist sub-contract companies.

Gear Honing/Deburring.

This process is used by only a small number of modern companies, the main advantage is that it can be used on hardened gears to remove nicks & burrs or heat-treatment scale. (With good house keeping, damage to the gear form should not occur thus reducing the need to hone)

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3.0 TOOL MATERIALS In the UK most standard and ‘stock’ type gear cutting tools are manufactured from M2 HSS (6-5-2) but to achieve improved tool life and more importantly much greater productivity we generally manufacture tools to order and select a powder metalurgy grade of HSS.

Powder metalurgy grades of HSS are produced in a different manor to conventional grades.

With a conventional grade of HSS all the alloys are thrown into the furnace with the basic melt and then poured to form a large solid ingot, which cools slowly. In the cooling process the centre section cools at a slower rate and as a result a very poor structure is obtained in this area, the ingots are then forged and rolled which improves the structure, but there still remain large carbides. These large carbides can cause two major problems, their hardness makes grinding difficult and if one of these long carbides should be close to the tool cutting edge the softer matrix can chip away in use causing premature wear. Conventionally produced HSS will also have a greater and less predictable movement in the hardening and tempering process of tool manufacture than the PM grades.

Powder metalurgy HSS is produced in an induction furnace by melting the mixture of steel scrap plus the required alloys. This is then cleaned in a ‘Tundish’ system, which has been specifically designed to guarantee ‘super clean’ products. The mixture is then poured through an inert gas and is atomised to form a powder. Each small powder grain has cooled rapidly giving it a fine even structure. These tiny particles are then poured into a capsule made of steel plate which is vibrated to pack the particles as tight as possible. A cover is welded onto the capsule after the air has been drawn out. The capsule and its contents are then compacted in two stages by isostatic pressing. First of all, at room temperature at a pressure of 4000 atm. Then, at 1150 degrees C at a pressure of 1000 atm. forming a completely dense ingot. The ingot is then subsequently processed in the conventional manner by forging and rolling the bar to the sizes required.

The new structure of Powder metalurgy HSS gives improved grinding characteristics improved strength, and makes it possible to increase the alloy content to give improved performances.

There are a number of options we can chose from each with its own benefits. However for most gear tool applications Erasteel ASP 2030 or other suppliers equivalent cover the majority of applications.

We can see from the next over heads how the various properties of the different grades of material compare. In the first case we are considering ‘HOT HARDNESS with WEAR RESISTANCE’ if a tool is to operate at a high surface speed, it must have good hot hardness. M35 has basically the same alloys as M2 except it has an added element, namely 5% Co. We can see both have a similar wear resistance but the extra 5% cobalt gives it greater ‘Hot Hardness’ so it can there fore be used at higher cutting speeds. The same applies with ASP2023 and ASP2030 where the cobalt additive is 8.5%.

ASP 2060 would appear the best choice when only considering ‘Hot Hardness’ and ‘Wear Resistance’ but moving on to the next overhead we see we are loosing out to ‘TOUGHNESS’ so a balancing act to suit the application has to be made.

We need a combination of all but a lack of toughness results in brittle tools that can easily chip; also with out hot hardness a tool will very quickly wear if operating at too higher surface speed.

So what do we need for the modern gear tool? How do specific elements influence the HSS?

By comparing the chemical analysis of different grades of HSS we can decide which will suit a given requirement.

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RELATIONSHIP BETWEEN WEAR RESISTANCE AND HOT HARDNESS

K 94 ASP2060

92

90

88

86

84

82

80

78

76

74

72 T15 ASP2052

70 REX T15 REX T76

68 M15

66

64 CPM M4

62 M4 T6

60 ASP2023 ASP2030

58 M3

56 T4

54 T1 M7

52 M42

50 M2 M35

48

We 25 30 35 40 45 50 55 60 65

K= Wear Resistance We = Hot hardness

It should be noted that the materials indicated in bold type are the preferred powder metallurgy grades and these should be used wherever coatings are to be added. PM HSS has a better structure than conventional grades of HSS – it is tougher, performs better and grinds more easily.

It is important to consider each material's toughness in addition to the above information. For example, ASP 2060 has a higher hot hardness coupled to wear resistance but it also has a poor toughness that can result in tools chipping under certain circumstances.

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TOUGHNESS COMPARISON OF POPULAR HSS

S690 & CPM M4

S390 & ASP 2052

ASP 2060

ASP 2030

ASP 2023

M35

M2 0 5 10 15 20 25 30 35 40 45 50 55 60

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MATERIAL REQUIREMENTS FOR MODERN GEAR TOOLS

HIGH STRENGTH Obtained from CARBON + ALLOYS

GOOD TOUGHNESS Obtained from the STRUCTURE OF POWDER METALURGY HSS

HIGH WEAR RESISTANCE

High Vanadium gives HARD CARBIDES

GOOD HOT HARDNESS

Obtained from high COBALT

GOOD COATING SUBSTRATE

Even grain structure with hardness and strength, again from POWDER METALURGY HSS

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INFLUENCE OF ELEMENTS IN HSS

CARBON (C) ESSENTIAL FOR MARTENSITE TRANSFORMATION - THE ULTIMATE HARDNESS OF HSS, ALSO THE AMOUNT AND TYPE OF CARBIDE.

CHROMIUM (CR) HARDENABILITY – WEAR RESISTANCE

TUNGSTEN (W) CARBIDES _ WEAR RESISTANCE SECONDARY HARDNESS MOLYBDENUM (MO) 1% MO = 2% W

VANADIUM (V) STABLE HARD CARBIDES

COBALT (CO) HOT HARDNESS – IMPROVES TEMPER RESISTANCE

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COMPARATIVE HSS CHEMICAL ANALYSIS

C Cr Mo W Co V M2 0.9 4.2 5.0 6.4 Nil 1.9 M3-2 1.2 4.1 5.0 6.2 Nil 3.0 M4 1.3 4.2 4.5 5.5 Nil 4.0 M35 0.9 4.2 5.0 6.4 4.8 1.8 M42 1.1 3.8 9.4 1.5 8.0 1.2 T1 0.75 4.1 Nil 18.0 Nil 1.1 T4 0.75 4.1 Nil 18.0 5.0 1.1 PM 23 GRADES 1.3 4.1 5.0 6.4 Nil 3.1 PM 30 GRADES 1.3 4.2 5.0 6.4 8.5 3.1 S390 GRADES 1.6 4.8 2.0 10.5 8.0 5.0 PM 60 GRADES 2.3 4.2 7.0 6.5 10.5 6.5

Items shaded yellow

These are the important powder metalurgy HSS grades.

Items shaded red Have cobalt for hot- hardness.

Items shaded blue These have vanadium content for hard carbides, giving wear resistance.

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4.0 TOOL COATINGS

For gear tools we have until now generally accept that TiN coating applied by the PVD process gives the most uniform results. Under certain conditions and for specific materials we have the choice of other PVD coatings and current developments of multi-layer coatings.

Coatings are applied to the surface of the material in a vacuum chamber using the physical vapour deposition (PVD) process. The coatings are only microns thick and do not significantly effect the tool geometry. The hardness of TiN is approximately 2300 Vickers when measured on a perfectly prepared sample. The substrate to which the coating is to be added must be consistent which is why the powder metalurgy grades of HSS are preferred. As we have seen the PM HSS have a very even structure which can be ground safely, conventional grades on the other hand have large carbides distributed throughout them and these often cause grinding burns and therefore a martensitic areas on which the coating does not adhere well. Tools, which have been coated, improve productivity in two major ways. Being so much harder than the substrate, tool life is extended considerably and the hard surface with reduced friction characteristic permits the surface speeds to be increased with dramatic results.

Tools should be pre-paired correctly prior to coating, they must be well cleaned and all burrs removed. Under normal conditions burrs are usually knocked off in the cutting process but if they are coated over they become very hard areas which will cause damage to surface being cut.

COATING COMPARISON

TiN TiCN TiAlN ‘FUTURA’

TiAlN

‘X.TREME’

Hardness Vickers 2300 3000 3000 3500

Max. Operating Temp. 600 C 400 C 680C 800C

Colour yellow/gold blue/grey violet/grey purple/grey

TiN Coating is used as the standard coating

TiCN Coating is used more for abrasive materials and higher performance

TiAlN ‘FUTURA’ is ideal for high temperature conditions, and is one of the coatings being used in dry hobbing trials. It is a multi layer type coating.

TiAlN ‘X.TREME’ is recommended for skive hobbing (carbide tools). This is a mono coating.

Futura & X.Trame are Balzer trade names. Multi Arc. TecVac. & J J Castings etc. are all coating companies, and all have their own variations for you to chose from.

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5.0 SPEEDS & FEEDS The surface speed at which the gear tool operates is critical to tool life; we must therefore select the surface speed to suit tool life/productivity for the gear material being cut. In a few cases tool life is secondary to cycle time, but this can be a dangerous scenario as once the cutting edge becomes dull the wear rate starts to climb rapidly and tooth failure can then occur.

To chose the optimum surface speed we must consider the machineability of the gear material and then decide on which grade of HSS and coating are to be used.

From this information we can study the attached graph and chose one of the available speeds. For ASP2030 we can see there are three speed zones to chose from.

1) Using the material, without coating. 2) Coated. 3) Again coated but used with the modern CNC type machines.

I am also assuming all tools, which have been coated, are re-coated after they have been sharpened.

I have found in the past that surface speed/tool life can be related to a frequency curve.

We find that at point ‘A’ tool life is very poor but at ‘B’ it is good, there are a number of points along the curve at which this occur.

The frequency curve gives an example for one of the common gear steels namely SAE8620 with a tensile hardness of 40/45 tonnes/in squared which is around 150/180 BHN.

For ASP 2030 we require a surface speed of 36 Mtrs/Min.

For ASP 2030 with TiN we can now increase to a surface speed of 73 Mtrs/Min. Also for ASP 2030 with TiN coating on CNC M/Cs a surface speed of 135 Mtrs/Min can be achieved.

This area is ideally suited to the modern hob, small diameter long length with multi gash and multiple starts.

An interesting point occurs at the higher surface speeds in that wear patterns change. With high speed hobbing it is more common to see crater wear, and the wear rates when this occur are much less than is normally seen.

Currently trials of special materials and with new coatings are being conducted in the UK & USA where speeds of 300 to 350 Mtrs/Min are being tried for dry cutting. One requirement for these tools is believed to be a super finish on the cutting face. But from this we can see that the frequency curve probably continues on for more cycles. With these new HSS grades and multi-layer coatings it will be interesting to see how much further productivity can go.

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SUPER HIGH SPEED DRY CUT HOBBING

I have accumulated some information on the topic of ‘super high speed dry cut hobbing’ which you may find interesting.

1) Very High Surface speeds Currently around 240M/M

2) High Chip Loads

3) Increased Tool Life

4) Environmentally Friendly

The lack of fumes produced compared with normal conditions, and combined with not having to dispose of used cutting oil helps the environment.

1) Reduced Cost Per Part

2) Reduced Tool Cost Per Part

3) Higher Productivity than Carbide

The ‘Futura’ coated hobs are produced from a new grade of PM HSS where the powder is much finer than that of previously tried and tested PM HSS. It is much tougher and even in the very adverse conditions of ‘super high speed dry cut hobbing’ it out performs Solid Carbide hobs in similar circumstances.

When cutting a 56 tooth 2.54 Mod Helical Gear with a 25 mm face width in SAE 8620 (225 BHN 21 HRC) an increase of over 20% was recorded in the number of parts per regrind, but more importantly the cutting time was reduced by 29%.

Overall cost savings of 25/30% are being achieved over the gears being produced by carbide hobs.

The initial hob cost is approximately half that of the carbide hob for the same job.

Carbide tooling also requires much greater handling care as it is very prone to chipping.

Resharpening of the new generation HSS tools is much easier than for carbide, but it is still recommended they are returned to the original manufacturer to be sharpened to ensure the optimum required surface finish is obtained prior to re-coating.

It is important to have, and use, the tool supplier’s recommendation for each hob and gear combination as the chip load also plays a very important part at these very high speeds. (Higher chip loads than those used by carbide tools are recommended.) The tool supplier will take into account the power requirement and the gash design to ensure adequate swarf removal.

Removal of the swarf from the machine has to be swift as the heat generated during ‘super high speed dry cut hobbing’ is mainly in each chip. The hob and component temperature are well within that which can safely be handled.

As with carbide hobbing it is important to use the new generation hobbing machines dedicated to dry cutting. These machines are equipped with special stainless steel swarf conveyors and vibrators for swift clearance of the swarf away from the hobbing machine.

These hobs are suitable for cutting abrasive and harder type gear materials.

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‘Super high speed dry cut hobbing’

• Very High Surface speeds

• High Chip Loads

• Increased Tool Life

• Environmentally Friendly

• Reduced Cost Per Part

• Reduced Tool Cost Per Part

• Higher Productivity than Carbide

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COMPARISON TEST DATA ON LIEBHERR LC153 HOBBING MACHINE

Gear Data

2.54 Mod 20 deg. PA. 56 Teeth 14.133 deg. HA

151.76mm Outside Diameter 25mm Face width

Material SAE 8620 (225 BHN 21 HRC)

Hob Material Solid Carbide Acealloy 240

Coating TiN Balinit Futura

Hob Diameter 80mm 75mm

Hob Tooth Length 162mm

Number of Gashes 19 16

Number of Starts 2 2

Cutting Speed 280 M/M 240M/M

Hob RPM 1114 1019

Feed Per Rev. 2.25mm 3.10mm

Chip Thickness 0.131mm 0.180mm

Cutting Time 35Sec. 24.9Sec.

Parts Per Hour 86 118

Parts Per Regrind 774 942

Tool Life 4.0M 5.0M

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6.0 QUALITY Gear quality will depend upon a number of factors

• The type of process used

• The quality of the tool being used

• How well the tool has been set up

• How well the tool has been maintained

• The condition of the machine tool

The attached chart gives a guide to the type of process or quality of tool to achieve tolerances to AGMA and DIN specifications. I should now introduce an additional column for Grade ‘AA’ Multi-Gash hobs as a higher quality gear can be produced with these tools. A number of companies are now finish hobbing in place of shaving and getting better involute profiles then on their previously shaved gears.

One of the most common reasons for rejection of a gear by quality control is the poor involute profile. These errors can usually be related back to tool setting either on the gear-cutting machine or in tool resharpening.

To demonstrate the importance of tool set up I can show you a specific problem one customer experienced.

I received a phone call from a customer who was obtaining profile errors, which were so bad that they could be seen by eye. The background was that the customer was using an involute spline hob, which had previously performed correctly and it had not been sharpened since it was last used. One component flank looked OK but the other had excessive trimming at the tip. My first comment was that this was the result of poor generation caused by run out. It transpired that on the previous job there had been a setting error that resulted in the hob feeding excessively into the work piece resulting in a badly bent arbour. The TIR was 0.180mm and as there was no spare available and delivery of the parts was critical a delay of 2/3 weeks to obtain a new arbour was not acceptable.

What could be done?

The design was simulated on computer and then we introduced the 0.180mm run out and ran the generation sequence at 90 degrees out of phase, (i.e. The point of maximum run out was generating the involute at the pitch line.) we found the results were exactly as described. By re-running the generation sequence, but with the run out in phase, (i.e. The point of maximum run out was now generating the involute on the centre line of the tooth space.) we obtained a balanced profile. Whilst the involute profile was not exactly correct it was acceptable in this instance for a spline coupling. The customer then reset his machine to the new recommended position, and by hob shifting an exact tooth pitch every time hob shift was required a status quo was achieved.

BGA Training Notes



26

GEAR QUALITY FROM VARIOUS GEAR PRODUCTION METHODS

AGMA DIN

quality profile run-out pitch variation lead t/t composite t/t total CBN grind

grind shave gear shape

hob 'AA' hob 'A' hob 'B'

15 3 4 3 to 4 4 3 to 4 4 to 5 Normal 14 4 5 4 to 5 4 3 to 4 5 to 6 Ideal Normal 13 5 6 5 to 6 4 4 to 5 6 to 7 Normal Normal Ideal 12 6 7 6 to 7 4 5 to 6 7 to 8 Normal Normal Normal Ideal 11 7 8 7 to 8 5 6 to 7 8 to 9 Normal Normal Normal Ideal Normal Ideal 10 8 9 8 to 9 6 7 to 8 9 to 10 Normal Normal Normal Normal Normal Normal Ideal 9 9 10 9 to 10 6 8 to 9 10 to 11 Normal Normal Normal Normal Normal Normal 8 10 11 10 to 11 7 9 to 10 11 to 12 Normal Normal Normal Normal Normal 7 11 12 11 to 12 8 10 to 11 12 Normal Normal

Ideal Achievable with ideal conditions

Normal Achievable with normal conditions

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Page 27

In the same way when resharpening a hob if it is not set up to a TIR of 10 microns both ends, with the high spots in line, the run out sharpened into it will cause generating errors when next used. A more detailed explanation will be given in tool maintenance shortly.

As we saw from the chart showing tool quality & types of tool, higher quality can be achieved by purchasing higher-grade cutting tools. Tool design also can have a marked effect on quality.

UK manufacturers usually supply hobs and Gear Shaper Cutters unless requested otherwise to grade ‘A’. However, a few companies supply tools, which are only grade ‘B’. It is in recent times becoming more important to purchase tools of grade ‘AA’ or better to achieve the ever increasing quality requirements.

Where tool design is used to improve quality we can use the multi-gash hob as an example.

From experience we have found that to give the customer the benefit of multi-gash design and still maintain a reasonable tool life then increasing the number of gashes by 1.3 to 1.5 times that of a conventional hob works best.

The following sequences of graphics enable us to consider a conventional hob design that is

90mm diameter, 5.21DP with 10 gashes. This we follow with designs where the numbers of gashes have been increased to 12 and 14 respectively.

First we consider the design with 10 gashes.

The end view of the hob looks good and we have 45% of the gash pitch available for sharpening. The generated involute profile is as one would expect under normal hobbing conditions. We now show the generated surface finish magnified and simulated on to the involute profile, this shows facets up to 0.0002” (5 microns). Next we see the magnified involute chart and shortly we will study a centre section of this.

Now we will consider the design with 12 gashes.

The sharpening life has reduced to 39% of a smaller gash pitch but the generated surface finish has improved.

Now we consider the design with 14 gashes.

The sharpening life has reduced further to 34% of yet a smaller gash pitch and the tooth begins to look spiky, but we can see the generated surface finish magnified onto the involute profile is only 0.0001” (2.5 microns).

Referring back to the centre section of the involute chart and superimposing all three designs together we can clearly see the advantages.

• Surface finish of 0.0002” for 10 gashes has halved to 0.0001” with 14 gashes

• The length of cut has reduced in relationship to the number of gashes

• The amount of stock removed in this area is greatly reduced, and this has the advantage of reducing tooth load and tooth wear. It is normal to find that tooth wear is reduced to such a level that the multi-gash hob actually produces more parts than the conventional design.

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7.0 GEAR TOOL MAINTENANCE

GENERAL MAINTENANCE

This is an area, which is the cause of so many problems and that with a little more care in sharpening and handling these would be eradicated.

When you first receive your new tool it will have been checked to ensure it is to the standard ordered and will there fore produce good quality gears. It will be dipped in a peelable, oil-based coating and packed in its own box for protection, when ever possible these protection methods should be maintained throughout the tool life.

Correct handling and storage on the shop floor is critical, HSS tools are all brittle and any knocks can be the cause of chipping, the tools should be kept well apart.

Extra care is needed if you send tools to another company to be sharpened. I have seen tools returned for sharpening which have just been wrapped in newspaper and loosely packed. Once in the hands of the transport company they get thrown around, the impacts allow the tools to cut their way through the paper and as they come into contact with each other they become chipped and in the worst cases teeth actually get knocked out.

A little care and attention in handling can save a lot of money in the long run!

Special tools, such as shaving cutters, that require their involute profiles reground and the outside diameter reduced to a specific size to maintain a balanced condition, are usually returned to the manufacturers for sharpening. We will therefore concentrate on gear shaper cutters and hobs, and in the main, hobs because of their special geometric needs.

SHARPENING OF GEAR TOOLS

Wherever tools are being used, it is important to ensure the operators know exactly how much wear is permitted on the tooth prior to being sharpened. Once a tool is blunt the rate of wear escalates, and using a tool with a dull edge can result in rapid and excessive wear. Excessive wear is expensive to remove and reduces tool life considerably. For example a tool could produce 500 gears and show only 0.25 microns of wear, but cut just a few more and this wear can quickly escalate to 0.5 microns. Excessive wear will have generated considerable heat, which will have likely damaged the substrate, probably reducing the hardness and creating a white layer.

Once you have established the optimum point at which a tool should sharpened a system on how to carry this out should be established.

All to often staff in tool rooms are requested to sharpen gear generating tools with out having first had the importance of the specific elements and their functions explained. Just to sharpen away the wear to obtain a clean edge can have disastrous results on the finish cut gear if the geometry is not correct.

There are specific areas, which greatly effect the way in which a gear tool performs after sharpening.

BGA Training Notes



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RE-SHARPENING

• THE CORRECT SHARPENING ANGLES

• RUN OUT

• QUALITY OF GRINDING

• SURFACE FINISH ON THE CUTTING FACE

WITH GEAR HOBS THERE ARE EVEN MORE FACTORS TO CONSIDER

• GASH RADIALITY

• GASH TO GASH SPACING

• GASH LEAD

BGA Training Notes



Page 30

GASH RADIALITY.

Most hobs are sharpened with a radial face, and this must be maintained throughout the hob life. In a few instances, a negative or positive rake is used, and this must also be maintained throughout the hob life. Where a rake is used, sharpening data is etched on each hob.

The original hob design takes into account the type of sharpening, and any deviation outside the permitted sharpening tolerance can result in generating involute profile errors. Diagram ‘A’ indicates what to expect from hob sharpening. In this narrative, we are only dealing with hobs that have been designed with radial sharpening, as hobs designed with a rake are very much the exception.

Figure 1. This shows that a correctly sharpened tooth presents a true basic rack. Assuming everything else is correct, it will generate a true involute profile.

Figure 2. This shows a sharpening error where a positive rake has resulted. The cam relief on the hob causes the constant profile to drop from front to back on the hob tooth. Whilst the tip of the hob tooth is correct, an error has been created at the root of the hob. The cam effect is such that the tooth depth is enlarged, and the pressure angle thus reduced. A hob sharpened like this will generate an involute profile error with plus metal at the tip of the gear.

Figure 3. This shows that a negative rake will give the reverse effect to fig. 2.

Figure 4. This shows one of the most common errors in hob sharpening. It normally results when sharpening a hob with spiral gashes, using a large diameter grinding wheel that has been dressed with a straight-line dresser. The helical interference as the hob spirals past the wheel will generate a convex cutting face, and this produces hollow involute profiles.

GASH TO GASH SPACING

Figures 5 & 6 show eccentricity, and will again generate involute errors. These eccentricities are caused by uneven stock removal from gash to gash. To clarify what is happening, diagram ‘B’ shows how each gash generates the involute. If one gash is incorrect, it will leave a profile error at some point, either leaving too much stock or removing too much metal. Sharpening, when the hob is not running true introduces errors on to each gash division, which make each face cut in an incorrect position this we will see shortly.

GASH LEAD ERRORS

If the hob is not sharpened at the gash lead specified on the end of the hob, the cam relief will result in a taper on diameter. The incorrect gash lead will also result in the hob lead being increased on one side and decreased on the other.

This effectively results in a hob with a different pitch on each side of the tooth. It will, there fore, generate a gear with a reduced pressure angle on one side, and an increased pressure angle on the other. As hob shift is used to move the hob from one end to the other, variations on thickness and tooth depth also occur.

THE EFFECTS OF RUN-OUT

Run-out is a problem, both during normal use of hobs, and during sharpening. Many problems can occur if a hob is not clocked up correctly. The involute errors that occur follow various patterns, and can result from several types of run-out. You can have run-out one end only; Out-of-phase run-out,

BGA Training Notes



Page 31

wobble; and general run-out from bent arbors. Maximum run-out should be within 10 Microns each end and it is important where run-out occurs that the high points are in phase with each other.

Earlier in the section on QUALITY we saw the example where a customer was using a hob with 0.180mm run-out and the effect this had. His hob was correctly sharpened, however, similar results would have been seen if the error had been sharpened into the hob and then used correctly on the hobbing machine.

Diagram ‘E’. This shows a hob that was returned to the supplier after one third of its life had been used. A true run-out had been sharpened into the hob, and the top graph shows the cyclic error per rev. In fact, whilst producing what were considered poor quality gears, the hob was still just grade ‘A’ After correct sharpening, the lower graph was obtained, and high class gears were again produced.

MINIMISING RUN-OUT PROBLEMS.

To overcome the run-out factors, the tool room should regularly check:

A. The sharpening machine arbor run-out and the condition of the location face.

B. The divide plates are clean and division is correct

C. The spacers are in good condition

It is critical to have good spacers, if they are bruised or dirty, they can create run-out. If the faces are not parallel, flat, and square, then the arbor distorts as soon as the clamping nut is tightened, giving rise to all the problems outlined previously.

QUALITY OF GRINDING

Ideally, hobs should be sharpened on a specialised hob-sharpening machine. Most of these are fully enclosed wet grinding machines, however there are still many older dry grinding machines being used. Wet grinding is the preferred method, as it minimises problems from burning etc. However, dry grinding is acceptable if used with care.

On a dry grinding machine, the grinding wheel must be regularly dressed to ensure it does not become glazed and the stock removal should be kept down. Either of these can result in the possibility of burning the hob. Burning can create a number of problems. Very severe burning will cause fine grinding cracks, resulting in teeth dropping out.

Less severe burning will cause damage to the material structure. The heat generated will form a ‘White Layer’, (Diagram ‘F’ shows this condition, but in this example the heat generation came from the hobbing process as the chips rub the cutting face causing crater wear). Where this ‘White Layer’ occurs, there is a zone of newly formed Martensite, which will give a reduced hardness at this point. There fore when sharpening, sufficient material must be removed to take out all the wear plus the damaged substrate, once this is done the material will back to its original hardness and structure.

With many tools now being re-coated after sharpening it is more important that the tool room is aware of these factors, Coatings do not adhere as well to un-tempered Martensite, and for optimum results from coating the ‘White Layer’ must be fully removed.

In recent trials, it has been noted that the surface finish on the tool can make large differences to tool life and tool operation. It has been claimed for dry hobbing using coated HSS, that unless the surface finish is very fine the tool will break down much more rapidly.

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Page 32

IN CONCLUSION

Tool maintenance is a very important part of production. High cost savings can be achieved when tool grinding is correct.

Tools sharpened to their original condition will:

• Produce gears to the correct profile.

• Maintain their maximum wear resistance.

• Accept coating after sharpening.

• Reduce scrap levels.

• Save time in setting (as the setter is not resetting time and time again trying to improve quality from an in built run-out).

BGA Training Notes TN5: Steel Selection: The Manufacture of Engineering Steels


Page 1

BGA Training Notes TN5 Steel Selection

The Manufacture of Engineering Steels

Author: G. Haywood Macreadys Steels Paynes Lane, RUGBY, CV21 2UW [email protected] http://www.macreadys.co.uk



Page 2

Index

THE MANUFACTURE OF ENGINEERING STEELS ....................................................................................4 MELTING PROCESSES.....................................................................................................................................4 CASTING PROCESSES .....................................................................................................................................4 ROLLING............................................................................................................................................................4 TESTING.............................................................................................................................................................5 FINAL PRODUCT ..............................................................................................................................................5 HEAT TREATMENTS .......................................................................................................................................5 BRIGHT BAR .....................................................................................................................................................5

FINER FINISHES............................................................................................................................................6 THE TESTING OF STEEL...............................................................................................................................6

AN OVERVIEW OF ENGINEERING STEELS................................................................................................7 INTRODUCTION ...............................................................................................................................................7 1.0 CARBON STEELS..................................................................................................................................8 2.0 FREE MACHINING GRADES...............................................................................................................9 3.0 CASE HARDENING STEELS .............................................................................................................10 4.0 DIRECT HARDENING STEELS .........................................................................................................11 5.0 STEELS FOR INDUCTION/FLAME HARDENING ..........................................................................12 6.0 NITRIDING STEELS............................................................................................................................12

THE ROLE OF ELEMENTS IN STEELS........................................................................................................13

DIRECT HARDENING STEELS ......................................................................................................................16 DESIGN RELATED TECHNICAL CONSIDERATIONS...............................................................................17 SOME PROBLEMS ENCOUNTERED WITH DIRECT HARDENING STEELS..........................................18

END CRACKING...........................................................................................................................................18 SURFACE DEFECTS/IMPERFECTIONS.....................................................................................................18 HEAT TREATMENT CRACKS DIRECT FROM MILL .................................................................................18 HEAT TREATMENT CRACKING - SUB CONTRACTOR HEAT TREATMENT ..........................................18 FURTHER SAFEGUARDS............................................................................................................................19

ADDITIONAL TREATMENTS .......................................................................................................................19 UNDERSTANDING CASE HARDENING (CARBURISING) .......................................................................20

HISTORY..........................................................................................................................................................20 FACTORS AFFECTING SELECTION............................................................................................................21 DISTORTION ...................................................................................................................................................22 INDUCTION/FLAME HARDENING STEELS...............................................................................................23 CRACKING ......................................................................................................................................................24 APPARENT INABILITY TO HARDEN..........................................................................................................24 LOCALISED HARDENING.............................................................................................................................24 CONCLUSION..................................................................................................................................................24 CRACK DETECTION ......................................................................................................................................24 STEELS WHICH SUIT NITRIDING, CARBONITRIDING AND NITROCARBURISING..........................25 STEELS USED FOR NITRIDING....................................................................................................................25 MACHINABILITY ...........................................................................................................................................26 STEELS FOR NITROCARBURISING ............................................................................................................26 CARBONITRIDING .........................................................................................................................................26 PRE-TREATMENTS ........................................................................................................................................26

MACHINABILITY ADVANCES IN ALLOY STEELS..................................................................................28



Page 3

MACHINABILITY OF ALLOY STEELS........................................................................................................28 OBJECTIVES OF THE STEELMAKER..........................................................................................................29 MACHINING ADVANTAGES OF INCLTISION MODIFIED ALLOY STEELS .........................................30 SPECIALISED MACHINING OPERATIONS ................................................................................................30

Later Developments - mild/carbon steels.......................................................................................................30 ACTUAL CASE STUDIES - WHAT'S THE PROBLEM? .............................................................................31

CASE 1: CUTTER BLADE ..............................................................................................................................31 CASE 2: STEERING BUSH: ............................................................................................................................32 CASE 3: SECURITY SAFE COMPONENT ....................................................................................................33 CASE 4: FINAL TEST RIG FAILURE ............................................................................................................34 CASE 5: PINION GEARS - HEAVY MACHINERY.......................................................................................35 CASE 6: CONTROL RACK SHAFT................................................................................................................36 CASE 7: FILTER PLATE .................................................................................................................................37 CASE 8: SPLINED SHAFT - GLASS MAKING MACHINERY....................................................................38 CASE 9: SLIDE WAY - ENGINEERING INSTRUMENTATION .................................................................39 CASE 10: RACK SUSPENSION STUD...........................................................................................................40 CASE 11: TRACK PIN - HEAVY OFF ROAD VEHICLE (CIVIL) ...............................................................41





Page 4

THE MANUFACTURE OF ENGINEERING STEELS Engineering Steels are manufactured from two sources of raw materials; iron ore or re-cycled scrap.

MELTING PROCESSES

By far the greatest bulk production is by reacting iron ore, coke and limestone with hot air in a Blast Furnace (Fig. 1 ). This produces pig iron, which is then refined in the Basic Oxygen Steelmaking furnace to produce steel (Fig. 2). One such steel plant may have a manufacturing capability of 15,000 tonnes PER DAY.

The alternative production method is by using steel scrap rather than ore.

Unlike aluminium, steel is almost entirely re-cyclable. Steel scrap is graded according to purity, and melted in an electric arc furnace (Fig.3). The liquid steel is then refined and combined with alloying elements in order to make the required specification. Such a process has a greater energy dependence and the steels manufactured by this process are usually more expensive. This process is more flexible and can make smaller quantities, coupled with more sophisticated alloy combinations. In comparison to Blast Furnaces, an Electric Arc Melting plant would probably produce 3000 tonnes per day.

In the last twenty years, secondary steelmaking (ladle furnace - Figs. 4,5) has now complemented both types of steelmaking plant.

The final optional molten metal processing procedure is vacuum degassing (Fig. 6). This is only employed when exceptionally low gas levels are vital and very low non-metallic inclusion levels are required. Such steels give superior fatigue strength properties, which in tum are applied to improve performance of sophisticated high tensile components such as bearings, high performance gear boxes or engine components.

CASTING PROCESSES

More than 95% of steel is continuously cast (Figs. 7, 8) the remainder being poured into ingots. The steel is passed through water-cooled copper moulds and is usually poured into rectangular or square blooms (typical size 350max 250mm), billets (typical 100 mm sq. to 135mm sq.) or slabs (Fig. 9) ( 1200mm x 200mm). Cast products of this type have an extremely coarse crystal structure, has virtually no strength and must be hot worked to break down the structure. Thus the next stage is hot rolling; the hot blooms (700 oC when leaving the casting machine) are re-heated to 1200 oC or 1300 oC in order that rolling can commence.

ROLLING

The steel has to be reduced in cross sectional area, ideally by 7 to 1, or the rolled product is only 14% of the cross sectional area of its original size.



Page 5

TESTING

The product has to be "Fit for purpose" and undergoes a series of tests, depending on the application and quality levels required. It must be stressed that each such test is expensive to carry out and only tests necessary for the application should be called for:

1. Analytical 2. Surface conditions of the billets 3. Internal defects - ultrasonic testing 4. Mechanical testing. Tensile, impact, hardness range, bend tests are fairly

common. Hardenability testing is done mainly for transmission components. Fatigue and creep testing would also be essential in the case of aircraft steels.

Having thus been tested to the properties required by the customer, a certificate of conformity would then be required to be issued in order to fulfil the Standards of Directorate.

FINAL PRODUCT

Such blooms and billets are termed "semis" meaning that there is more processing to be done. This could be in the form of forging, drop forging or further rolling an a smaller ("jobbing") mill, which would roll the billets into rounds, hexagons, squares or flats. This would be termed secondary rolling.

HEAT TREATMENTS

Following secondary rolling, the products are termed "black" (i.e. black bar, black flats etc...). The lower carbon (lower strength) materials are usually unheat treated, whereas medium medium/high and direct hardening alIoy steels would be heat treated. Therefore, the customer has an option to order steel in one of the following executions:

a) As rolled. b) As rolled/reeler straightened. (Fig.10) c) Annealed (softened) d) Normalised (air cooled) e) Hardened and tempered

Black products may vary in size and contain minor surface imperfections. Thus a closer tolerance BRlGHT bar may be sought, albeit at a higher price.

BRIGHT BAR

In short, the black bar is descaled by either chemical (acid) pickling or mechanical (shotblast) means, after which it is drawn through a die (Fig. 11 ) (bright drawn bar) or the surface is "peeled". Peeled bar is more expensive in that approximately 6-11% of the bar weight is lost in turnings. However, this method removes the surface imperfections which are found in a drawn bar (which originate from rolling irregularities previously mentioned). Bright bars are available in rounds, hexagons, squares and flats.



Page 6

NOTE:

Drawn bar has a higher strength due to the deformation imposed when going through the die. Peeled bar rarely * exhibts any increase in hardness from the original black bar hardness previous to peeling.

* Occasional increase in surface hardness due to roller pressure but it is usually controllable and confined to the surface layers.

FINER FINISHES

From the drawn or peeled cold finished bar, it is then possible to grind in order to obtain finer surface finishes. Such finishes are expensive, may constitute further production problems and due consideration and understanding should be given before taking this route.

(Figs.12,13)

THE TESTING OF STEEL

(Figs.14-22)



Page 7

AN OVERVIEW OF ENGINEERING STEELS

INTRODUCTION

Engineering steels, a term given to steels used by design engineers in order to fulfil technical needs of the designed component within the remit of minimum cost.

Minimum cost should be, (but not always so!) related to the actual cost at the point of application of the component coupled with the average life of the component in service.

NOTE: There is little point in buying the cheapest steel per tonne if this constitutes poor machining performance, lack of response to any subsequent treatment and the creation of excessive rejections in the final stages of production.

Hence the objective of this course is to simplify the choice of available options from available materials. In the commercial reality of today there is little point in choosing a steel if it is not available in the required quantities. Unless large tonnages of relatively few sizes are needed, the days when a steelmaker would make 100 tonnes, sell ten tonnes and stock ninety tonnes have gone!

Thus in spite of thousands of steels being published in specification books, tomorrow's engineer is faced with a diminishing range of steels with which to work due to the burden of stocking costs. It is, therefore, vital to know which steels do which "jobs" in order to understand that steels which are stocked will cover most applications. A simple understanding is, therefore, vital before more complicated issues can be approached.

Engineering steels fall into the following categories:

I. Carbon steels II. Free machining grades III. Case hardening steels IV. Direct hardening steels V. Other surface hardening steels (flame, induction, nitriding etc.) VI. Special requirements (low temperature, high temperature grades)

NOTE: The following steels cannot be covered in this course in view of time available. Stainless Tool Steels; Hot and Cold Work Die Steels; High Speed Steels; Spring Steels.

(Fig. 23)



Page 8

1.0 CARBON STEELS

(a term given to steels containing basically carbon with small amounts of silicon and manganese)

In short, the cheapest to produce and the cheapest to buy, but have limiting factors, which have historically led to development of steels which overcame the limitations; each one being at a somewhat inevitable increase in manufacturing costs.

Available carbon steels fall into three main groups, generally classified by carbon range:-

1. Low Carbon (Mild) Steels below 0.25% carbon. The softer steels not necessarily

good for machining. Good for welding and extremely ductile for drawing to small sizes for screws, nails, staples, wire etc.

2. Medium Carbon Generally 0.3-0.45% carbon. Form mid range (500-750 n/mm2) tensile strength options. May be hardened and tempered in small sections. Lower carbon range will weld whilst others require preheating. Possible flame/induction hardening in medium (40-48RC) range. Used for shafts, surface hardened worm screws low duty gears. Boron is sometimes added for improved hardenability.

3. High Carbon Generally 0.5% - 0.8% carbon. Much higher tensile strength. Poor hardenability restricts ruling section. Only lower range (0.5-0.6) practical for general engineering. Above this, uses are small hand tools, rope wires, low grade springs. These steels should not be welded without specialised procedures as their Carbon Equivalent is too high.

(Fig. 24)

Carbon Equivalent is a guide to a steels weldability and calculated by the formula:-

Carbon Equivalent is below 0.35% - readily weldable Carbon Equivalent between 0.35-0.55 - pre heat recommended Carbon Equivalent above 0.55% - preheat and post heat vital Notes: Specialised advice should be sought from welding engineer re filler rods and temperatures needed with Carbon Equivalent above 0.55%.

BS970 080A 15 070M20 (En3B) 150M19 (En14A)

080M40 (En8)

070M55 (En9)



Page 9

2.0 FREE MACHINING GRADES

The addition of sulphur to steel is well known to increase machinability. Note there is a corresponding increase in manganese in proportion to the amount of sulphur. This is to create formation of manganese sulphide particles which facilitate chip breaking, corresponding increased machining speeds and hence lower production time for a given component.

A corresponding addition of lead further increases this.

(Fig. 25)

Limitations

Such additions mean that impact strengths of these steels are inferior to lower sulphur grades and the welding of any free machining grade is not recommended due to formation of brittle networks of sulphide which in turn can cause cracking.

Available free machining steels fall into three main groups, with slightly different classes than carbon steels:

1. Low Carbon (Mild) Steels below 0.15% - possibly the largest group of steels produced for engineering purposes. Applications are for components where little strength would be required e.g. small instrument parts, pressure switch bodies, low grade fasteners, screws, hose couplings etc. They may be case hardened for increased surface wear resistance.

2. Medium Carbon - non heat treated (supplied "as drawn") Generally 0.4-0.45% carbon, the higher sulphur and corresponding manganese (hence tensile strength around 10% higher) mainly due to the high manganese/carbon. Uses, low grade gears, lead screws, steering/machine racking, steering pump shafts etc. They will induction harden in 45/52RC range*.

3. Medium Carbon - heat treated (hardened & tempered) grades Generally 0.4% carbon with higher manganese/sulphur. These will harden and temper as high as 850-1000 n/mm2 606M36/USACUT 55* range. There is, however, a "penalty" of reduced impact strength. They will induction harden in the 45-S2HRC range *.

4. Medium Carbon/Heavy worked high tensile grades Materials with superior yield/ UTS ratios due to special drawing techniques involving patent die designs coupled with greater cold or warm reduction ratios.

* See notes on induction hardening.

230M07(En 1A)/ 230M07Pb (En 1 A Pb) 210M15 (En32M)

212A42 (En8M) 226M44 (HITENSPEED/ USACUT 45)

SAE 1144



Page 10

3.0 CASE HARDENING STEELS

High dynamic surface loading of a component requires a correspondingly high surface hardness of controlled depth, backed by a pre-determined core strength. Thus a range of steels evolved which were relatively easy to machine before case hardening, but developed the required case/core properties by controlling alloying elements (core strength) and case hardening practice (case hardness and depth).

Because of such a variation in HERTZIAN (dynamic) loading stress, a range of steels to satisfy low loading conditions (mild steels) to high loading (steels with up to 5% total alloy content) are available. In the interest of economy, steels should be selected to satisfy the necessary properties, with minimum cost, in order to produce a given component.

080M15 (En32B) No added alloy D.I.=0.52 c/s 350-450 n/mm2 665M17 (En34) Alloy content 2.0% D.I.=1.13 900-1050 n/mm2 635M15 (En351 ) Alloy content 1.6% D.I.1.5 900-1000 n/mm2 805M20/SAE8620 Alloy content 1.4% D.I. 2.3510'50-1150 n/mm2 * 655M13 (En36) Alloy content 4-4.5% D.I. 4.141150-1250 n/mm2

*Higher core strength developed by slightly higher carbon in original steel.



Page 11

4.0 DIRECT HARDENING STEELS

Steels which are generally hardened by heating to austenitising temperature (8100C - 920 oC), soaking at temperature for the required time, quenching in oil, water or polymer (oil/water emulsion) and tempering to the required hardness range. None of these steels are weldable unless specialised techniques are strictly adhered to.

These steels generally fall into two ranges of carbon content but additions of nickel, manganese, chromium, molybdenum are commonly used in different combinations. It is also possible for relatively small additions of other elements for various additional benefits:-

a) Vanadium

b) Titanium

c) Niobium

V and Ti increase nitrided layer hardness. Specialised high temperature e.g. bolting grades for power stations.

d) Aluminium (specialised nitriding steels with high hardness in nitrided layer)

a) Medium tensile steels (up to l000n/mm2) Carbon varies from 0.25 up to 0.45% generally. The

ruling section/maximum size that properties can be achieved is usually controlled by alloying additions - higher additions give greater ruling sections.

b) High/ultra high tensile steels (Up to 1550 n/mm2 dependant on ruling section)

APPLICATIONS

Medium/high performance requirements such as structural, engine and transmission components. Steels may be rolled, forged, drop forged before heat treatment and final machining. Surface hardening techniques (flame/induction) are possible BUT post stress relieving (i.e. heating to 100 to 200 oC for at least 1 hour is VlTAL in order to reduce possibilities of cracking.

Note: Such high tensile levels have greater "notch sensitivity" (i.e. prone to embrittlement). This may thus require to undergo further expensive processing (vacuum arc re-melting (VAR); electro slag re-melting ( ESR) in order to improve this. )

605M36 R, S 708M40 R, S, T 817M40T/Annealed

817M40 U to Z (En24) 826M40 U to Z (En26) (bigger ruling sections than En24)



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5.0 STEELS FOR INDUCTION/FLAME HARDENING

When outer layers of steel are rapidly heated to austenitising (usually above 800oC) temperature and rapidly quenched to give increased hardness in the outer layers of the components. Such practice is economic in components manufactured in high numbers; an example would be automotive components (steering pump shafts where contact bearing surfaces prevail).

Carbon is the vital element for the rapid hardening and only steels with carbon contents between 0.4 and 1% are usually applied to the process. Hardness increases progressively with increases in carbon, together with greater risk of cracking and distortion. Such risks can be anticipated and avoided with good practices.

Sulphur bearing steels with like carbon levels can also be hardened, but risks of cracking can be greater.

080M40 (En8) 212A42 (En8M) 226M44 606M36/USACUT 55 605M36 (En16) 708M40 (En19) 817M40 (En24) 535A99 (En31 )

6.0 NITRIDING STEELS

Steels which respond to nitriding - a relatively low temperature (500 - 550 oC) surface hardening process involving exposing the surfaces of the components to nitrogen ions, which diffuse into the surface causing a layer of alloy nitrides. This process also gives less distortion to the components, compared with iii to v. Such steels should ideally, be low (around 0.3%) carbon and contain chromium, molybdenum, vanadium and sometimes aluminium.

Such steels generally fall in the medium stress performance areas, but exhibit advantages of moderate corrosion resistance and improved performance in service at high (up to 500 oC) temperatures compared with classes üi to v.

708M40 (En 19) 817M40 (En24) 722M24 (En4OB)



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THE ROLE OF ELEMENTS IN STEELS 1. CARBON (up to 1 %) The very basis of steels. Always present as Fe3C but in

various forms - some softer (pearlite) than others (bainite or martensite). The more carbon present, the steel is generally harder.

2. SILICON (up to 0.4% in general engineering steels) Special steels up to 2%) Present in most steels (not necessarily all) in small quantities (up to 0.4%). It is a deoxidiser and an increase tends to increase the steels hardness. SPRING STEELS have 0.7-1% silicon, other steels do not, except for special applications such as valve steels (up to 2%).

3. MANGANESE (0.5-1.7%) Vitally present to combine with sulphur to form manganese sulphide (MnS) which lessens sulphur's embrittling effect. (Without manganese present, iron sulphide would form, giving hot working problems and embrittlement during use.) Further additions render higher strengths coupled with good toughness, particularly in low temperature (-40oC) applications.

4. SULPHUR In truth is undesirable in engineering steels but is widely added to improve machinability coupled with a proportional increase in manganese. Such sulphur additions increase tendency to embrittle and should not be welded by any process.

5. PHOSPHORUS Inevitably present as an impurity, up to 0.04%. Special applications (aircraft) limit to .O10% with high consequent cost. This has an embrittling effect on the ferritic phase in steel. It is not visible under normal microscopic conditions.

6. LEAD (up to 0.3%) At this stage of freemachining steels, additions (0.3% max.) of lead greatly improves machinability (giving an increase of 30% on sulphurised grades); lead does not usually exist in free form in the micro structure, but associates with the manganese sulphide inclusions in freecutting grades). It is environmentally sensitive and must be added under strict conditions. It is currently only added to the free machining mild steel grade (230M07). Availability in other grades is rare.

7. NICKEL (up to 2.1 /2%) Is added up to 2.1 /2% generally with 1% alone greatly increasing tensile strength, impact toughness and low temperature mechanical properties. It is usually (but not always) combined with a chromium addition and often



Page 14

further molybdenum addition also. Increase in nickel content increases the hardenability of the steel.

8. CHROMIUM (0.5 - 3.0%) Great benefits abound by chromium additions in engineering steels for differing reasons to nickel. Chromium is a powerful carbide former which:

I. Increases the strength and toughness II. Improves high temperature tensile strength (up to

300 oC). III. It is also a powerful nitride former, thus lending

itself to nitriding applications. IV. Combinations of molybdenum and vanadium (up

to 0.5% of each) also enhances the high temperature performance in addition to higher surface hardness after nitriding.

9. MOLYBDENUM (0.2 -1.0%) As stated, added up to 0.5% with occasional 1 %. Greatly improves hardenability, strong carbide former, also forms nitrides and tends to improve properties at elevated temperatures.

10. VANADlUM (0.20 - 0.50%) Another carbide former, good at creep (extension under load at high temperature) resistance, due to its very finely divided nature in carbide form - appears as "clouds" in microstructures Another boost to nitriding processes.

11. BORON A lesser used element in engineering steels but greatly enhances hardenability in straight carbon steels. Ideally used in heat treated medium tensile bolts. An addition of .0008 - .005% is equivalent to 0.5 -1 % of other alloying elements. There are other limitations in properties and also in production.

12. ALUMINIUM May be used as a deoxidiser if silicon free steel is required. It is added in small quantities (0.025% max. ideally) to steels for grain size control in order to obtain the necessary mechanical properties. However, excessive nitrogen levels can cause ultrafine grains, which are detrimental to hardening or induction hardening (particularly in plain carbon steels with no compensatory alloy additions).

13. COPPER An impurity derived from the use of recycled scrap with adherent copper based components. The element cannot be removed by refining and consequently any copper added is reported as an impurity in the analysis report. Copper, along with tin, can adversely affect the hot reduction processes (rolling, forging) causing hot bursts (hot shortness) on the surface.



Page 15

14. TIN Derived in steels in similar way to copper. The effect of tin on mechanical properties is that it is more powerful than copper. i.e. small amounts can affect hot working. Tin and copper together can affect impact strength, hence their control is important. They cannot be removed by refining during steelmaking, thus control of incoming raw materials is very strict.

15. CALClUM Small amounts of calcium (usually in the form of calcium silicide powder) are sometimes added to steels prior to (and sometimes during) the casting process. It is primarily added to increase fluidity in the casting of steel. Secondly calcium silicide injection techniques are used with other compounds in order to modify (soften) abrasive alumina based inclusions for machinabitity improvements.

16. NITROGEN inevitably absorbed from the atmosphere during the melting process. Generally desirable to be less than .010% in general engineering steels and .007% in welding grades - particularly plate.

17. OXYGEN Likewise absorbed during melting, but being more active, will combine with most elements present to form oxide inclusions, some of which are hard (aluminium oxide) and some relatively softer (compound mixtures of manganese, silicon, iron etc.) Whilst there is a need for close control for high performance engineering components, their normal levels in general engineering applications do not cause many failures.



Page 16

DIRECT HARDENING STEELS This term describes steels which are capable of being quenched from high temperature, resulting in a controlled/predictable increase in hardness throughout the section of steel in question. Normally the carbon range lies between 0.35% and 0.55% with certain special exceptions to this being spring steels (0.6-0.75%C), bearing steels (0.9-1.1 %C) and tool steels (0.8 - 2%C) but do not feature in this seminar.

Alloy contents vary according to application but are generally in a total alloy content level of 1-5% with a general rule that the higher the content of alloy the greater is the HARDENABILITY.

General Rule: Carbon determines the attainable tensile strength, alloys determine the depth to which this tensile strength may be achieved.

Such steels must be processed by quenching (i.e. rapid cooling from high temperature) followed by tempering, (secondary lower temperature treatment) in order to achieve the desirable combination of mechanical properties. (Fig.26)

Quenching involves accelerated cooling from a pre-determined (austenitising) temperature and this is usually air, water, oil or polymer (combinations of oil/water emulsions).

Tempering is then carried out at a pre-determined temperature in order to attain the desired properties, which should have been previously specified by the design engineer of the component in question. These temperatures vary generally between 550oC and 650oC and involve a time for heating to temperature and then 2-3 hours (usually depending on steel grade) at the temperature required.

Fig. 27 shows an example of a tempering graph for BS970 1983 708M40/SAE/AISI 4140/42CrMo4.

NOTE: Both quenching and tempering temperatures vary with each variation in steel analysis. ALL have been pre-determined by research and are well known to heat treatment subcontractors. Further detailed information may be obtained in large reference libraries if required.

RULING SECTION - must be clearly understood and this is plainly influenced by alloy content. A general rule is that limiting ruling section increases with corresponding increases in ALLOY content. Some alloying elements though have a greater effect than others and nickel and molybdenum play a greater part than does chromium.

BS970 does play a role in guiding the selection by indicating both minimum and maximum diameters in millimetres to which a range of mechanical property options may be obtained. This is termed "Ruling Section". Many publications illustrate this and are available for the engineer.



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DESIGN RELATED TECHNICAL CONSIDERATIONS

The decision on which steel to select may lie on many other considerations. Even when a seemingly correct decision has been made, have all other factors been which may influence these properties within a component.

a) Does the application involve working in corrosive environments? This may well affect the life expectancy of the component. If this is undesirable then means of extending this life must be considered:-

I. Applying a deposited protective layer - e.g. electro plating, phosphate or PTFE II. Consider a thermally applied coating - e.g. nitriding. üi] III. If these are not practical, reconsider the material to be used.

b) Will the component be used at elevated temperatures? If so, what is the maximum temperature and how long will this be sustained?

I. Long times at temperature - is there a danger of fatigue? II. How will the component perform? Tensile strength falls with increase in tensile

strength; what will the l.ITS be at the working temperature? III. Does the material have to be reconsidered?

c) Will the component need to work in low temperature conditions? If so, will there be any impact strength problems?

NOTE: High/low temperature data is available and guidance can be sought IF the problem is anticipated before the steel is used.

d) Are there any areas where high local loading or contact stresses may prevail? e.g. contact with roller/ball bearing rolls will be a source of high wear:-

I. Should localised hardening techniques be considered? (e.g. flame or induction hardening) to sustain these loads.

II. Should a stress relieving operation be carried out after induction hardening? III. If this should be so, do we need to crack detect the component after induction

hardening? IV. Will the interface between hard/soft areas be a problem?

NOTE: Such surface hardening requirements should be added to the drawing showing:- * The area to be hardened * * The hardness (specified in hardness units) required. * * * The depth at which a minimum hardness is attained.

V. If such a hardness is not attainable then another steel with superior induction hardening properties would be needed and the selection process would need to be restarted.



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SOME PROBLEMS ENCOUNTERED WITH DIRECT HARDENING STEELS

END CRACKING

Bars to be direct hardened usually have sawn ends. These sawn ends may radiate quenching cracks due to their being stress raisers. These rarely extend more than 50mm into the bar and this is usually the "bar end" discard (scrap) anyway. Buying cut-to-length blanks invariably means that they rarely occur.

SURFACE DEFECTS/IMPERFECTIONS

Both black and bright drawn bars when purchased are NOT defect/imperfection free. Clear understanding of defect/imperfection levels should be sought before deciding on the size of bar for a component in order that they are moved by machining. Alternatively, peeled bar may be bought but this carries price penalties. Consequently, removal of inherent defects is usually the most practical option to take.

HEAT TREATMENT CRACKS DIRECT FROM MILL

There are no requirements in international standards to crack detect bars. Consequently, unless requested on ordering (with a price premium) the bars are not guaranteed crack free. Any such heat treatment crack is large and usually extends from the outside of the centre of the bars. Such defects are not common with a good supplier and with an incidence usually less than 0.5%.

HEAT TREATMENT CRACKING - SUB CONTRACTOR HEAT TREATMENT

Subcontractor heat treatment is usually adopted when the steel purchased does not possess the final specified mechanical properties. There may be several reasons in direct hardening steels, but the two main ones are:-

I. When the steel needs to be machined at a lower tensile/hardness level for reasons of economy.

II. When in any event a steel with the required tensile/hardness is simply not available.

All too often components appear to have cracked on return and the principal defence argument is that the cracks were in the material before heat treatment. This "no win for all" situation should be avoided at all costs by:-

a) Auditing the subcontract heat treatment company. b) Specifying the EXACT procedure required (a copy of which is sent with the order)

stating:-

I. Max./min hardness levels required. II. Preheat procedure - NO direct hardening steel should be placed directly

into a hot furnace. Critical heating (650oC-800oC) should be slow to allow volume expansion stresses to equalise.



Page 19

III. Specify immediate temper after hardening (i.e. components at least 70-l00oC minimum).

IV. Automatic request of photocopies furnace charts for every heat treatment batch.

V. Send certified crack detection details with the batch!

NOTE: These matters generally induce more attention during processing. Another point could be to involve the heat treatment company, listen to their suggestion, ask for confirmation and thus add to the order; in that way they are working to their recommendation.

FURTHER SAFEGUARDS

If large numbers of relatively valuable components are to be produced, heat treatment problems are very costly. Therefore, there may be further considerations which reduce the risk of cracking:

a) Order crack detected (and certified) bars from the supplier. b) Crack detect the components after machining. c) When large numbers of components are involved, consider two contractors, if

cracking problems have been experienced in the past. In that way it is a control which can help if one sub contractor cracks the bars and attempts to point the fault at the steel supply.

NOTE: The cheapest heat treatment may prove to be the most costly in the end!

ADDITIONAL TREATMENTS

There may be additional requirements in order to improve fatigue life in direct hardening steels. These usually involve additional hardening or inducing compressive stresses into the surface layers of components:

1) Induction Hardening - applied to increase surface hardness. This may be applied locally (i.e. where a particular area would be subjected to wear - splines or surfaces in contact with bearings being examples) or entirely along the length in some cases.

2) Shot Peening (laser peening) - Components subjected to controlled bombardment of shot have an increased fatigue resistance. The shot imparts compressive stresses in the component surface and has been proved to be extremely effective in reduction of component fatigue failures.

3) Grinding - Improved surface condition by grinding to a closer surface (cla) finish will improve fatigue life. Great care must be taken to avoid overheating at all costs as grinding cracks, or softening of the surface, will most definitely form if disciplines are not followed. It is strongly advisable to crack detect, following grinding operations, particularly in safety critical components. Grinding cracks follow a characteristic, different to all other cracking, in that they usually are "crazy paved" in nature or transverse to the longitudinal axis of the component.



Page 20

UNDERSTANDING CASE HARDENING (CARBURISING)

HISTORY

From ancient times, blacksmith's discovered that if a soft iron implement was left in the hot charcoal/wood fire for a night and water quenched from red heat, they would become very hard. So hard in fact that fine edges could crack easily. It, therefore, became practice to return them into a cooler part of the fire for one or two hours and this would prevent cracking, but retain most of the hardness. Swords, axe heads, ploughshares and other tools were made in this way.

They had, in fact, carburised hardened and tempered the steels (but probably never knew this!).

Solid (activated charcoal) carburising has been largely superseded by liquid (cyanide) and gas techniques, the liquid one also having greatly diminished for obvious environmental reasons.

These modem techniques involve furnaces equipped with sophisticated computerised gas/temperature control, giving pre-calculated diffusion of carbon into the component surfaces. The depth of case is dependant on time/temperature/gas control but in the case of continuous operations (i.e. automotive transmissions, shafts etc.) an amazing degree of consistency can be achieved. Thus a furnace containing components of similar size and composition will indeed have similar case hardnesses and core strengths. (Fig. 28)

A furnace containing different carburising steels, however, will doubtless have differences in case hardnesses and core strengths. This variation is completely due to differing chemical analysis, possible variation in grain size, both of which in turn affect the hardenability and transformation characteristics.

Imagine, therefore, three gears of the same size and design, but different steels, carburising together in the same furnace, temperature gas composition and time. All are quenched in oil and tempered at I8OoC. After cooling they are sliced across the axis and hardness tested at 5mm intervals across the section. Figs. 29-36 show the wide contrast in case and core hardnesses.

Note that steels with softer cores, generally have harder cases than alloy steel cases, unless additional techniques are used; although this does not necessarily mean that performance is impaired - hardness differences referred to are marginal and post treatments (i.e. refrigeration) can increase this if it is required.



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FACTORS AFFECTING SELECTION

The engineer should be able to take information needed from drawings or from calculations. What is basically required is:

I. Depth of case and hardness requirement II. Core hardnesses required at specified distances below the surface. These

figures are based on hardenability details which are obtainable (often at extra cost) on request. Altematively, testing houses carry out a test on the actual cast.

NOTE OF CAUTION: Hardenability tests are conducted by water quenched hardness figures. Case hardened steels are usually oil quenched (sometimes polymer) and therefore actual hardnesses will be lower. Again, there is no substitute for sectioning the case hardened component for true values.

Meaningful comparison, as to how a particular steel will perform within a given specification, can be done by calculation. Much argument has taken place, but practical experience shows that the calculation of D.I. value of a given steel can be useful:-

D.l. Value =

where f(C)

(when Mn<1.2%) (when Mn>1.2%) (when Ni 1.80%)

(when Ni > 1.80%)

For example, every steel has a range of analysis in which there is a theoretical high range (i.e. where every element is on top level). Likewise, there is a theoretical low range (i.e. where every element is on the bottom level).

Steels generally lie towards the mean, but if several casts are to be chosen from, the D.I. value might well help.

In the case of a small diameter shaft, a low D.I. value would suffice, but in the case of a larger diameter, where better hardenability is sought, a higher D.I. value should be used.

The engineer can easily find this by requesting the cast analyses of all casts available in that size and calculating the values using the above formula.



Page 22

DISTORTION

Invariably the steel is blamed, but the probable truth is that less than 10% of distortion is from the steel. Often (unnecessary) normalising (i.e. air cool from 850-900oC) treatments are requested and distortion may still take place. It is, therefore, important to list factors which may affect distortion and ask questions of the heat treatment contractor* relating to these:-

a) Time, temperature and control of gas composition are vital. Any variations can lead to variation in case depth, carbon level and hardness.

b) Quenching oil properties vary and should be examined. Life of oil is vital and should not be allowed to age.

c) Temperature of quenching oil, before and after the quench, should be monitored. Does the oil return to original temperature before the next quench?. . . Are oil coolers in operation? - if so are they 100% operational?

d) Directional circulation of oil can play a significant part. Is it taken into consideration? If so, are circulation pumps operating effectively?

e) Are the components stocked in the furnace in order to minimise distortion i.e. long products of more than 3 x D should be quenched vertically as opposed to horizontally.

f) It should be appreciated that quench hardening of steel involves volume change. Many factors effect this, but generally hardening causes a volume expansion. Thus some dimensional change is inevitable.

NOTE: Components of high volume should be measured before and after case hardening. If carefully controlled disciplines are followed, such changes should be constant and predictable - even to the point of changing component dimensions to meet these changes - provided D.l. values are also considered.

* Many heat treatment contractors operate very sophisticated, internally audited procedures. Such modern plant often lose out to the lower priced contractor with greater problems arising.

080M15 (En32B) No added alloy D.I=0.52 c/s 350-450 n/mm2 665M17 (En34) Alloy content 2% D.I.=1.13 900-1050 n/mm2 635M15 (En351) Alloy content 1.6% D.I.=1.5 900-1000 n/mm2 805M20/SAE8620 Alloy content D.I. =2.35 1050-1150 n/mm2 655M13 (En36) Alloy content 4-4.1/2% D.I. =4.14 1150-1250 n/mm2



Page 23

INDUCTION/FLAME HARDENING STEELS

These processes do NOT change the chemical analysis of the surface layers, both processes involve rapid heating of the component surface and quenching in order to form a hardened wear resistant layer of controlled depth and hardness. The surface hardness is thus controlled by selection of steel with the requisite carbon level.

Carbon is by far the most influential element and the maximum attainable hardness varies from 45RC up to 64RC. Such hardness levels are attained by medium (0.40%) carbons up to hyperutectoid ( 1 % carbon). Alloying elements play a part of increasing the depth of the layer (as well as the depth being also controlled by induction coil frequency). A broad guide of hardness levels achievable would be in the following table.

DESIGNATED STEEL SPECIFICATlON

PREFERRED CARBON RANGE

ACHIEVABLE HARDNESS LEVELS (RC)

080M40 (EN8)* 0.40-0.44 45-50

070M55 (EN9) 0.50-0.60 55-60

605M36 (EN 16) * 0.36-0.40 52-57

708M40 (EN19) 0.40-0.44 52-57

817M40 (EN24) 0.40-0.44 52-57

535A90 (EN31) 0.90-1.05 59-64 * Free machining (high sulphur) versions of these alloys also may be hardened, but are more notch sensitive.

Case depths - vary from 0.5mm to 10mm normally depending on what is required.

NOTE: Volume changes occur in the hardened layers, these are due to martensitic transformation. Such changes can affect the length of shafts.

One typical example being a 60mm bar of 535A99 (En31 ) from which a roll 2.3 metres long was manufactured. After induction hardening the roll was 1 mm shorter!

This feature is not always the case; a combination of carbon content of steel, depth of case, diameter of bar, may even cause growth.

SPEClAL NOTE:

One advantage though is that once this is fully understood, then the growth/shrinkage is entirely predictable. - Large production runs of manufacturing a particular product can render close control IF the incoming material is carefully specified:-

a) Method of manufacture (Electric Arc? Blast Furnace?) * b) Close control of composition (tighter limits than entire range) * c) Close control of pre-heat treatment (i.e. closely specified hardness limits on the "as

received" bars) * d) Close dimensional tolerance *

* Such parameters often mean that the base material carries a price premium. However, elimination of rejections due to soft spots, uncontrolled growth and expensive rectification at the finished stage more than compensates for the purchase price premium.



Page 24

CRACKING

Induction/flame hardened components have high internal stresses. The higher the carbon, the higher these stresses are, hence it is vital to stress relieve WITHOUT DELAY! Low temperatures are sufficient ( 1 hour at 160oC) without losing the hardness appreciably.

APPARENT INABILITY TO HARDEN

Many times engineers complain of lack of hardening response and immediately blame the steel. Before this should go any further it is suggested that the following SIMPLE principles should be understood.

I. Sufficient temperature is essential - this varies with carbon content and advise should be sought. Wrong temperature lack of response.

II. Although rapid heating, sufficient time is needed to dissolve the carbon in the austenite (high temperature phase).

NOTE: Chromium bearing steels feature chromium carbides. These are more difficult to dissolve and more time/temperature MAY be needed.

LOCALISED HARDENING (i.e. Gear teeth, bearing areas on shafts)

Localised hardening is possible leaving the remainder of the component in the original condition. Special consideration should be given to the consequences of this; in cases of high tensional stress, environments failure can begin at the hard/softer interface i.e. the "peak" of the torsional stress. One series of failed shafts (15mm dia. x 120mm long) showed such a failure. This was easily averted by a change - by induction hardening the entire shaft at little extra cost!

CONCLUSION

These factors, when taken into careful consideration BEFORE production begins, render induction/flame hardening to be of great use in producing low cost, high duty components.

If greater understanding was applied, these processes would be used to far greater advantage and involvement of the relevant sub contractors at the earliest stage possible would be an invaluable asset.

CRACK DETECTION

All such processes court positive potential dangers of cracking. Crack detection should, therefore, be seriously considered, particularly after finished grinding. If disciplines are rigidly followed, however, the risk of cracking should be minimised and statistical random sampling should suffice.

It is worthy to mention once again:

Increase in carbon = increase in hardness = increase in likelihood of cracking.



Page 25

STEELS WHICH SUIT NITRIDING, CARBONITRIDING AND NITROCARBURISING

There is a separate B.G.A. course for surface treatment processes which does involve these processes in excellent detail. However, it is important to have a brief, clear understanding of the different chemical mechanisms under which these processes achieve the surface hardness increase. Likewise, it is equally important to understand the times, depths of case achievable and any other features.

Surface Treatment

Treatment Temp. Range

Duration of Treatment

Depth Ranges of

cases Produced

Hardness Range

Ideal Steels for Treatment

NITRIDING (GAS) 490-560 0C UP TO 90 HRS

.05-0.75MM 550HV -1150 HV *

STEELS WITH 0.20-0.4% CARBON WITH CR, MO,V AND A1 ADDITIONS

NITRIDING (PLASMA)

400-590 0C UP TO 20 HRS

UP TO O.3MM 550HV -1150 HV *

STEELS WITH 0.20-0.4% CARBON WITH CR, MO,V AND A1 ADDITIONS

FERRITE NITRO- CARBURISING (SALT)

560-580 0C 1/2 TO 5 HRS.

.012-.020MM 350-650- MICRO HV

MOST STEELS

FERRITE NITRO- CARBURISING (GAS)

560-590 0C 1/2 TO 5 HRS.

UP TO 1.5MM UP TO 350 HV

MILD STEELS. CASE HARDENING STEELS. CAR COMPONENTS

AUSTENITIC NITRO- CARBURISING GAS)

800-940 0C 5 HRS. UP TO 1.5MM UP TO 800 HV

MILD STEELS CASE HARDENING STEELS

CARBO- STEELS NITRIDING (GAS)

800-940 0C 5 HRS. 0.1-0.76MM MOST STEELS

CARBO- NITRIDING (CYANIDE)

800-950 0C 5 HRS. 0.5-L.5MM UP TO 800 HV

MOST STEELS

STEELS USED FOR NITRIDING

Steels for nitriding logically contain elements for which nascent nitrogen has an affinity in the requisite temperature range. Carbon is required in order to give core strength, but in fact is an obstruction to nitrogen diffusion. A balance has, therefore, to be struck, hence low carbon/high chromium (722M24) gives a harden case than higher carbon/lower chromium (708M40). It can thus be seen that classified nitriding steels are modelled on this principle.

NOTE: Nitriding 708M40 hardened/tempered to `T' condition gives lower surface hardnesses after nitriding than the same alloy in the 'V' condition. This is due to lower tempering temperatures for 'V' which suppress formation of chromium carbide, thus allowing more chromium availability for nitrogen. This is only a moderate increase and should not give the impression that hardnesses over 600 VPN can be easily achieved.



Page 26

MACHINABILITY

Nitriding steels are notoriously bad for machinability. This is due to the elements present which are not only strong carbide formers but also chemically hard themselves. They are usually improved by heat treatment - FULLY hardened and tempered:-

a) Austenitise at a high (920-950 0C) temperature to dissolve the carbides in the austenite.

b) Temper at a high (700-720 0C) temperature to obtain a uniform distribution of carbides throughout the structure.

NOTE: Whilst this may not remove all machining difficulties, it is a step forward, by removing coarse carbide deposits formed after hot rolling/forging.

STEELS FOR NITROCARBURISING

Unlike nitriding, the nitrocarburising layer is NOT for load bearing. It is to provide a good resistance to wear, galling, scuffing, seizure. It also provides a porous surface which readily accepts further aesthetic treatments i.e. chemical blacking and polishing after which gives good corrosion resistance coupled with low coefficient of friction. From mild steel washers to alloy steel tappets the surface layer's compressive stresses offer improved fatigue resistance in addition to the other features. Pressed components in mild steel are undergoing nitrocarburising treatments, giving strength and rigidity. e.g. automotive bumpers, brackets.

IN SHORT: All steels may benefit by nitrocarburising - but they won't all nitride!

CARBONITRIDING

As in nitrocarburising, all steels may possibly be treated but as this process is carried out at austenitising temperatures, greater case depths can be achieved. The gaseous carbonitriding involves introduction of nascent nitrogen (i.e. cracked ammonia) into the carburising gases. Such introduction increases the hardenability of the carburised layer and accelerates case depths in steels with no alloying elements.

The liquid form of carbonitriding is by salt bath (sodium cyanide/sodium carbonate) in which the components are immersed in temperatures of 800-950 0C. Very deep cases may be obtained by this method.

PRE-TREATMENTS

It is important to have the required surface finish prior to surface hardening, such that post machining temperatures (i.e. grinding) may be eliminated or absolutely minimised (grinding merely removes the hardened layers ).

"STOPPING OFF" - apply a chemical compound coating to surfaces not requiring to be hardened. Such coatings are metallic suspensions in solvent base. Carburising stop off compounds are based on copper suspensions, whereas nitriding stop off compounds are tin based.



Page 27

Ineffective stopping off is expensive - hard spots formed cause tool breakages or impossible machining situations, therefore, an audit * of the requisite stop off operation is vital on the following guidelines:

• Is the compound fresh? - If not - scrap it and start again.

• Has the stop off compound been mixed? Has the sediment been fully agitated?

• Is the application brush good? Splayed/hard bristles - scrap it!

• Is the applied coating even? Uneven coating means problems.

• Is the applied coating completely dry before going into the furnace? If it isn't, there is a danger of blowholes/blowing off as the solvent boils out.

• Check if operative applying the coating has been trained.

* Visiting Quality Managers tend to audit paperwork systems/records, but their audits rarely extend to stop-off practice.



Page 28

MACHINABILITY ADVANCES IN ALLOY STEELS The advent of ladle metallurgy commenced with the development of the SKF-ASEA ladle furnace (Fig. 37) in the early 1970's. The need for this development was to reduce expensive electric arc steelmaking; using the arc furnace for melting/preliminary refining and diverting the semi refined steel into the less expensive ladle furnace. By application of two ladle furnaces in parallel with an arc furnace meant that capital cost reduction was possible without loss of tonnage produced. The "birth" of the ladle furnace was therefore much welcomed and within fifteen years almost every alloy steel production plant in the world employ ladle techniques.

Ladle steelmaking techniques enable the following processes to be undertaken:-

I. Careful control and equalisation of temperature by arc heating couple with induction stirring.

II. Introduction of inert gas stirring. III. Addition of final alloying elements. IV. Advanced metallurgy by feeding highly exothermic oxide/silicide mixtures in

order to modify the morphology of inclusions.

MACHINABILITY OF ALLOY STEELS

Machinability of any given alloy steel can vary dramatically. Factors which affect this are:-

I. Chemical analysis within the specification. II. The condition in which the steel is required:- as rolled; annealed; normalised;

hardened and tempered; tensile strength; surface condition. III. The morphology (chemical content coupled with crystallographic forms) of the

inclusions.



Page 29

OBJECTIVES OF THE STEELMAKER

The objectives of the steelmaker focusing on better machinability are:

a) Control of critical elements within a specification - by keeping carbon, chromium, molybdenum and other carbide forming elements towards the lower end of specifications; by intentionally controlling sulphur towards the upper half of the specified level.

b) Inclusion modification - Grain control is vital to achieve requisite mechanical properties, hence an aluminium is inevitable. Such addition is a potential danger to machinability as the oxide found is very hard and possesses a high melting point. If alumina (oxide) contents are high, both fatigue strength and machinability are adversely affected, hence this is chemically "modified".

A feedwire, containing mixture of lower reactive oxides and calcium silicide is introduced. The thermo chemical reactions result in exceedingly high temperatures, the resultant effect being:-

mpt 2400 oC mpt 1250-1350 oC

hard white "coral-like" morphology

A "softer", lower melting point class to which manganese sulphide also combines to advantage

(Figs. 38-40)

Past research and market development has seen inclusion modified steels develop dramatically. Since a 1980 beginning, well in excess of 1,000,000 tonnes are now produced.

[ ][ ] ( )( )( )( )23232 SiOCaOFeOOAlCaSiFeOOAl ⇒+



Page 30

MACHINING ADVANTAGES OF INCLtISION MODIFIED ALLOY STEELS

Whether direct hardening, case hardening or higher tensile applications increase in turning and drilling speeds between 40% and 100% could reasonably be expected. All significant usages should be preceded by measured trials in order to re-educate machine operators and tool setters to feed/speed combinations which simply are outside their understanding based on conventional machining speeds.

(Figs. 41-44)

SPECIALISED MACHINING OPERATIONS (Milling, Hobbing, Shaping, Broaching, Reaming etc...)

In specialised machining areas, as above, some increase in speeds may be achievable, but the principal objectives should really be to extend the life of the expensive tool. However, there have been many cases when the difference has been dramatic.

What should be enjoyed from inclusion modified steels?

• At least 20% reduction in production costs; achieved by increase in speeds up to 100%.

• Greater consistency in product - giving confidence to increase speeds on a constant basis.

• Resultant confidence giving company estimators the confidence to quote lower prices in order to become more competitive, without giving away all the savings. Hence more competitive + greater profit must be the direction in which to go!

LATER DEVELOPMENTS - MILD/CARBON STEELS

Engineers are persistently frustrated by poor machinability in mild/low carbon steels, which invariably are produced from the iron ore (low impurity) routes. Poor turning, very poor drilling (built-up-edge problems - particularly when using high speed steel drills) mean low productivity and poor profit margins.

Significant progress has been made in order to produce inclusion modified mild and carbon steels. Whilst time savings are not so significant, reduction in down time for drill, tool re-grinding means, once again, reduction in production costs.



Page 31

ACTUAL CASE STUDIES - WHAT'S THE PROBLEM?

CASE 1: CUTTER BLADE

BS970, OSOM4O 1991 (EN 8) - FLAT BAR 6OMM X 2OMM

PROBLEM: After considerable machining, the cutter blade did not respond to flame hardening - achieving only 38RC when previous experience had shown 45-50RC to be achievable.

Analysis showed that an excessively high aluminium prevailed which resulted in minute grains (10-12 ASTM rating) when 5-8 would have given better results. Also the analysis showed that the steel had been made via the Blast Furnace/B.O.S. route (i.e. very low residual Ni, Cr, Cu, Sn) hence there was no extra hardenability which would compensate to a degree for the high aluminium.

N.B. Blast Furnace/B.O.S. route - total "residuals" around 0.2% Basic Electric Arc (i.e. scrap charged) route - total residuals around 0.6%.

SOLUTlON: Only practical answer is to request analysis before purchase and also do a flame hardening test on a small sample before manufacture of the component.

If no aluminium level available, insist on obtaining figure from the original supplier (steelmakers always have this but do not put it on most certificates of analysis!)

SPECIAL NOTE: Out of all steel specifications in the world, almost none have a limit on aluminium.

(Figs. 45,46)



Page 32

CASE 2: STEERING BUSH:

BS970 1991 070M55 3OMM DIA. HARDEN/TEMPER TO 48-52 RC.

PROBLEM: Customer complained of cracked bushes after heat treatment. Close examination revealed that no external machining had taken place, so the original drawn bar surface (including the "as drawn" surface irregularities) acted as a nucleus fault which the cracking propagated.

NOTE: BS 9701991 shows that the allowable defects could be as deep as 1 % of the bar diameter. Hence removal is vital if such hardening treatments are to take place.

SOLUTlON: Purchase a larger diameter bar and remove surface imperfections by turning.

Crack detect components after heat treatment and prior to assembly. Understand steel specifications before buying!!



Page 33

CASE 3: SECURITY SAFE COMPONENT

PROBLEM: Customer complained of unusual bubble-like defects after electro-plating. This had never been experienced before (even though material choice was BS 970: 230M07).

REASON: Excessive sulphur segregation had reacted with the acid solutions during plating. This in turn had released hydrogen sulphide (H2S) gas which was trapped under the chromium layer, exhibiting large blisters.

This illustrates why sulphur bearing steels should not be electro-plated. Whilst some days may be successful a heavily segregated steel (or indeed newly replaced plating solutions) will cause gas generation.

SOLUTION: Switch to BS9701983 080A15 (En3) which has low (0.04% max.) sulphur. Such choice is unpopular with machinists, due to inferior machinability.

NOTE: Recent developments in inclusion modified mild steel grade showed considerable improvement in machining and has been adopted (albeit with slight increase in material cost compared to standard 080A15) with confidence.

NOTE: The increased cost meant that the steel was still cheaper than 230M07!



Page 34

CASE 4: FINAL TEST RIG FAILURE

BS 970 1983 817M40 ANNEALED (EN24)

PROBLEM: Customer experienced heavy failure rate on final test rig, when torque load of 10% higher than normal load was applied. After manufacture, the components were hardened and tempered to 42-53RC i.e. very high tensile coupled with an extraordinary tensile strength range.

REASON: Excessive allowable hardness range, plus tight corner radii in machined head meant high notch sensitivity. Hardness checks revealed that the components were excessively high in the range.

SOLUTlON: Modification of the corner radii was effected along with a change in heat treatment practice, such that 42-44 RC was achieved.

RESULT: No failures have occurred with the modified practice for at least one year.



Page 35

CASE 5: PINION GEARS - HEAVY MACHINERY

PROBLEM: Pinion gears failed in service when manufactured in BS970 1983 722M24/En4OB) 'T' condition and nitrided. Urgency of failure demanded two attempts with different steels, 655M13/En36 and 835M15 (En39). First pair failed in service after 3 hours in operation!

REASON: Detailed examination was carried out revealing that heavy Hertzian stresses had caused heavy fatigue load failure, which caused teeth to be thrown into other teeth causing heavy shear failure. Cause was insufficient (0.6mm) case depth. Also teeth were excessively hard (840 VPN) which would not allow bedding.

Most significant factor, however, was that only 60% of gear faces were damaged, clearly showing that severe misalignment had meant 40% of faces were not in contact. This alone would increase tooth loading to impossible surface loads and would contribute most to the failure.

SOLUTlON: Case depth specified to 1.0-1.2mm, case hardness tempered back to 700-720 VPN (aim). This would, therefore, support twice the Hertzian stresses on the case and the lower tooth hardness would be more suitable for "bedding in".



Page 36

CASE 6: CONTROL RACK SHAFT

PROBLEM: Material BS970 1991 605M36T (En16T) Customer complained "material too hard".

REASON: The material was checked - was within specified range. There did not appear to be any experience of hobbing racks, nor of medium tensile alloy steels. Their choice of material for the hob was high speed steel - BS4692 M42 Grade. High vibration rendered redressing/coating of the hobbing tools.

SOLUTION: Change to sintered tool steel (ASP 30) and excellent results were achieved.



Page 37

CASE 7: FILTER PLATE

PROBLEM: A visit to a customer resulted in a courtesy tour round his works. A pile of broken plates was observed where severe breaking had occurred.

Questions revealed that BO1 was used for filter plates, where many holes were drilled to allow waste to pass through. Breaking up had occurred where holes had not been drilled in a regular manner - fractures occurred between holes separated by the least amount of material. High hardnesses were needed - hence the need for BO1 (aiming 55RC and higher).

SOLUTION: It was suggested that hole spacings needed to be more regular. In addition to this, lower carbon alloy or tool steels should be adequate to achieve the hardness requirement.

An additional consideration would be to use a chrome bearing steel, harden and temper at 520 0C and nitride. Only trials would show if the superior hard surface (60RC) would be a better resistance to abrasion coupled with some resistance to corrosion.



Page 38

CASE 8: SPLINED SHAFT - GLASS MAKING MACHINERY

PROBLEM: Material BS9701991605M36T. Customer complained of numerous breakages in shafts and requested an explanation.

REASON: The failure was typical torsional fatigue failure which originated at the base of the splines which exhibited very poor surface finish.

SOLUTION: Recommendations were made to improve surface finish by careful grinding and consider surface hardening of the splines.



Page 39

CASE 9: SLIDE WAY - ENGINEERING INSTRUMENTATION

PROBLEM: Material BS9701991080A15 (En3) Bright Flat. Customer requested a visit due to "surface imperfections" which developed during machining.

REASON: It was evident that lof the drawn surface revealed random patches of non metallic inclusions. These are caused by elongation of (usually) silicate type steel making derived inclusions during the rolling process.

The actual patchy nature is caused within the casting process and their differing density causes them to "Float" or rise towards the upper bloom face; known as "north wall effect". They are unpleasing to the eye, uncontrollable in commercial grades of steel, but not detrimental to the performance.

NOTE: They do not always appear - usually in certain casting conditions

SOLUTION: The easiest solution is to turn the steel through 1800 (i.e. upside down) before cutting the guide groove and in almost every case they are usually not present in the other side.



Page 40

CASE 10: RACK SUSPENSION STUD

PROBLEM: Material BS970 1991 605M36T - 1/2" Bright. Customer asked if it could be "tack welded" - he was strongly advised not to; he insisted so written recommendation was faxed imploring them to "follow to the letter".

Six months later, catastrophic failures occurring everywhere. Steel had cracked - customer complained steel was defective. Sample of stud examined and brittle fracture was seen.

REASON:

Written instructions totally ignored. I. The "tack" weld had heated stud to red heat. II. Welding stopped - red hot stud "quenched" by surrounding (cold) metal. III. Stud hardened from 261 VPN to 43SVPN in extreme area. IV. Further welding done immediately adjacent the hardened stud with disaster

following. V. Crack was not evident until component in service - further movement caused

total failure.

(Fig. 47)



Page 41

CASE 11: TRACK PIN - HEAVY OFF ROAD VEHICLE (CIVIL)

PROBLEM: Material BS9701991 817M40 Annealed (En24) After many years of supply, total rejection by OEM - pins were bent. Customer desperate as more than 1/2 of the business were the pins.

REASON: Process route (unchanged for years)

1) Machine pin 2) Send to Heat treatment for harden temper to 'X' condition. 3) Deliver to contract grinders. 4) Back to manufacturer before despatch to OEM.

Something had changed! - not the steel:

a) Heat Treatment - Absolutely nothing since beginning. b) Contract Grinders - only thing to change was that a grinding machine operator

had retired after 50 years service. c) Back to Heat Treatment - were pins hardened whilst they were horizontal or were

they suspended vertical - horizontall!

Pins were checked before and after h/t - bent after heat treatment.

SOLUTION: The old (experienced) grinder has ALWAYS received them bent, but was skilful enough to set his machine to grind them parallel! New man obviously didn't know this.

From thence, pins treated by vertically suspending them and bending was eliminated.

The real reason was that when horizontal, the lower half of the pin was quenched in cool oil in relation to upper half, hence transformation (expansion) was uneven. When the pins were vertical, transformation is even.

CONCLUSION: Problem completely eliminated.

BGA Training Notes

TN6-1: Gear Geometry optimisation for low noise and stress using 3D F.E. Mesh Modelling


Page 1

BGA Training Notes TN6-1 GEAR NOISE AND VIBRATION GEAR GEOMETRY OPTIMISATION FOR LOW NOISE AND STRESS USING 3D F.E. MESH MODELLING

CURRENT DESIGN PRACTICE FOR INVOLUTE AND LEAD CORRECTION FOR SPUR AND HELICAL GEARS

Author: D.A. Hofmann, Design Unit, Gear Technology Centre, University of Newcastle upon Tyne, Stephenson Building, Claremont Road, Newcastle upon Tyne, NE1 7RU [email protected] http://www.newcastle.ac.uk/~nmecheng/

BGA Training Notes



Page 2

Index

CURRENT DESIGN PRACTICE FOR INVOLUTE AND LEAD CORRECTION FOR SPUR AND HELICAL GEARS. ...................................................................................................................................... 3

1. LOAD AND TRANSMISSION ERROR IN SPUR GEARS......................................................... 3

2. INVOLUTE CORRECTION IN SPUR GEARS ........................................................................... 5

3. LEAD CORRECTION, CROWNING AND OTHER LEAD MODIFICATIONS..................... 9 3.1 LEAD CORRECTION...................................................................................................................................9 3.2 CROWNING .............................................................................................................................................10 3.3 THE EFFECT OF LEAD CORRECTIONS ON TE IN SPUR GEARS..................................................................11

4. CONTACT IN HELICAL GEARS ............................................................................................... 12

5. TRANSMISSION ERROR IN HELICAL GEARS..................................................................... 13

6. INVOLUTE AND LEAD CORRECTION IN HELICAL GEARS................................................ 15

REFERENCES............................................................................................................................................ 18



BGA Training Notes



Page 3

CURRENT DESIGN PRACTICE FOR INVOLUTE AND LEAD CORRECTION FOR SPUR AND HELICAL GEARS. This is a brief summary of involute and lead correction in gears, to provide some background to the use of more sophisticated analysis techniques to optimise gear geometry for low noise and vibration. Both spur and helical gears are considered, and the difference in their behaviour is examined.

It should be noted that involute and lead correction are considered in the traditional manner, that is separately, in the case of spur gears. They must however be considered in combination for helical gears.

1. Load and Transmission Error in Spur Gears In normal spur gear drives, with a transverse contact ratio less than εα = 2.0, load is transmitted by two teeth at the start and end of contact, but by only one tooth over part of the mesh cycle around the pitch point. This is clearly shown in Fig. 1, with 2 teeth in contact at the start of active profile on the pinion, and only one in contact at the pitch point.

Even in a perfectly accurate gear, a tooth pair experiences a step change in load four times, at the start and end of contact, and at the change over from double to single tooth contact. Since the teeth are not perfectly rigid, these changes in load result in different deflections of the gear teeth, with a resultant change of wind-up or Transmission Error (TE) between pinion · and wheel. This effect is shown diagramatically in Fig.2, where the load on a tooth (Fig. 2a) and the TE (Fig. 2b) are plotted against length of roll.

(The total length of the line of contact from start to end of engagement is distance A - E in Fig.1).

Fig. 1.a - Start of Contact 2 teeth load

carrying

Fig. 1.b - Contact at Pitch Line 1 tooth load

carrying

Fig. l.c - End of Contact 2 teeth load carrying

BGA Training Notes



Page 4

We note therefore that spur gears of perfect accuracy, have ZERO transmission error at no load When transmitting torque, a variable component of TE equal to approximately half the mesh deflection occurs.

At a speed when gear element inertia becomes significant, this variable component of transmission error causes acceleration and deceleration forces, which result in gear noise and vibration. Profile correction was introduced to reduce loaded transmission error, and gear noise and vibration.

Fig. 2 - Load and TE against Length of Line of Contact

BGA Training Notes



Page 5

2. Involute Correction in Spur Gears Involute correction has two objectives

i. to prevent tip contact, that is non-involute contact ahead of point A on the path of contact. (Fig.1)

ii. to minimise transmission error, and hence gear noise and vibration.

The first objective, to prevent tip contact, is easily attained. The mesh deflection at a given load plus pitch and profile errors can be calculated, and appropriate tip and root relief applied to pinion and/or wheel, so that the combined relief at start and end of contact is equivalent to the total mesh deflection plus tolerances. This will avoid tip contact and non-involute action.

The second objective to minimise TE is more difficult to attain, since it requires the extent of involute correction to be defined. Two limiting conditions can easily be conceived:

i. To modify the involute profile such that at least a base pitch of true involute profile is retained, as shown in Fig.4a. In this case, referred to as 'short' relief there is zero transmission error at no load, but full transmission error still occurs at the design torque (for which the amount of relief is appropriate. (See Fig. 4)

Fig. 3 - Tip Relief a) shown on gear tooth b) shown on involute diagram

Fig. 4 - Short Relief

BGA Training Notes



Page 6

ii. To modify the involute profile so that relief extends from the start of active profile (SAP) to the point of theoretical transfer from double to single tooth contact, i.e. relief extends from A to B and from D to E on the roll diagram (Fig.5a). In this case we have a short involute with εα < 1.00, and transmission error occurs at no load, and is reduced to almost zero at the design torque, (where the extent of relief = elastic mesh deflection).

Munro [1] has summarised the approximate transmission error characteristics for spur gears for the full range of possible options, from zero relief to very long relief. This extremely useful compilation, shown in Fig. 6, shows clearly that, for normal spur gears with contact ratio εα < 2, it is impossible to achieve very low transmission error over a wide torque range. Examining Fig. 6, it is clear that intermediate relief, extending to about 0.75 pe (εα -1) gives the best compromise for the extent of tip and root relief over the torque range from 0 to 100% of FLT, (assuming tip relief appropriate to FLT).

Fig. 5 - Long Relief

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Page 7

The simple analysis on which Fig. 6 is based is borne out by more accurate FE elastic mesh analysis. Figure 7a and 7b show the transmission error v. phase of mesh for a pair of 8mm module spur gears,

z1 = z2 = 20, b = 25, a =160, x = 0,

for a) short relief, and b) intermediate relief.

This confirms the trends shown in Munro's summary, Fig. 6.

Note:

The value of mesh stiffness to be used in the calculation of the amount of tip and root relief is not well documented. From the FE analysis, C' = 10 .. 12 N/mm/µm which is significantly less than predicted by BS 436 or DIN 3990.

Fig. 6 - Relief of Transmission Error Summarised By Munro

BGA Training Notes



Page 8

Fig. 7a - TE for Short Relief (Computed by DU-GATES) 160 Centres Test Riq 8 Mod. 20 T / 70mu SHORT Tip & Root Relief

Transmission Error vs. Phase vs. Torque

Fig. 7b - TE for Intermediate Relief (Computed by DU-GATES) 160 Centres Test Riq 8 Mod. 20 T / 70mu INTERMEDIATE Tip & Root Relief

Transmission Error vs. Phase vs. Torque

BGA Training Notes



Page 9

3. Lead Correction, Crowning and other Lead Modifications Lead correction is applied to all gears to correct for elastic deflection, and other systematic errors, as detailed in 3.1 below. Lead crowning is used to make the gears less susceptible to random alignment errors, resulting from manufacturing tolerances, assembly and mounting tolerances. Depending on the type of gearbox, the random errors can be lesser or greater than the systematic errors.

Lead correction and crowning have a significant effect on TE and noise and vibration in helical gears, but it also has a secondary effect on noise and vibration in spur gears.

3.1 Lead Correction

It is usual to provide lead correction on only the pinion. The total correction is made to compensate for:

i. the bending deflection of the pinion.

ii. the bending deflection of the wheel (where significant).

iii. the torsional wind-up of the pinion

These deflections can be summed to give the total lead error due to elastic deflection as shown in Fig.8. The lead correction should compensate for the total elastic deflection, e.g. the form of the lead is the mirror image of the deflection curve.

(Note: In many cases it will also be necessary to allow for the flank misalignment caused by bearing clearance, and bearing and housing compliance).

Fig. 8 - Bending and Torsional Deflection

BGA Training Notes



Page 10

For symmetrically straddle-mounted gears only, the effective bending deflection, bf∆ , is:

4

12712 C

bEbFf tD

b

−=∆

π

The effective torsional wind-up tf∆ is:

24 CGbFf tD

t π=∆

3.2 Crowning

Crowning is used to allow for manufacture and mounting deviations. With a total probable deviation fma between pinion and wheel flanks, crowning to a height of 1/2 fma should be applied to either the pinion or wheel. This crowning should be superimposed on the lead correction for full compensation of elastic deflection and lead deviations.

With medium facewidth gears asymmetric crowning, where hc2 - hc1 = ∆fb + ∆ft can give an adequate approximation for the sum of ideal lead correction and crowning.

To a first approximation, at design torque, ideal crowning will reduce a load distribution factor KHβ for the uncrowned gear with alignment error fma to:

( )1211 −+= ββ HCrH KK

Fig. 9 - No-Load gap and Crowning

BGA Training Notes



Page 11

3.3 The Effect of Lead Corrections on TE in Spur Gears

The interaction of misalignment and lead correction in spur gears affects the load intensity and thus the elastic mesh deflection and hence the noise and vibration characteristics This can best be demonstrated qualitatively by considering a perfectly aligned and a misaligned gear pair. Assume a load distribution as shown in Fig.10b, with say a load distribution factor KFβ = 2, i.e. zero load one end, 2 x mean load at the other end of the tooth.

If we assumed that long tip and root relief has been applied to these gears to match the mean load occurring under ideal conditions, as shown at 'a'. In that case, the TE variation for perfect alignment 'a' would be zero. For the misaligned gear, case 'b', thinking of the gear tooth as a series of thin slices of spur gear, the TE variation will vary from zero at the centre to extreme value of 50% of the mean mesh deflection at the ends.

Summing these transmission errors across the facewidth will then give a mean TE characteristic for the misaligned gear as shown, with a TE range of 25%, and severe rates of change of TE.

The simplified 2D analysis is obviously inadequate for dealing with the real case of misaligned spur gears with lead corrections. Only a full 30 FE mesh model will be able to predict TE in such a case with any accuracy, especially when the gears concerned have more complex lead correction - for example asymmetric crowning and. end relief.

Fig.10 TE in Misaligned Spur Gears

a) Load Distribution in perfectly aligned gears

b) Load Distribution in misaligned gears

c) Transmission Error at 200%. 100% and 0% of FLT at sections I, II and III, and resultant T.E. Long Tip and Root Relief

BGA Training Notes



Page 12

4. Contact in Helical Gears A helical gear can conveniently be thought of as a number or slices of spur gear, where the slices L1 .. L5 are rotated relative to each other (Fig. 11).

Fig. 11 - Helical gear considered as staggered slices of spur gear (from [2])

Contact still takes place in the plane of action, A - E (Fig.11c), but the phase of mesh varies from slice L5 to L1. Fig.12 shows the position of contact lines on a helical gear at a particular phase of mesh.

Fig. 12 Contact lines on a helical gear

Assuming perfect gears, perfectly aligned, on pinion tooth I contact will initially have been made at 'A' in the root, the entry point, and the contact lost at the tip, point 'B', the

BGA Training Notes



Page 13

exit point. In a perfect gear at zero load, initial and final contact will be on point like contact lines. Under load, elastic deflection of the preceding teeth will however cause some tip-interference at entry and exit of mesh.

5. Transmission Error in Helical Gears In spur gears of normal geometry, the primary cause of TE in the change from two teeth in contact to one tooth in contact, with the resultant change in mesh stiffness. The change of mesh stiffness can be thought of as being proportional to the length of the line of contact, which changes from

Lmin = b - in single tooth contact, to

Lmax = 2b - in double tooth contact.

In helical gears, the average length of the line of contact, irrespective or helix angle, is

bLav ∗= αε

The variation of the length of the line of contact depends on the helix angle, facewidth and module, that is the face-contact ratio εβ, where

( )πβε β

nmbsin=

Fig.l3 (from Niemann-Winter [2)) shows how the total length or the line of contact varies with phase of mesh for spur gears, and helical gears with εβ = 0.5, εβ = 1.0, εβ =1.5.

The variation in length of line of contact can be equated to the variation in mesh stiffness, and thus directly to TE.

Fig. 13 - Lines of contact in the base tangent plane at 6 phases of mesh, and the corresponding total length of the lines of contact at that position.

BGA Training Notes



Page 14

It is noted that:

The variation of the length of the line of contact for the spur gear, b < L < 2b is identical to the TE curves shown in Fig. 2.

• the length of the line of contact for the helical gears with εβ = 0.5 is 1.2b < L < 2.0, . i.e. the TE is not significantly better in amplitude than for a spur gear but the change of TE is quite gradual.

• At εβ = 1.0, the length of the line of contact remains constant, and thus, to a first approximation, one could expect zero transmission error under all load conditions.

• At εβ =1.5, the length of the line of contact varies gradually from 1.47pe, < L < 1.73pe,. The variation of TE in this case would be just 25% that of the same spur gear at the same load.

The solution to TE, and hence noise and vibration, in helical gears would therefore appear to be very simple - design gears with integer face contact ratios, e.g. εεεεββββ = 1.0, 2.0, 3.0. Such gears do not require tip and root relief for constant mesh stiffness and minimum transmission error. Unfortunately, in real gears, the design of optimal gear geometry is not so easy.

• Mesh stiffness of an uncorrected involute profile varies with phase of mesh by ±15% to ±25%, leading to not insignificant transmission error if the profile is not corrected.

• It is not possible to manufacture perfect gears, far less bring them into mesh without some misalignment With misaligned gears, whether lead corrected or not, transmission error greater than shown in Fig. 13 will occur. The analysis of transmission error in helical gears with lead correction and involute corrections requires sophisticated computation and is described elsewhere [3]. Some general factors to consider when specifying involute and lead correction in helical gears are however considered in Section 6 below.

BGA Training Notes



Page 15

6. Involute and Lead Correction in Helical Gears Traditionally, helical gears are ground with the same tip and root relief as spur gears. Some authorities specifically recommend long relief for helical gears, since kinematically correct transmission is maintained even when εα < 1.0, provided εγ = (εα+ εβ) > 1.0.

As seen in Fig. 13 above, this justification for tip and root relief in helical gears is unfounded. Tip and root relief will generally decrease the effective εα, and thus decrease the mean length of line of contact, which will result in higher specific load and larger total mesh deflection and TE. Thus classic tip and root relief increases the contact stress, as shown in Fig. 14, and reduces the load capacity of a gear pair.

Fig. 14 - a) Conventional tip and root relief b) MAAG Generated End Relief showing load distribution with and without tip relief.

Conventional tip relief is quite unnecessary in helical gears. Even generated end relief as proposed by MAAG is difficult to justify in terms of reduced TE, but will be useful in reducing peak contact stress at entry and exit.

In helical gears, there is no "entry" and "exit" shock such as occurs in spur gears, since initial contact is quasi point contact at the start of engagement. A local high load intensity ("shock-load") on a very short contact length will not generate significant TE.

However, some tip and root relief should be provided to avoid the high stress intensity at the two ends of the line of contact, where tip and root area of mating gears contact.

BGA Training Notes



Page 16

Fig. 15 - Contact Stress on Finite Length Contact Lines a) Real contact stress in Helical gear without tip relief b) Contact stress on sharp cornered roller c) Contact stress on radiused roller

Sharp edges on a gear, (Fig. 15a) like sharp edges on a rolling element bearing (Fig. 15b), lead to high contact stress at the edge, which can be avoided by careful radiusing of the comer. The extent of tip relief can be quite small. Experimental helical gears have been run very successfully, with a nominal load intensity of 700 N/mm facewidth where the amount of tip relief on pinion and wheel is equal to the mesh deflection (35 µm) but the extent was only 0.4 module, compared to the more usual 0.9 to 1.8 mod (in the roll or involute diagram).

The design of lead correction in helical gears should be the same as in spur gears, with helix correction for 'systematic' errors such as elastic torsional wind-up and shaft deflection, and crowning to compensate for random misalignment. In a helical gear, crowning and lead correction introduce no-load gaps which affect the load intensity in the same way as tip and root relief. Fig. 16 shows the no-load gap for a lead crowned helical gear tooth, and the resulting load distribution along 3 contact lines, drawn for the sake of simplicity on the one tooth.

BGA Training Notes



Page 17

Fig. 16 Lead Crowning and Load Intensity along Contact Lines in a Helical Gear

It is clear that the analysis of TE for such a relatively complicated load distribution, with complex variable mesh stiffness, requires the use of full 3D FE mesh modelling.

To summarise therefore:

• Tip and root relief of the type used on spur gears is inappropriate for helical gears.

• Tip and root relief affect primarily the transverse contact ratio in helical gears, and as such do not have as significant an effect on TE as helix angle and lead correction.

• Some tip relief is required to avoid high stress at the end of the line of contact.

• Lead correction for systematic (elastic) and random (tolerance) effects is an absolute necessity on helical gears, both:

a) to maximise load capacity, and b) to minimise the effect of misalignment on the effective face contact

ratio and hence on the TE and noise and vibration.

• The amount of lead correction should always be a minimum consistent with avoiding high loads at the ends of gear teeth.

• Helix angle and tip relief can be optimised to achieve minimum TE for a given lead. correction using a full elastic mesh analysis.

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REFERENCES 1. Munro R.G., Yildirim N. "Some measurements of static and dynamic

transmission errors in spur gears". Proc.1994 International Gearing Conf. Newcastle, pp.371 .. 3io

2. Niemann G., Winter H., "Machinenelement Vol. II", Springer 1985, pp.61 .. 64

3. DU-GATES - Design Unit Gear Analysis for Transmission Error, University of Newcastle upon Tyne,1995

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TN6-2: A new tool for designing quiet, low vibration main propulsion gears


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BGA Training Notes TN6-2 GEAR NOISE AND VIBRATION

A new tool for designing quiet, low vibration main propulsion gears

Authors:

P. Maillardet BSc, MSc, DMS, RCNC Ministry of Defence (Navy)Naval support Command, UK. D.A. Hofmann BSc, CEng, MIMechE, MIEE M E Norman BSc, MSc, BA Design Unit, Gear Technology Centre, University of Newcastle upon Tyne, Stephenson Building, Claremont Road, Newcastle upon Tyne, NE1 7RU [email protected] http://www.newcastle.ac.uk/~nmecheng/

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Index

A NEW TOOL FOR DESIGNING QUIET, LOW VIBRATION MAIN PROPULSION GEARS ...... 3

BACKGROUND ........................................................................................................................................... 3

INTRODUCTION ........................................................................................................................................ 4

SPECIFICATION FOR A GEAR DESIGN OPTIMISATION TOOL ................................................... 5 TRANSMISSION ERROR AND BEARING LOAD VARIATION ..................................................................5 ELASTIC MESH ANALYSIS ............................................................................................................................7 FE MODEL AND SOLUTION OF COMPATIBILITY AND EQUILIBRIUM EQUATIONS.........................8 THE SOFTWARE PACKAGE .........................................................................................................................10 EXAMPLE GEAR PAIRS ................................................................................................................................12 MINIMISING TRANSMISSION ERROR........................................................................................................13 OPTIMISING LOAD DISTRIBUTION FACTOR...........................................................................................17 MINIMISING QUASI-STATIC BEARING LOAD .........................................................................................21 DESIGN OF A WELL-BALANCED GEAR PAIR ..........................................................................................23 EXPERIMENTAL VERIFICATION OF ELASTIC MESH ANALYSIS ........................................................25

Bearing Load ........................................................................................................................................ 26 Transmission Error............................................................................................................................... 26

CONCLUSIONS ...............................................................................................................................................27 ACKNOWLEDGEMENTS...............................................................................................................................27 REFERENCES ..................................................................................................................................................27

NOMENCLATURE ............................................................................................................................... 28



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A new tool for designing quiet, low vibration main propulsion gears Low noise and vibration is a primary requirement in naval gearing. A calculation procedure, Design Unit-Gear Analysis for Transmission Error and Stress (DU-GATES), is described, which facilitates the optimisation of gear geometry for low noise at the design stage. The requirements for a design tool are outlined, and the concept of transmission error and quasistatic bearing load variation introduced. The elastic mesh analysis and the solution of the equations equilibrium and compatibility and the calculation procedure are described. The capabilities of the calculation procedure are demonstrated by example for a low and a high helix angle gear, and the improvements in static TE and quasi-static bearing load variation, which can be achieved, are detailed. The design of well-balanced gear is discussed, and the proving of the procedure against experimental data outlined. It is concluded that significant reductions in gear noise and vibration can be achieved by appropriate optimisation of gear geometry.

BACKGROUND Naval main propulsion gearing must be both reliable and very quiet. To this end, the Royal Navy has supported gear research since the Second World War, through the Admiralty and Vickers Gear Research Association (AVGRA) and the Navy and Vickers Gear Research Association (NAVGRA). This research investigated gear manufacture, metrology and gear fatigue strength, and established significant relationships between gear noise and gear accuracy and transmission error1,2,3,4. Even after the cessation of NAVGRA research in 1979, the Royal Navy has continued to support gear research both at major gear manufacturers such as David Brown, GEC Alsthom Gears, VSEL and Allen Gears and at UK universities.

Since 1989 the Royal Navy has funded a programme research at the Design Unit, University of Newcastle. A major gear research rig has been commissioned5, which has been used to establish empirical design rules for marine gears, and for fundamental research into gear dynamics6 The FE based elastic mesh analysis tool for gear design optimisation described in this paper was developed in conjunction with this programme.

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INTRODUCTION There has been considerable research into the relationship between the kinematic behaviour of gears and gear noise and vibration. In the late 1950s and early 1960s Kohler7 and Munro8 confirmed a dear relationship between quasi-static transmission error (TE) and gear noise and vibration in spur gears. Recent work on the Marine Gearing Research Rig9 has substantiated this relationship between static transmission error under load and gear vibration and noise for large, wide facewidth, high contact ratio helical gears. Since TE represents the principal excitation, minimising TE will minimise gear noise and vibration irrespective of shaft, bearing and gearcase dynamic response.

Conventional gear design techniques for naval gearing, based on Naval Engineering Standards10 or ISO standards11 or Lloyds Rules12, size gears to transmit a specified torque without surface fatigue (pitting) or tooth bending fatigue failure. These traditional design techniques do not consider gear noise and vibration.

Modern powerful gear analysis techniques are based on 3D FE modelling of gears. In addition to very accurate gear stressing, these can carry out a full elastic mesh analysis which calculates loaded TE, and can thus be used to optimise gear geometry for minimum noise and vibration. The elastic mesh analysis described here, DU-GAT'ES, has been especially developed to analyse wide facewidth gears typical of naval gearboxes, with total contact ratios up to 9 (i.e. high helix gears with up to nine teeth in contact). An important feature of DU-GATES is that it has been extensively validated against measured data from real, full size gears on the Marine Gearing Research Rig5 This paper considers primarily the use of DU-GATFS for design optimisation for low noise and vibration, but will include some examples of design optimisation for minimum stress.

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SPECIFICATION FOR A GEAR DESIGN OPTIMISATION TOOL Any new calculation procedure for optimising the detailed geometry of gears must be based on achieving minimum transmission error at acceptable stress levels taking full account of:

1. the range of torques over which quiet running must be achieved;

2. the mesh misalignment due to elastic deflection of gear bodies, shafts, bearings and gearcase;

3. manufacturing tolerances for gears, shafts, and gearcase.

To be useful to the gear designer, the procedure must be:

1. fast and easy to use and applicable to all possible forms of involute and lead correction and modification;

2. capable of very high discrimination in the calculation of TE (TE < 0.1 µm);

3. able to analyse load distribution and TE rapidly for many mesh positions (typically 32 steps are required to represent accurately the TE characteristics over a single base pitch of gear movement).

The procedure described meets the requirements of this specification.

TRANSMISSION ERROR AND BEARING LOAD VARIATION When two perfect (i.e. error free) involute gears are meshed together under no-load conditions, rotation of the driving gear results in uniform rotation of the driven gear in proportion to the number of gear teeth in the two gears. When two real gears, that is gears with pitch, profile and lead deviation and mounting errors (within specified tolerances) are meshed together at zero load, with particular shaft misalignment, when the rotation is traced against time (or phase of mesh), it is observed that there is a deviation from the uniform motion defined by the simple gear ratio due to these geometric deviations, which is termed the 'unloaded or geometric transmission error'.

When torque is applied to a gear pair, there is torsional wind-up of one gear relative to the other in proportion to the load and the elastic stiffness of the gear teeth, the 'meshstiffness'. Variations in mesh stiffness at different phases of mesh result in elastic TE. This is best illustrated by considering the case of a spur gear pair. As shown in Fig 1(a), at the start of engagement of pinion tooth II, tooth I and II transmit the torque.

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When tooth I leaves mesh, the torque is transmitted only by tooth II, until tooth III comes into mesh when the torque is again transmitted by two teeth. Tooth loading therefore changes two-fold from initial engagement at the start of involute to contact at about the pitch line and then to contact at the tip at end of engagement, with an equivalent change in elastic mesh deflection or transmission error. As can be seen schematically in Fig 1(b), the TE, as uncorrected spur gears pass through mesh, is approximately half the total elastic deflection of the gears in single tooth contact.

Fig. 1 - Load sharing during the engagement of spur gears

(a) Tooth contact sequence for tooth passing through mesh from entering base tangent (or

contact) plane at A to leaving at E

(b) Tooth load and transmission error (TE) as a function of phase of mesh

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The combination of the geometric transmission error and the elastic mesh deflection is the total or kinematic transmission error. Under dynamic conditions, i.e. when gear inertia is considered, TE generates excitation forces in the gear mesh, which excite a dynamic response of shafts and bearings and transmit dynamic loads into the gearcase, the response of which then results in gear noise being radiated from the gearcase and vibration being transmitted to the gearbox mounting (Fig 2).

In wide faced gears with relatively small bearing span, a further source of dynamic excitation is the fluctuation of bearing loads with phase of mesh due to variation in axial position of the resultant of the distributed load along the contact lines between mating gear teeth. This results in quasi-static variation of bearing load at tooth contact frequency, which also excites gearbody and shaft and thus gearcase vibration.

ELASTIC MESH ANALYSIS

The relatively complex problem of determining the. load distribution between mating gear teeth and the elastic deflection of these can be solved by setting up equations of compatibility of displacement and equilibrium of forces for a sufficient number of discrete points representing the contact region between meshing gear teeth. In addition to the known no-load geometric deviation at each point, the formulation of suitable equations requires the direct and cross compliance between discrete points to be calculated and any load-dependent geometric changes due to other mounting compliances. These equations are solved for the discrete mesh loads, which can be re-expressed as specific load distribution; and for kinematic transmission error. The

Fig. 2 - Sources of gear noise and vibration

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veracity of the analysis will depend on the accuracy with which the mating gear surfaces are described, and the number of discrete contact points. In this analysis up to 120 contact Points are used.

The calculation procedure DU-GATES is a development of the approach used by Steward13,14 for spur gears, who identified that the overall gear body deflection contributes significantly to mesh compliance. Haddad15 extended Steward's work to helical gears, with a full 3D FE model which included the gear body, main tooth and two first-adjacent teeth. The compliances due to further adjacent effects were estimated. The present implementation benefits from the speed and capacity of more recent computers, and the FE model now includes not only the gear body and main tooth but also up to 17 adjacent teeth, resulting in an FE model with typically 15 000 deg of freedom. The curve fitting procedures, outlined under the following heading, have also been significantly enhanced.

FE MODEL AND SOLUTION OF COMPATIBILITY AND EQUILIBRIUM EQUATIONS

For a given pair of gears the first stage of the analysis is to generate and store, for each gear, a set of compliance influence coefficients such that combined compliance can be reconstructed at a general point on the contact line with respect to concentrated load at any general contact point. A finite element model is devised such that local contact deformation is excluded to give the compliance at the tooth flank due only to bending and shear deformation of the gear body and teeth. This is achieved by extracting the displacements at corresponding points projected normal to the flank lying on the tooth centreline enabling the non-linear (Hertzian) contact deflections to be introduced later in the tooth contact analysis program where the non-linearity with load can be taken into account It has been found adequate to apply unit point loadings to the finite element model at five radial positions and eight axial positions, closer spaced towards the ends of the teeth where displacement gradients are expected to be higher. Displacements are output at the same five radial positions for the main tooth and at two radial positions per adjacent tooth, up to 17 teeth. Across the facewidth, 27 nodal positions are used for displacement output, providing a refinement of the spacing in the region of the load positions, where displacement gradients are high. In the radial direction, linear interpolation between the discrete load positions and the displacement position can be shown to be suitable. In the axial direction, the use of curve-fitting functions is inevitable. Exponential functions have been developed with 12 terms, modelling not only the underlying symmetry of displacements about the loaded position but also the end effects as buttressing reduces toward the ends of the tooth and asymmetry of end effects resulting from the 'sharp' end and 'blunt' end of the helical gear. Also, both constant and asymmetric effects of gear body movement appear in the functions, which thus allow for out-of-plane tilting and rotation of the gear.

If root stressing is required, stresses may be output from the FE analysis at the 30 deg tangent points and at the root apex. Similar curve fitting gives a set of stress influence coefficients such that the root bending stresses can be reconstructed at a general axial position due to concentrated load at any general contact point.

The stored influence coefficients will be applicable for contact analysis at any phase of engagement and for any combination of operating conditions, mounting conditions, misalignments and tooth modifications or errors.

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At each phase step the (multiple) contact line will be in a new position and a new tooth contact analysis will be required. The program repeats the analysis at sufficiently close phase steps to plot continuous graphs of transmission error, load distribution factor, quasi-static bearing loads and root bending stresses versus time.

In each contact analysis the equations are set up as follows: In unloaded contact, only one point (k' th) is in contact, such that

( )iikk ectect −=− min i = 1 to n (1)

(Note that ctk - ek may be +ve or -ve since conventions for elements of ei are arbitrary).

Then no-load gap is:

( )kkiii ectectg −−−= (2)

Under load, the unknown elastic deflection, is:

∑ == n

j jjiji IwCd1

(3)

Then, at each point, di = de - gi or, rearranging with unknown = known:

iie gdd =− (i = 1 to n) (4a)

Equation (4a) expresses the conditions for compatibility of displacement at the n discrete points and represents n equations in n + 1 unknowns.

The n + 1'th equation is provided by the equilibrium condition, that is, the total normal load, F, being known:

FIw jj

n

j=∑

=1 (4b)

Strictly, the set of equations (4) is non-linear with load because the compliances, Cij have been adjusted before this stage to include a load-dependent theoretical contact 'compliance'. Thus, an iterative solution process is necessary, the contact 'compliances' being revised as estimates of the unknown loads change. Evidently, there is a further nonlinearity since a solution containing negative values of loads would be invalid. Therefore, negative loads are set to zero at each iteration of the quasi-linear solution, which typically converges in about seven iterations.

In addition to the specific load distribution, the solution yields a single value of the elastic mesh deflection, de.

Defining positive transmission error such that the driven gear lags the driver, positive de gives a positive elastic component of transmission error. Similarly, positive ctk - ek gives positive unloaded geometric transmission error. However, the latter includes an arbitrary constant component (independent of the phase of engagement) since the

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geometric error term, ek is partly independent of the gear topography (due, for example, to mounting misalignment). Thus, when expressing transmission error versus phase, a mean value can be deducted to leave only the alternating component· The loaded transmission error is given by the sum of the unloaded transmission error and the elastic mesh deflection. Again, only the alternating component is meaningful.

THE SOFTWARE PACKAGE

DU-GATES can be used for the analysis of all parallel axis gears, i.e. spur, helical, double helical and single helical gears with thrust cones. The package is shown schematically in Fig 3, and can be divided into two parts:

1. finite element analysis and generation of compliance and stress coefficients: for a given macro-geometry this part of the analysis is run only once; and

2. tooth contact analysis which is run repeatedly to optimise the micro-geometry to achieve minimum TE and gear stress.

The software runs under UNIX with interactive data input and job control via a graphical user interface under X Windows. The macro-geometry for the mating gears needs to be entered (e.g. number of teeth, helix angle, module, tooth depth, addendum modification, protuberance, face-width etc). A dedicated pre-processor automatically generates the FE model of the gear, and an internal FE solver and a curve fitting procedure generate the compliance and stress coefficients described in the section called 'FE model and solution of compatibility and equilibrium equations'.

The analytical contact analysis is then carried out for a specified micro-geometry (e.g. tip relief, root relief, profile crowning, lead correction, end relief, face crowning or any defined surface topography) at a number of phases of mesh (typically 16 to 32 increments in a base pitch) to determine kinematic transmission error, load distribution and root bending stress at given torque(s) and misalignment(s). The micro-geometry can then be varied to achieve minimum TE and stress for the full range of operating conditions.

Output from the calculation procedure - kinematic TE, bearing load, tooth load, contact and bending stress -can be output either as a function of phase of mesh or, in the case of load and stress, as a distribution across the face-width at a particular phase of mesh.

The computation time, depending on the work-station used, is typically about 3h per gear for the FE analysis and curve fitting, and about 1 min for the elastic analysis per phase increment.

Under the forthcoming headings 'Minimising transmission error', 'Optimising load distribution factor' and 'Minimising quasi-static bearing load', the use of DU-GATES is demonstrated for a typical low and high helix angle gear, varying both macro-geometry (helix angle) and micro-geometry to explore the effect on TE, quasi-static bearing load variation and load distribution factor

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Fig. 3 - Block diagram of DU-GATES software package

.

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EXAMPLE GEAR PAIRS

As an example, DU-GATES has been used to analyse the following two typical gear pairs used on the Marine Gearing Research Rig at Newcastle. The gear details are shown in Table I.

To emphasise the importance of making realistic assessments of the actual operational mesh misalignment in wide facewidth single helical gears with thrust bearings, the majority of the analysis has been carried out for the condition of maximum expected operational misalignment, in this case 70µm over 200 mm facewidth, with some examples of elastic mesh analysis at lower and zero misalignment. Since naval gearing must achieve the lowest noise levels at part load, i.e. at patrolling speed, all of the analysis has been carried out at 50% of maximum engine revolutions, i.e. 25% of maximum rated torque. Gear stress and TE must of course also be calculated at maximum torque and other part loads in any full gear analysis.

The analysis of transmission error, load distribution factor and quasi static bearing load has been carried out for the following conditions:

1. torque: 4000 Nm (i.e. 25% of full load torque);

2. gear element misalignment: 70 µm hard forward, i.e. toward the driven end of the pinion,

investigating the effect of varying:

a) macro-geometry: helix angle variation by 2o about a nominal helix angle;

b) micro-geometry: lead crowning (from 0 to 120µm); tip and root relief (from 0 to 40µm).

The results of this analysis are discussed below.

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MINIMISING TRANSMISSION ERROR

Figure 4 shows the TE for the maximum expected alignment of 70 µm for the low helix angle gear plotted as a function of tip relief from 0 to 40 µm and crowning from 0 to 120 µm. Figure 4(a) shows TE for a helix angle of 6o, while Figs 4(b) and 4(c) show TE for 80 and 100 respectively. It is observed that:

1. the minimum achievable transmission error reduces significantly from 1.4 µm at β = 6o to 0.3 µm at β = 10o;

2. the 'shape' of the area over which minimum TE is achieved varies with helix angle (Figs 4(a) and 4(c));

3. at 6o helix angle the minimum transmission error occurs at close to zero tip relief, with little change between 10 and 80 µm crowning. At 10o helix angle, minimum transmission error is very Iocalised at zero tip relief and 40 µm crowning, but in this case a low value of TE is possible at about 40 µm crowning for all values of tip relief from 0 to 40 µm.

Table 1 - Gear details

High helix gear Pair Low helix gear pair

Number of teeth z 29 87 33 99

Module mn =6 =6

Pressure angle αn 20o 20o

Helix angle β 28 o /30 o /32 o 6 o /8 o /10 o

Facewidth b 200 200

Addendum correction x 0 0 0 0

Tip diameter dn =212 =612 =212 =612

Tooth depth mn 2.4 2.4 2.4 2.4

Transverse contact ratio εα 1.45/ 1.42 /1.37 1.75/ 1.74/ 1.72

Face contact ratio εβ 4.91/ 5.23/ 5.77 1.10/ 1.50/ 1.85

Figure 5 shows TE as a function of crowning and tip relief for the high helix angle gear at 28, 30 and 32 degrees respectively at 70 µm misalignment. In this case:

1. The minimum transmission error possible is 0.25 µm at β = 28o and 0.31µm at β= 32o compared to 1.4 µm at β = 6o and 0.3 µm at β = 10o. With TE as little as one tenth of that of the low helix gear, the dynamic mesh forces and the noise and vibration will be significantly lower.

2. For all three helix angles, TE is predominantly a function of crowning, with tip relief having only a secondary effect.

3. Minimum TE at β = 32o is very localised at zero tip relief and 65 µm crowning.

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Fig. 4 - Low helix gear pair misaligned 70 µµµµm hard forward variation of transmission error

with profile and lead modification and helix angle

Fig. 5 - High helix gear pair misaligned 70 µµµµm hard forward variation of transmission error with

profile and lead modification and helix angle

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The effect of misalignment on TE for the low helix gear (β= 10o) and the high helix gear (β= 32o) is dearly shown in Figs 6a and 6b respectively, which show TE as a function of crowning and tip relief at zero and at 70 µm misalignment.

It is noted that:

1. For the low helix gear (Fig 6a(1) and Fig 6a(2)), minimum TE is achieved with extremely small crowning when there is no misalignment, and TE is halved compared to the best design which can be achieved at 70 µm misalignment.

2. For the high helix gear (Fig 6b(1) and Fig 6b(2)) the effect of eliminating misalignment is to reduce TE significantly for all gear geometries with low crowning, and to shift optimum crowning from 65 µm to 25 µm, with a minimum TE of less than 0.03 µm.

Fig. 6a - Low helix gear pair: variation of transmission error with profile and lead

modification and misalignment

Fig. 6b - High helix gear pair: variation of transmission error with profile and lead

modification and misalignment

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3. For gear geometries as currently used, i.e. 35 µm crowning and 55 µm tip relief (see forthcoming section called 'Design of a well-balanced gearpair'), the effect of eliminating misalignment would be to reduce TE by a factor of 4.

Fig. 7 - Low helix gear pair misaligned 70 µµµµm hard forward: variation of load distribution factor with


Fig. 8 - High helix gear pair misaligned 70 µµµµm hard forward: variation of load distribution factor with profile and lead modification and helix angle

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Fig. 9 - high helix gear pair: variation of load distribution factor with profile and lead modification and misalignment

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OPTIMISING LOAD DISTRIBUTION FACTOR

The total load distribution factor KHβα is the ratio of peak load intensity Pmax at any phase of mesh to the load intensity which would be observed in perfect and inelastic gears, perfectly aligned. With tangential load Ft:

αβα εbF

PK

tH ∗= max (5)

The load distribution factor is primarily affected by mesh misalignment, crowning and helix angle, and to a lesser extent by tip relief.

Investigating first the effect of tip relief and crowning on the high and low helix angle gears at the worst expected mesh misalignment of 70 µm, Figs 7 and 8 show the load distribution factor for the low helix angle gears with β = 6, 8, 10o and for the high helix angle gear with β = 28, 30, 32o respectively. It is noted that:

1. For all six helix angles KHβα is greatest with zero crowning, as would be expected.

2. Minimum KHβα is achieved with a crowning height of about 45 µm, and zero tip relief, when KHβα is 2.6 for the low helix angle gears and 2.1 for the high helix angle gears.

3. Comparing Figs 7 and 8, the range of crowning and tip relief over which low load distribution factor is achievable is far greater in the case of the high helix angle gears than the low helix. In particular at β= 28o (Fig 8(a)) low KHβα is achievable for alI crowning heights from 35 to 120lm for a range of tip reliefs from 0 to 30 µm.

4. Comparing Figs 8(a) and 8(c), the worst load distribution factor for zero tip relief drops from 8.4 at β = 28o to 7.5 for the higher helix angle gear at β = 32o.

Improving the alignment of gears has a significant effect on the achievable load distribution factor. This is shown in Fig 9, where the load distribution factor is shown as a function of tip relief and crowning for the 32o helix angle gear at misalignments of 0.5 µm and 70 µm, all still at 25% FLT.

It is noted that:

1. At zero misalignment, the minimum load distribution factor, i.e. the minimum stress, is not achieved at zero crowning as would be intuitively expected, but at 25 µm crowning when KHβα = 1.7, compared to KHβα = 2.5 at zero crowning and any value of tip relief. Optimal tip relief is again in the range 0 to 10 µm, much less than would normally be specified for a gear of this type.

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2. At 35 mm misalignment, optimal corrections are 55 mm crowning, and zero tip relief, when KHβα =1.8, i.e. no significant increase in stress when the gearbox is optimally crowned for 35 mm misalignment.

3. At 70 µm misalignment, optimal corrections are 40 µm crowning with zero tip relief, when KHβα = 2.1.

It can therefore be concluded that, although the 'shape' of the load distribution graph varies significantly with misalignment, the effect of misalignment on load distribution and hence stressing is not so significant provided the gear is optimally designed for that misalignment. However, if the gear is not designed to accommodate misalignment, and is of typical current design, for example zero crowning and 60 µm of tip relief, then the load distribution factor will increase from KHβα =25 at zero misalignment to KHβα = 75 at 70µm misalignment.

Fig. 10 - Specific load distribution on wide-faced helical gears

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Fig. 11 - Variable bearing load component versus phase

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Fig. 12 - Low helix gear pair misaligned 70 µµµµm hard forward: variable bearing load component versus


Fig. 13 - High helix gear pair misaligned 70 µµµµm hard forward: variable bearing load component versus


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MINIMISING QUASI-STATIC BEARING LOAD

In addition to noise and vibration excited by dynamic bearing loads resulting from dynamic forces caused by TE, the bearings are subjected to quasi-static load variations resulting from changes in the axial position of contact between gear teeth at different phases of mesh. These quasi-static bearing load variations occur once per tooth mesh, i.e. at tooth contact frequency, and can be significant in the total dynamic load spectrum

In Figs 10(a) and 10(b) the load distribution along two teeth in contact on the low helix angle gear (β= 8o) is shown at two phases of engagement (0.1 and 0.8 base pitches). With only two teeth engaged, the 'main' tooth is seen to be sharing load alternately with the preceding and succeeding tooth. This results in a cyclic variation of 9 mm in the axial position of the resultant, and changes in the bearing reaction forces. For the high helix angle gear (β = 30o) Figs 10(c) and 10(d) show the load distribution along four contact lines at phase 0.0 and 0.5. Sharing the load between more teeth results in a much smaller change in bearing reaction forces, the axial position of the resultant varying by only 1 mm.

Figure 11 shows the quasi-static bearing load as a function of phase of mesh. The high helix angle gear generates a bearing load variation of only 30N, while the low helix gear generates ± 320N at tooth contact frequency (Nb: The nominal Load at each bearing at 4000 Nm torque is 20 000 N. The quasi-static bearing load variation in the low helix gear is thus still only ± 1.6% of bearing load but this is quite large compared to other dynamic forces).

There is significant scope for optimising gear geometry to minimise quasi-static bearing load variations. Figures 12 and 13 show the rms quasi-static bearing loads for the low and high helix angle gear, for 6o, 8o and 10o and, 28o, 30o and 32o helix respectively, again as a function of crowning and tip relief (Nb: quasi-static bearing load amplitude

2= . rms quasi-static bearing load).

Comparing Figs 12 and 13, it is noted that

1. All three low helix angle gears achieve minimum quasistatic bearing load variation at very low tip relief with a crowning of 70 to 120 µm. In this region, the bearing load is below 75 Nm rms (± 106N).

2. For the more usual design of such a gear, say 30 to 40 µm crowning and 30 µm tip relief. the quasi-static bearing load variation at β = 8o is about 220 rms (± 311N).

3. The high helix angle gear Fig 13 shows a completely different pattern of quasi-static bearing load variation with tip relief and crowning. Minimum bearing load (about 4N rms ± 6N) is achieved at about 80 µm crowning and β = 30o, which does not change significantly with changing hp relief

4. With the well designed high helix angle gear, the quasistatic bearing load is only 5% that of the optimised low helix angle gear.

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DESIGN OF A WELL-BALANCED GEAR PAIR

In the examples considered above, it has been easy to select a gear geometry for minimum transmission error, minimum load distribution factor and minimum quasi-static bearing load variation considering each parameter individually. In designing a real gear, a compromise must be achieved between conflicting requirements for high load capacity, low noise and good insensitivity to misalignment over the full operational speed-torque envelope.

The requirements for the gear design will depend on operating mode.

In the 'patrol' speed range, typically to 25% or 30% of full power, the emphasis is on minimum noise and vibration, but strength considerations are secondary.

In the 'sprint' speed range, towards 100% full power, noise and vibration are not so important, but strength considerations, i.e. a minimum load distribution factor, are paramount.

Considering the design of the gears in relation to the 'patrol' speed range, i.e. 25% of full power, firstly, it is very clear that the high helix angle gear is very much quieter than the low helix gear, so that only the optimisation of gear geometry for this gear will be considered.

In patrol mode, minimum TE and minimum quasi-static bearing load variation are primarily required. From Figs 5 and 13, it is seen that in areas of low KHβα and low quasi-static bearing load, tip relief generally has no significant effect other than at β = 32o (Fig 5(c)) where an absolute minimum TE is achieved at zero tip relief and 70 µm crowning. It is noticed that 'bands' of lowest TE correspond to 'bands' of higher bearing load and vice versa. Since in this case quasi-static bearing load is quite small, it is appropriate to minimise transmission error and select a gear with:

Helix angle β = 32 Tip relief = 5 µm Crowning = 6 µm

For this gear geometry:

TE =0.13 µm (< 0.3 arc sec) Quasi-static bearing load variation =18N rms

For this 'optimised' gear geometry, the load distribution factor, which is of secondary importance at part load, would be (from Fig 8(c)):

KHβα = 2.2

It is interesting to compare the performance of this 'optimised' gear design with the typical gear geometry. which would have been selected for a high helix angle gear based on typical current design rules, viz:

1. integer face contact ratio, in this case β = 5 at β = 28.11o ;

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2. crowning height = 0.5 x design misalignment, e.g. 0.5 x 70 = 35 µm;`

3. tip relief equivalent to maximum elastic mesh deflection at full load plus pitch and involute profile tolerance, e.g. in this case = 55 µn tip relief.

For this gear geometry, from Figs 5,13 and 8:

TE = 0.55 µn Quasi-static bearing load variation = 50N rms Load distribution factor KHβα = 3.5

The gear designed to current 'rule of thumb' is thus 50% worse in terms of load distribution factor, but TE is four times greater and quasi-static bearing load variation is about three times greater. It could thus be expected that the 'optimised' gear would have a longer life and be significantly quieter than a gear designed to current 'best practice'.

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EXPERIMENTAL VERIFICATION OF ELASTIC MESH ANALYSIS

The predictions of static transmission error, quasi-static bearing load variation and load distribution described above have been checked against experimental data measured on the Royal Navy's Marine Gearing Research Rig5.

Fig 14 General arrangement of test rig (part section)

This research rig has tested the low and high helix gears, with the geometry detailed in Table 1, at torques to 15000 Nm and powers up to 8MW. This test rig, shown schematically in Fig 14, has the facility for measuring both static and dynamic transmission error to very high accvracy16 and also for measuring both static and dynamic bearing forces with a discrimination of 0.1N at high frequency5. In addition, all the test gears used were strain-gauged with 24 or 32 gauges across the facewidth to measure operational load distribution.

Good correlation was established between predicted and measured static transmission error, bearing load and load distribution. Dynamic transmission error and dynamic bearing forces were, however, dominated by torsional and flexural modes of vibration which enhanced and attenuated the static mesh excitation, resulting in very complex dynamic behaviour. In spite of this it was found that certain general relationships held true. In particular, the validity of the assumption that calculated, static TE correlated well with gear noise and vibration as characterised by the dynamic bearing forces, was proven to hold true for a wide range of gears and operating conditions. Figure 15 shows the dynamic, out of resonance, bearing load plotted against calculated (static) TE, for the low and high helix angle gears considered in this paper. In spite of a ratio of about 10:1 between TE in low and high helix angle gears, and a ratio of nearly 10:1 between TE with 0 and 100 µm misalignment, the measured dynamic bearing load per unit of calculated TE varied by less than 3:1. Over a wide range of misalignment, for both high and low helix angle gear, one micron of calculated TE resulted in a dynamic bearing force of (on average) 600N. This linear relationship has been proved to be true for two

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types of gear geometry (with different lead and profile modifications), two pinion bearings (forward and aft) and two torque conditions (2000 and 4000 Nm).

Ih - Low helical gears ..................... 600 N rms / µm line hh - High helical gears h - Ahead s - Astern pf - Pinion Forward Bearing ps - Pinion Aft Bearing

BEARING LOAD

• Test measurements taken away from resonance conditions - rms value at tooth passing frequency

• Results include fully aligned, -50µm/200mm, -l00µm/200mm, and +l00µm/200mm misalignment conditions at a pinion torque of 2000Nm and 4000Nm.

TRANSMISSION ERROR

• Computed FE model results

Fig. 15 - Measured variation of bearing versus computed transmission error

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CONCLUSIONS The FE based elastic mesh model described has been shown to be a useful tool for improving the geometric design of marine gearing. In particular, tooth profile and lead modification may be optimised for a range of torques and misalignments to minimise quasi-static transmission error and bearing load variation, while maintaining acceptable load distribution and hence gear stress. The validity of the premise that static TE is a valid criterion for gear noise and vibration has been confirmed, and the calculations validated against experimental data.

'The technique described ran be applied to single and double helical gears, with thrust bearings or thrust cones, and is capable of modelling all significant geometric and elastic conditions to tooth contact behaviour. Properly applied, it can achieve significant reduction in gear noise and vibration.

ACKNOWLEDGEMENTS

The authors wish to thank the MOD(N) for their assistance with the development of the calculation procedure described in this paper.

REFERENCES

1. 'The effect of gear accuracy on noise and vibration ; NAVGRA Report No 68/19.

2. 'The dynamic performance of a small gearbox with helical gears', NAVGRA Report No 70/3.

3. 'Noise trials on the AE18"centres test rig', NAVGRA ReportNo 73/10.

4. 'Dynamic analysis of the 2000 K gear test rig', NAVGRA Report No 74/7.

5. S J Thompson ct al, 'A four megawatt test rig for gear noise and vibration research', Proc International Gearing Conf, pp 445/451, Newcastle (1994).

6. J Rosinski et al, 'Development of a new three dimensional dynamic model of helical gears', Proc International Gearing Conf, pp 517/524, Newcastle (1994).

7. H K Kohler, 'The measurement and mechanism of dynamic loading in spur gears', PhD Thesis, Sheffield University (1959).

8. R G Munro,'The dynamic behaviour of spur gears', PhD Thesis, Cambridge University (1962).

9. DU unpublished report

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10. 'Requirements for gearing in main propulsion', Naval Engineering Standard NES 305 Issue 4 (September 1991).

11. 'Load capacity of parallel axis gears', Draft Standard ISO 6336, International Standards Organisation.

12. 'Gearing', PartS, Chapter 5, Main and Auxiliary Machines, Lloyds Register of Shipping (LRS) Rules and Regulations for the Classification of Ships, Revised (1990).

13. J H Steward, 'Elastic analysis of load distribution in wide-faced spur gears, PhD Thesis, Universitv of Newcastle upon Tyne (1989).

14. J H Steward, 'The compliance of solid wide-faced spur gears', Trans ASME /n! Mech Des II2, pp 59�595 (December 1990).

15. C D Haddad and J A Pennell, 'A PC based program for three dimensional elastic analysis of load distribution in wide-faced spur and helical gears', Proc 3rd World Congress on Gearing & Power Transmission, pp 20I-211 (Paris 1992).

16. J Rosinski et al, 'Dynamic transmission error measurement in the time domain in high speed gears', Proc International Gearing Conference, pp 363-370 (Newcastle 1994).

NOMENCLATURE

n number of intervals of contact line (mid point will be contact points in a discretised model)

i index of compliant point

j index of loaded point

k index of no-load contact point

ei total geometric error at i' th point (+ve = material added)

cti theoretical clearance at i' th point (only non-zero outside of theoretical contact path)

gi no-load gap at i' th point

di unknown elastic deflection at i' th point

Cij compliance at i' th point due to load at j' point

wj unknown specific load on j' th interval of contact line

lj length of j' th interval of contact line

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F total normal load

Ft total tangential load

b facewidth

Pmax maximum specific load intensity

β helix angle

εβ face contact ratio

εα transverse contact ratio

KHβα total load distribution factor ( = KHβ . KHα Q to ISO 6336)

BGA Training Notes

TN7: Controlled Shot Peening of Gears and Transmission Components


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BGA Training Notes TN7 CONTROLLED SHOT PEENING OF GEARS AND TRANSMISSION COMPONENTS

Authors: Peter O'Hara and Graham Hammersley Metal Improvement Company Inc. Hambridge Lane, Newbury, Berkshire, RG14 5TU. United Kingdom [email protected] http://www.metalimprovement.com

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Index

CONTROLLED SHOT PEENING OF GEARS AND TRANSMISSION COMPONENTS ................. 3

1.0 CONTROLLED SHOT PEENING THEORY............................................................................... 3 INTRODUCTION ...................................................................................................................................... 3

a) Residual Compressive Stress ...................................................................................................... 4 b) Surface Roughness ...................................................................................................................... 5 c) Surface Hardness ........................................................................................................................ 6 d) Grain Size ................................................................................................................................... 6 e) Microstructure. ........................................................................................................................... 6 f) Ultimate Tensile Strength............................................................................................................ 6

2: CONTROLS OF SHOT PEENING ...................................................................................................... 8 INTRODUCTION ...................................................................................................................................... 8

1.0 Media control.............................................................................................................................. 8 2.0 Intensity Control. ...................................................................................................................... 10 3.0 Coverage Control...................................................................................................................... 12 4.0 Equipment Controls. ................................................................................................................. 13

3.0 APPLICATIONS OF CONTROLLED SHOT PEENING......................................................... 15 INTRODUCTION .................................................................................................................................... 15

1.0 Surface Treatments ................................................................................................................... 16 2.0 Welding ..................................................................................................................................... 17 3.0 Electro-discharge machining (EDM or “Spark Erosion”) ....................................................... 17 4.0 Grinding.................................................................................................................................... 18

EXAMPLES OF COMPONENTS WHICH HAVE BENEFITED FROM SHOT PEENING................. 19 COMPRESSOR/TURBINE BLADES .................................................................................................... 19 CONNECTING RODS. ......................................................................................................................... 20 CRANKSHAFTS.................................................................................................................................... 20 SHAFTS AND AXLES........................................................................................................................... 22

4.0 TRANSMISSIONS AND CONTROLLED SHOT PEENING.................................................... 24 1.0 BENDING AND TORSIONAL FATIGUE.............................................................................................. 24 2.0 PITTING OR CONTACT FATIGUE. .................................................................................................... 25 3.0 CARBURISED AND NITRIDED STEELS. ............................................................................................ 25 4.0 DECARBURISATION........................................................................................................................ 27 5.0 GRINDING ABUSE. ......................................................................................................................... 29 6.0 SHOT HARDNESS. .......................................................................................................................... 29 7.0 RETAINED AUSTENITE. .................................................................................................................. 31 8.0 CORRECTION OF DISTORTION BY SHOT PEENING........................................................................... 32



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CONTROLLED SHOT PEENING OF GEARS AND TRANSMISSION COMPONENTS

1.0 CONTROLLED SHOT PEENING THEORY

INTRODUCTION

Controlled Shot-Peening is the cold working of a surface with spherical particles of known mass, impinging at a pre-determined velocity.

The effect of the technique is to:

a) Induce a residual compressive stress. b) Alter surface roughness. c) Increase surface hardness. d) Reduce grain size. e) Alter microstructure. f) Increases Yield Strength.

These surface modifications are effective in preventing surface failures from taking place and has been used successfully against:

• fatigue • fretting fatigue • fretting • stress corrosion cracking • corrosion fatigue • galling • spalling • cavitation erosion • porosity

These will be discussed later.

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Fig. 3 - EXAMPLE OF RESIDUAL STRESS PROFILE INDUCED BY SHOT PEENING

A) RESIDUAL COMPRESSIVE STRESS

The action of bombarding a surface with shot stretches the material which yields at and near the surface but the core resists plastic work and a residual compressive stress is induced. The magnitude and depth of the compressive stress varies with materials, but generally a level of 50% - 60% of the Ultimate Tensile Strength of the base metal is achieved with the depth varying from 0.1 to 2.5 mm. See Figures 1 and 2.

Figure 3 demonstrates a typical stress profile created by shot peening on materials with no prior stress condition. Experience on many materials indicates that the four key points on the profile are; - surface stress / peak compressive stress / neutral and core tensile stress. Variations in the parameters and controls of the process will vary these levels and depths; therefore thorough specifications and execution of the technique is critical.

Fig. 1 - RESIDUAL STRESS INDUCED BY SHOT PEENING -v- TENSILE STRENGTH OF STEEL

Fi.g 2 - DEPTH OF COMPRESSIVE RESIDUAL STRESS -v- ALMEN ARC HEIGHT

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B) SURFACE ROUGHNESS

Impinging a surface with shot varying in size from 50 to 3000 micron and at velocities varying from 20 to 200 metres/second will result in a considerable variation in final roughness. However it is not just the roughness that is changing in magnitude, it is also the type of finish. A multi-directional finish is generated which is good in terms of fatigue especially when machining lines are at 90o to the applied stress and in cases where leakage is noted on high pressure pneumatic/hydraulic seals.

Another advantage is that surface growth will be noted which can be used to increase the size of a shaft or decrease the size of a bore. Applications exist where on repair of bearing bores that have gone oversize, peening generates sufficient growth that new parts can be pressed in satisfactorily.

Fretting occurs at the areas of contact of two metals which are essentially stationary with respect to each other, but in fact are in relative motion. This movement can cause those areas under pressure and contact, not separated by a lubricant film to micro weld. Fretting is also referred to as fretting wear, fretting corrosion, false brinelling, friction oxidation, chafing fatigue and wear oxidation. In these cases Shot Peening can have a beneficial effect as a result of work hardening the surface through cold work and roughening the surface to eliminate or reduce slip at the interface.

However roughening is perhaps an inaccurate term and generation of a random or multidirectional surface more appropriate. Two conventional machined finishes in contact can sit peak to peak and slip. Two Shot Peened surfaces will resist movement and with the increased skin yield strength, fretting resistance is improved. Examples of areas Shot Peened with success include flanged, splined or keyed assemblies. Automotive applications include shell bearing and the joint faces of connecting rods. Although fretting applications exist, fretting fatigue is the greatest concern as premature failure can result. Generally speaking initiation occurs through microcracks developed at the fretted surface and accelerated by wear particles and debris produced at the contact face.

The influence of surface finish is shown in Figures 4 and 5 on aluminium alloy. Trials were conducted in the plain and fretting fatigue conditions. Plain fatigue strength after Shot Peening improved by 70% and fretting fatigue strength after Shot Peening improved by 300%. However tests were also conducted removing the surface finish produced by shot peening by polishing. In the plain fatigue tests the results improved, whereas in fretting fatigue the results reduced. In plain fatigue a finer finish can enhance product life further, hence duplex Shot Peening. In fretting fatigue, loss of the surface 'key' is detrimental which is a point for Designers to recognise when fretting situations arise.

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C) SURFACE HARDNESS

Shot peening is a cold working process and therefore a change in surface hardness will result. For instance on EN58a with a core hardness of 230 Vickers, the surface hardness can be lifted to 370 Vickers.

On Stellite thrust runners with a core hardness of 415 Vickers the surface hardness was increased to 630 Vickers. On separate tests on Stellite faced rotary crushers made from Boiler Plate, used to process fullers earth, the following hardness figures were noted.

Boiler Plate - Before Shot Peening 165Hv - After Shot Peening - 272Hv

Stellite - Before Shot Peening 564Hv - After Shot Peening 786Hv

D) GRAIN SIZE

Cold work of a surface will result in the break up of the surface grains and grain boundaries and provide a multitude of slip planes and dislocations. It has been postulated that this in itself is beneficial in regarding fatigue cracks but one area where considerable success has been noted is in the prevention of Intergranular Corrosion.

I.G.C. of Stainless Steels results from heating that material to its 'sensitized' level around 480 to 800 oC which causes the precipitation of chromium carbides in the grain boundaries. Chromium is then depleted adjacent to the grain boundaries and the area is now susceptible to I.G.C. Cold working the surface prior to 'sensitizing' prevents precipitation of the chromium carbides and therefore chromium depleted boundary structures will not exist and I.G.C. will be retarded.

E) MICROSTRUCTURE.

Cold work of surfaces by Controlled Shot Peening or indeed machining, will result in microstructural changes in steels.

In austenitic stainless steels the level of ferrite will increase with extended levels of shot peening and in heat treated high carbon, or carburised steels, transformation of retained austenite to martensite will occur and is particularly advantageous on transmission components ie gears and splines. The amount of transformation is difficult to quantify and like determining the improvement in fatigue life with shot peening, it depends on where you start.

F) ULTIMATE TENSILE STRENGTH.

Controlled Shot Peening has no effect on the ultimate strength of the product as a whole, however cold working will increase its surface hardness as stated earlier and therefore the yield strength of that surface. That increase in surface yield must have a

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knock on effect in fatigue strength improvement alone in addition to induced residual compressive stresses.

The influence each of the above has on surface failure problems varies and as yet is not quantified, but it is generally accepted that the residual compressive stress has the greatest benefit. It is important to note when viewing the stress factors that the following simple formula applies in all cases.

ACTUAL STRESS = (APPLIED STRESS + RESIDUAL STRESS) x STRESS INTENSITY

where:

ACTUAL STRESS is that witnessed by the part. APPLIED STRESS is that applied by the operating load. RESIDUAL STRESS is that induced by the manufacturing technique. STRESS INTENSITY is the stress concentration factor from the component geometry.

This demonstrates that when assessments are being made of the benefits that can be attained from Shot Peening it really depends on the initial stress condition (often unknown) and the surface geometry. Components with a smooth, stress free condition can show marginal improvements whereas components with high manufacturing tensile stresses (not uncommon) and geometry with notch factors of 2 or 3 shows considerable improvements. Indeed the following chart highlights the geometry aspects very well.

Table 1 - Influence of pre-stretching on fatigue limit of notched and smooth bars.

Rotating bending staircase tests to 1 million cycles. Major dia. 0.843": root dia. 0.500": root radius 0.037" or 0.015": included angle 60 degrees. Material E4340 steel.

Code Notch Tensile Fatigue Limit Fatigue Limit

Strength Not Stretched Pre-stretched

ksi ksi % ksi %

201S 3.2 125 24 100 55 230

202S 2.15 125 30 125 58 240

201H 3.2 250 28 117 88 370

202H 2.15 250 35 146 90 375

200S 1 125 59 240 58 245

200H 1 250 91 380 92 385

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2: CONTROLS OF SHOT PEENING

INTRODUCTION

From what is known of the dimensions, material and service loadings on a given component we can determine an optimum shot peening specification for fatigue enhancement. However, since there is no non-destructive method to determine proper shot peening on a part, it is essential to ensure that the process is controlled to ensure repeatability. The four critical variables of the process which must be controlled are:

• MEDIA

• INTENSITY

• COVERAGE

• EQUIPMENT

1.0 MEDIA CONTROL.

There is a very wide range of shot peening media available and it is important to select the right shot for the job. Cast steel shot is used most commonly but glass beads, ceramic media and stainless steel cut-wire shot are also used.

Cast steel shot is available in sizes from 0.007" to 0.078" diameter, and comes in two hardness grades which are commonly referred to as "Regular" and "Hard". The former ranges from Rockwell "C" scale hardness of 45-52Rc and the latter from 55-62Rc. Generally speaking it is important to use a shot which is at least as hard as the component to be peened.

Glass beads range in size from 0.002" to 0.033" and ceramic media from 0.006" to 0.033".

For a given choice of media it is important that the size of shot is maintained within strict limits since the energy delivered, and hence the depth of compressive stress induced will vary with ball size if velocity is maintained and constant. The screening tolerances for cast steel shot are shown in Table II.

Equally important is the shape of the media - which should be as close to spherical as possible. Shot which is mis-shapen, elongated or indeed faceted or broken by the peening process must be extracted from the shot flow. This is best achieved by the use of a shot classifier which comprises a cambered spiral stainless steel chute somewhat reminiscent of a Helter-Skelter.

Shot is fed over a spreader cone at the top of the spiral chute, the dimensions of which are calculated to ensure that only the acceptable shapes (near spherical) will gain sufficient speed on its descent to achieve the escape velocity and ride over the

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cambered edge. Damaged, mis-shapen or broken shot will tend to slide, rather than roll, down the spiral and will emerge from the bottom into a waste bin.

1/ - Sieve numbers specified in RR-S-366, number in parenthesis represents sieve opening size (inches) Source AMS 13165

Modern shot peening machines are equipped with sieves and classifiers in their shot-return arrangements to ensure continuous control of the quality of the shot. Figure 4 shows poor size and shape distribution compared with an acceptable media sample and also magnified sections through the surface of a component peened by each type of media.

Fig. 4 - (A) UNACCEPTABLE SHOT PEENING MEDIA (B) ACCEPTABLE SHOT PEENING MEDIA

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Figure 5 shows the potential damage through incorrect shot control (A) and the surface finish feasible (B) through correct shot control.

Fig. 5. (A) SURFACE DAMAGE CAUSED BY POOR SHOT CONTROL (B) ACCEPTABLE FINISH THROUGH GOOD SHOT CONTROL

2.0 INTENSITY CONTROL.

Before peening components it is necessary to calibrate the shot peening machine to ensure that the correct impact energy will be imparted to the component to be processed. This is referred to as the peening Intensity and is a function of shot size, hardness, density, velocity and angle of impingement.

Since the alteration of any of these variables will have an effect on the peening intensity it is important to have a calibration method which takes account of all of them. This is achieved by peening test strips of spring steel which are held down onto a rigid steel block. Only the upper surface of the strip is processed, and since peening is a cold working process the surface is stretched as a result of the induced compressive stresses and plastic deformation of the surface. The result is that the test strip is curved - convex in the direction from which it was peened. The arc height at the centre of curvature is measured and is expressed in thousandths of an inch "Almen". J.O. Almen

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of the General Motors Research Laboratories developed the method which, along with the strips and the arc height measurement still bear is name.

Almen strips are 3" long 3/4" wide and come in three thickness grades depending on the intensity range and therefore the sensitivity of calibration required. The thickness grades of Almen strip are designated as follows:

N = 0.031" A = 0.051" C = 0.0938"

In practice it is necessary to peen successive Almen strips for increasing periods of time and to plot the arc height measured against time. This plot is referred to as a "Saturation Curve" and is shown in Fig 6.

Fig. 6. SATURATION CURVE

The term “Saturation is used because the initial linear relationship between arc-height and the exposure time reaches a limit beyond which further peening has a much diminished effect on arc-height. This "limit" is known as the saturation point and is defined as that time at which doubling the exposure time will affect the arc-height by no more than 10%.

When referring to the Almen intensity of shot peening we always mean the arc-height at the saturation point. The necessary adjustments must be made to the machine to ensure that the saturated intensity falls within the specified tolerance range of peening intensity.

Diminishing angles of impingement diminish the arc-height since the energy imparted by the shot stream to the surface is directly proportional to the sine of the angle of impingement. For parts of complex shape it is therefore necessary to position a number of Almen blocks so as to simulate the various faces which will require peening to the specified intensity.

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Fig. 7. THE ALMEN STRIP SYSTEM

3.0 COVERAGE CONTROL

Coverage is defined as the extent to which the original surface of the workpiece has been obliterated by the shot peening, and is expressed as a percentage.

It is important to note that this property must be measured on the workpiece and not on the Almen strip which could be different from the hardness of the part and hence exhibit a different degree of coverage.

Coverage determination is carried out visually, using ten-power magnification, or by a fluorescent dye process, PEENSCAN.

Fig. 8. SHOT PEENING COVERAGE: (A) PARTIAL (B) COMPLETE

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PEENSCAN is particularly useful for components which are very large, are complex in shape or have a very high surface hardness reducing visual evidence of peening.

The component is coated with the tracer liquid and is peened for a period of time, after which it is examined in a darkroom under ultraviolet (UV) light. Any bright spots witnessed signify incomplete peening coverage.

4.0 EQUIPMENT CONTROLS.

It is essential that the peening process is repeatable one part to the next, and to achieve this, shot flow, compressed air, mechanical movements, electronic and (increasingly) computercontrols must be set and maintained with failsafe intelligence built-in.

Computer monitored shot peening equipment which represents state of the art technology in repeatability of parameters have been developed and are used extensively in critical applications.

Figures 9 and Table III show the PEENAMATIC software flow path and a printout of the parameters set on the computer respectively

Fig. 9. SOFTWARE PATH FLOW DIAGRAM

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Table III - TYPICAL COMPUTER PRINTOUT DOCUMENT

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3.0 APPLICATIONS OF CONTROLLED SHOT PEENING

INTRODUCTION

There are numerous applications of the process of controlled shot peening, but mostly they are concerned with fatigue of one form or another.

The discovery that shot peening led to an improvement in fatigue properties stemmed from efforts in the late 1920's by the Buick Motor Company to remove oxide scale from valve springs by grit blasting. This improved the fatigue properties of the springs, and subsequent research showed that further improvements in fatigue resulted from the use of spherical media which was well controlled in terms of its size, hardness and angle and velocity of impact.

The research also proved that the improvement in fatigue was not solely the result of work hardening of the surface or scale removal, as had been the initial theory, but the induced residual compressive stresses.

Many metal-forming, welding and machining processes result in a combination of high residual surface stresses and sharp notches in the surface of components. The action of shot peening induces compressive stresses at and beneath the surface to a depth below the level of notch-like features from which cracks would tend to propagate.

From this it can be seen that components to benefit the most from shot peening are those which will experience a degree of bending involving high surface strains. Consequently springs exhibit the most dramatic improvement closely followed by gears and many other common engineering components.

We shall be discussing gears and transmissions in more detail later, but here we will consider the various manifestations of fatigue and how controlled shot peening has been shown to counteract them.

Most metals benefit from shot peening since they can all retain residual compressive stresses of varying magnitude, but it is useful to consider the various forming and heat treatment operations which have a bearing on fatigue performance.

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1.0 SURFACE TREATMENTS

The application of chromium and electro-less nickel plating has a serious deleterious effect on the fatigue life of components since the plating material is brittle and cracks very readily. These cracks can propagate rapidly into the underlying material and lead to premature failure. Shot peening prior to plating induces a compressively stressed region in the surface of the component through which cracks cannot propagate.

Fig. 10. PLATING CRACKS WILL NOT PROPAGATE THROUGH THE PRE-STRESSED BASE.

Fig. 11. ROTATING BENDING FATIGUE. CYCLES TO FAILURE OF 4340 STEEL (52-53 HRC)

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2.0 WELDING

Tensile stresses of very high magnitude - often approaching the yield point of the material - are present in and adjacent to the Heat Affected Zone (HAZ) of weldments.

Shot peening to induce compressive stresses is a very effective means of extending the fatigue life of welded components, and is even more beneficial if accompanied by "toe grinding" to reduce the stress concentration effect.

3.0 ELECTRO-DISCHARGE MACHINING (EDM OR “SPARK EROSION”)

Although spark-erosion does not of itself impart any force on the bulk of the material, molten metal solidifying at the surface leaves a hard, brittle skin and attendant tensile stresses approaching the Ultimate Tensile Stress (UTS) of the material.

Shot peening is used to good effect in restoring the fatigue strength of parts formed in this way.

Electro-Chemically Machined (ECM) and Electro-Polished parts are also candidates for shot peening to restore fatigue performance.

Fig. 12. SUMMARY OF HIGH CYCLE FATIGUE BEHAVIOUR OF INCONEL 718, SOLUTION TREATED AND AGED (HRC44)

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4.0 GRINDING

Residual tensile stresses and surface brittleness can be caused by the generation of high surface temperatures during harsh grinding operations with consequent reductions in fatigue life. Gentle grinding will improve the situation considerably, but the fatigue strength of components that have suffered severe grinding can be restored by shot peening to a degree higher than that associated with gentle grinding.

Fig. 13. SHOT PEENING IMPROVES ENDURANCE LIMITS OF GROUND PARTS - REVERSED BENDING FATIGUE OF FLAT BARS. HARDNESS 45HRC

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EXAMPLES OF COMPONENTS WHICH HAVE BENEFITED FROM SHOT PEENING.

The following examples are by no means exhaustive but serve to illustrate a cross section of engineering components which must endure a variety of differing loading regimes and for which shot peening has been found to increase life or reliability.

COMPRESSOR/TURBINE BLADES

Blades for use in both the compressor and turbine sections of gas turbine engines and in steam turbines benefit greatly from shot peening as do their respective location slots in the turbine disks.

Such blades are subjected in service to a variety of loadings and several different shot peening applications are applied to counteract the effects of each.

Blade roots are peened to prevent fretting, galling and fatigue. Broaching of the root form can leave tears, pits and scores which are stress raisers. Shot peening induces an envelope of compressive stress around the entire surface of these often complex shapes and refines the surface to a depth below which the machining marks have any influence. Aero engine blades are often peened during overhauls as well as when they are new.

The airfoils of blades are shot peened to provide an envelope of compression against the effects of impact damage from foreign bodies, pitting and stress corrosion cracking. Additionally many coating and plating processes essential to the performance of turbine airfoils tend to decrease fatigue life. Shot peening prior to these operations will often return a performance which excels that of uncoated blades in fatigue terms.

Fatigue cracking can initiate at the lacing wire holes on steam turbine blades and shot peening in these holes with a fine lance can obviate this problem.

Fig. 14.

SUPPRESSION OF FATIGUE DAMAGE OF INCONEL 713C TURBINE BLADES

BY SHOT PEENING

SHOT PEENING AS A MEANS OF OVERCOMING PRIOR FATIGUE

DAMAGE TO 4340 STEEL

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CONNECTING RODS.

Conventional wisdom would dictate that polishing to a very high surface finish is the route to be taken to ensure maximum performance from connecting rods. However it can be shown that a rough surface in compression will resist failure better than a smooth surface in tension. There is no advantage in producing a fine finish prior to peening, and indeed of the two operations, peening will tend to be the cheaper and certainly more cost effective solution. If a fine surface finish is felt to be necessary say for aesthetic purposes, then fine lapping, honing or polishing is possible after shot peening so long as removal is restricted to a depth which only just removes the witness of peening. In practice, for components of average hardness, this will be less than 10% of the compressive layer. Connecting rods are frequently shot peened before the bores are machined. A good shot peening procedure will avoid distortion of the rod enabling full machining first. This bestows additional benefits such as avoidance of fretting fatigue between the bearing shells and the large end bores of the con-rod.

CRANKSHAFTS

The most highly stressed area of a crankshaft is the crankpin bearing fillet which sees its highest stress on the underside when the pin is in the top dead centre position of the firing stroke.

Shot peening to a relatively high intensity in this fillet has a great improvement on fatigue strength, as instanced below in which a 6 cylinder crankshaft forged in 434011 material exhibited an increase in fatigue strength from 52KSI to 72KSI - an increase of 38%.

Shot peening can be applied to all sizes of crankshafts, from motorcycle engines to large marine diesels, and is effective on forged and cast steels, and nodular and Austempered Ductile Irons. Fatigue strength increases in design of up to 30% are permitted on shot peened crankshafts by marine underwriters Det Norsk Veritas.

Fig. 15. CRANKSHAFT AND DETAIL

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Fig. 16. INCREASE IN FATIGUE STRENGTH OF SHOT PEENED CRANKSHAFT

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SHAFTS AND AXLES

A wide variety of these components are shot peened to enhance fatigue properties. Areas such as splines, keyways, changes of section and fillets are typically of those which require peening.

Some shafts have a "shear section" designed into them to protect the equipment which they drive. These areas can lead to premature failure through fatigue cracking, but can be shot peened to enhance fatigue life while retaining the overload protection since the shear stress will not be affected by the shot peening process.

Fig. 17. FATIGUE TESTS ON REAR AXLE SHAFTS

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Fig. 18. FATIGUE TESTS ON NOTCHED SHAFTS

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4.0 TRANSMISSIONS AND CONTROLLED SHOT PEENING

Historically belief in the benefit and use of Controlled Shot Peening has been on materials where the dimpling effect is clearly visible and that "harder" materials would have little response to cold work. In fact in practice the complete opposite is the case. Softer materials, for example mild steel, improve in fatigue strength through cold work by perhaps 7% - 10%. High hardness materials can improve in fatigue strength by over 30%.

Fig. 19. demonstrates the above by showing that there is a reduction in fatigue strength with increasing surface hardness, especially when this exceeds 35 - 40 HRC. Even nitrided steels as hard as the high 60's HRC will respond well to shot peening, although to the untrained eye there is no discernable change in surface finish.

Fig. 19. COMPARISON OF PEENED AND UNPEENED FATIGUE LIMITS FOR SMOOTH AND NOTCHED SPECIMENS AS A FUNCTION OF ULTIMATE TENSILE STRENGTH OF STEEL

1.0 Bending and Torsional Fatigue

Torsion bars or shafts witnessing torsional loads can also improve in fatigue resistance as demonstrated by Figure 20. The principal reason bending and torsional applications respond well to shot peening is that in both cases the maximum applied loads are at the surface and this is the area affected by the process. Tensile fatigue cases can be treated by shot peening, however the amount of improvement is less noticeable as in

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these cases the applied load is uniform throughout the section and the surface is the only area affected by the process, albeit the most sensitive area.

Fig. 20. COMPARISON OF FATIGUE CURVES - BENDING v TORSIONAL LOADING

2.0 Pitting or Contact Fatigue.

Work on rolling element bearings has shown that increased residual compressive stresses will increase rolling element fatigue life. Traditionally this has not been an area where shot peening has been used but studies are indicating that the potential may be there for further research.

When surface pitting problems arise, not root bending, then the shot peening parameters must be reviewed to reduce surface roughening and increase surface compressive stress only. N.A.S.A. conducted work in this field and demonstrated that the Shot Peened gears exhibited pitting failure lives 1.6 times that the life of standard gears without Shot Peening under identical conditions.

3.0 Carburised and Nitrided Steels.

These heat treatments will result, if properly conducted, in induced surface residual compressive stresses. The action of Shot Peening on these materials will be to increase the magnitude of compressive residual stress further. Figures 21 and 22 highlight the improvements feasible on carburised and carbonitrided steels. These additional levels of stress in bending fatigue will show fatigue strength improvements of 20% -30%.

Figs. 21 and 22 are S/N curves on bending fatigue tests on case carburised and carbonitrided gears respectively.

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Fig 21. S-N CURVES OF BENDING FATIGUE TESTS ON CASE HARDENED GEARS OF 3, 5 & 8 MODULE, PEENED AND UNPEENED.

Fig. 22. ROTATING BENDING FATIGUE STRENGTH OBTAINED BY CARBONITRIDING AND HARD SHOT PEENING.

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4.0 Decarburisation.

Decarburisation is defined as the loss of carbon from the surface of a ferrous alloy as a result of heating in a medium that reacts with the carbon at the surface. It has been shown that decarburisation can reduce the fatigue strength of high strength steels (240KSI and higher) by 70% to 80% (Fig. 23) and lower strength steels (140KSI to 150KSI) by 45% to 55%. It is generally accepted that decarburisation is a surface phenomenon not particularly related to depth. A depth of .003inch decarburisation can be as detrimental to fatigue strength as a depth of .030 inch. However, the amount of decarburisation can have a dramatic influence on fatigue properties of carbon steels. Severe decarburisation can induce significant residual tensile stresses in the surface part as shown in Fig. 24.

Shot peening has proven to be effective in restoring most, if not all, of the fatigue strength lost due to decarburisation.

Decarburisation, though softer, is often not easily detectable and Shot Peening is a way of ensuring integrity of the parts if case problems arise. In addition the treatments act as a complete surface hardness tester and any areas suffering decarburisation will be highlighted. It has been known for the consistency of shot peening to be put in question and for the true reason to eventually be revealed, viz., hard and soft areas resulting from decarburisation. It is an excellent quality control check of the complete surface.

Fig. 23. EFFECT OF SHOT PEENING ON DECARBURISATION

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Fig. 24. EFFECT OF DECARBURISATION ON THE RESIDAL STRESS DEVELOPED IN CARBURISED & HARDENED PLATES. THE CARBON CONTENT AT 0,002MM WAS ESTIMATED TO BE 1%, 0.54% AND 0.35% FOR CURVES 1,2 & 3 RESPECTIVELY

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5.0 Grinding Abuse.

Grinding of surfaces is one of the manufacturing techniques that can lead to considerable fatigue scatter. Temper burns or rehardening cause grave concern to many organisations hence the obligatory demand for subsequent chemical etching to expose the problem on critical parts. The problems of grinding causing temper and rehardening causes variations in hardness and variation in surface stress.

Shot Peening will yield the surface material, even rehardened areas, and induce compressive residual stresses recovering the fatigue loss. Figure 25 demonstrates the ability of Shot Peening to recover S156 material from temper burns.

Fig. 25. RECOVERY FROM GRINGING TEMPER ON S156 MATERIAL BY SHOT PEENING.

6.0 Shot Hardness.

All metallic materials can and are shot peened for one reason or another. The action of causing surface deformation generates the benefits mentioned previously. However there comes a point when the transfer of energy from the dynamic component (the shot) is not transferred so readily to the static component (the product - gear, shaft, whatever). The result is that the shot deforms as much and in certain cases more than

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the product. Where deformation of shot and product are equal is approximately when each are of equal hardness. Standard shot is in the region of 45Rc which when used on case carburised surfaces (60Rc) results in well deformed shot. Although a beneficial residual compressive stress is induced it is not as expressed earlier in the levels of 50% - 60% of the U.T.S. Consequently industry has learnt that to achieve the best possible fatigue improvement the shot used must be as hard or harder than the substrate being treated. Hence the availability of high hardness shot in the region of 62Rc giving the benefits shown in Figure 26.

Fig. 26. PEENING 1045 STEEL AT HRC 48 WITH MI 330 SHOT

Fig. 27. PEENING 1045 STEEL AT HRC 52 WITH MI 330 SHOT

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7.0 Retained Austenite.

Retained austenite is essential in the microstructure of transmission components. What is in question is the amount remaining in the final state and its influence on performance. No attempt will be made to postulate on optimum levels as it is outside the remit of this presentation, however it is important to note that retained austenite readily converts to martensite under cold work. During transformation the volume expansion which accompanies the austenite to martensite reaction will in itself increase the level of compressive residual stress in addition to the normal stretching of the substrate traditionally experienced through shot peening. That conversion to martensite will also be noted as a considerable increase in surface hardness, of great benefit if wear is a problem.

Fig. 28. CHANGE IN RETAINED AUSTENITE CONTENT DISTRIBUTION OF CARBONITRIDED STEEL BY HARD SHOT PEENING

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8.0 Correction of Distortion by Shot Peening

Peen correction is a technique derived from Controlled Peen Forming which is a dieless forming process performed at room temperature. Selected shot peening parameters produce in combination compressive stress and plastic deformation which can develop compound, convex or concave curvature. Using additional shot peening techniques, single curvature, saddle back and conical profiles can be achieved.

After Peen Forming all parts also exhibit an increase in fatigue life or stress corrosion resistance, unlike many other forming techniques. Peen Correction is therefore applied as a reversal of the forming technique ie if parts distort on machining, heat treatment or fabrication they can be corrected using compressive stresses. Three or four point bending as an attempt at correction uses tensile stresses and yielding through a components section. On very high steels and complex sections this can result in residual tensile stresses or cracks in parts. Peen correction aims to achieve movement with induced compressive stresses and invariably an improvement in crankshafts and circular parts such as rings, gears, etc. which have become oval in manufacture.

Fig. 29. CONTOUR CORRECTION OF RING GEAR (A) (B) BEFORE PEENING (C) AFTER PEENING

(A) (B) (C)

M etal Im provem ent Com pany

Stretching Stress P eening

C urving

Shot P een F orm ing Princip les

MECHANICAL POWER TRANSMISSION SERIES

Lecturer's notes dealing with the manufacture and use of mechanical power transmissions.

Ref: Subject Author

TN 1-1 Introduction to Gear Noise J.D.Smith

Cambridge University

TN 2-1 Gearbox Systems: Miniature Instrumentation J. Rosinski, D.A. Hofmann

University of Newcastle upon Tyne

TN 2-2 Gearbox Systems: Problems & Solutions D.A. Hofmann

University of Newcastle upon Tyne

TN 3 Worm Gears Cedric Barber, Dr. Mike Fish

TN 4-1 Gear Cutting Tools – Part 1 W Clark *

* The presentation contains photographs and CAD data courtesy of: David Brown Gear Systems.

TN 4-2 Gear Cutting Tools – Part 2 Anthony Hardwick

TN 5 Steel Selection: The Manufacture of G. Haywood

Engineering Steels

TN 6-1 Bearing Noise and Vibration – Gear D.A. Hofmann

Geometry Optimisation University of Newcastle upon Tyne

TN 6-2 Gear Noise and Vibration – A new tool for P. Maillardet. D.A. Hofmann. M E Norman

designing main propulsion gears MoD, Uni of Newcastle,

TN 7 Controlled Shot Peening of Gears and Peter O'Hara, Graham Hammersley

Transmission Components Metal Improvement Co Inc

1

BRITISH GEAR ASSOCIATIONTECHNOLOGY TRANSFER PROGRAMME 2000

GEAR CUTTING TOOL SEMINAR

Held at David Brown Group Plc, HuddersfieldWednesday 25th October 2000

Speakers:

W. Clark C.Eng. MIMech.E

A. Hadwick C.Eng. MIEE

2

GEARSPOWER TRANSMISSION

3

GEARSTHE INVOLUTE

INVOLUTE CURVE

BASE CIRCLE DIA.

4

GEARSNON-INVOLUTE GEARS

CONFORMAL GEARS

WHEEL

PINION

5

GEARSINTERESTING QUOTATIONS

The practical consideration of cost demands the formation of gear teeth upon some interchangeable system, The cycloidal system cannot compete with the involute, because its cutters are formed with greater difficulty and with less accuracy, and a further expense is entailed by the necessity for more accurate centre distances. Cycloidal teeth must not only be accurately spaced and shaped, but their wheel centres must also be fixed with equal care to obtain satisfactory results.Wilfred Lewis.

There is no more need of two kinds of tooth curve for gears of the same pitch than there is need for two different threads for standard screws, or two different coins of the same value, and the cycloidal tooth would never be missed if it were dropped altogether. But it was first in its field, is simple, and has the recommendation of many well-meaning teachers, and holds its position in ‘human inertia’, or the natural reluctance of the average human mind to adopt to change, particularly change for the better.George B. Grant ‘Treatise on Gearing’ 1890

6

GEARSPITCH

CIRCULAR PITCH

PITCH CYLINDER (REFERENCE CYLINDER)

DP = π / CP MOD = CP / π

7

GEARSPRESSURE ANGLE

αα = PRESSURE ANGLE AT REF. DIA.

αα

αα

8

GEARSBASE PITCH

9

GEARSDIAMETER RELATIONSHIPS

CIRCULAR PITCH (d)

DEFINITION OF CP AND PA AT DIA. (d)

10

GEARSDEVELOPMENT OF

GEAR PITCH CYLINDER

11

GEARSAXIAL PITCH

12

GEARSGEAR DATA

No. Teeth

Normal Pi tch (Ref . )

Normal PA (Ref . )

Hel ix Angle (Ref . )

Hand of Hel ix

L e a d

Outs ide D iameter

Chamfer (EAP) Dia .

Star t of T ip Rel ie f Dia .

Pi tch Dia. (Operat ing)

Pitch Dia. (Ref.)

Start of Root Rel ief Dia .

Start of Act ive Prof i le Dia.

Root Dia .

Base D ia .

Normal Arc Tooth Thickness (Ref.)

No. Teeth

Outs ide D iameter

Chamfer (EAP) Dia .

Mat ing Centre Distance (Min. )

Gear Data

Mating Gear Data

13

GEARSGEAR PROFILE

14

GEARSGEAR PROFILE

15

GEAR MANUFACTUREMETHODS

FINISH CUT

CUT & SHAVED

CUT & ROLLED

CUT & GROUND

CUT & SKIVED

CUT SKIVED & GROUND

16

HOBBING

17

HOBBINGHOB TYPES

MONOBLOC GROUND FORM

MONOBLOC UNGROUND FORM

INSERTED BLADE (IB)

18

HOBBINGHOB DESIGN FEATURES

OUTSIDE DIAMETER

LENGTH

COLLARS/HUBS

BORE

GASHES

SUBSTRATE MATERIAL

COATINGS

BASIC RACK

QUALITY

MEANS OF DRIVING

NUMBER OF THREADS

HAND OF THREADS

19


20


21

HOBBINGTHREADS / GENERATION

1 Thread

22


2 Threads

23


3 Threads

24

HOBBINGFLUTES / GASHES

25

HOBBINGCAM

26

HOBBINGRELIEF

27

ALL TOOLSSUBSTRATE MATERIALS

GRADE APPROXIMATE COMPOSIT ION

C Cr M o V W Co

M 2 0.83 4.13 5.00 1.98 6.13

M 2, HIGH C 1.00 4.13 5.00 1.98 6.13

M 35 0.80 4.00 5.00 2.00 6.00 5.00

M 42 1.10 3.88 9.50 1.15 1.50 8.25

S 390 1.60 4.75 2.00 5.00 10.50 8.00S 590 1.30 4.20 5.00 3.00 6.30 8.40

S 690 1.33 4.30 4.90 4.10 5.90

S 790 1.30 4.20 5.00 3.00 6.30

ASP 2023 1.28 4.20 5.00 3.10 6.40

ASP 2030 1.28 4.20 5.00 3.10 6.40 8.50

ASP 2060 2.30 4.00 7.00 6.50 6.50 10.50

EM 2 0.83 4.13 5.00 1.98 6.13

EM 35 0.80 4.00 5.00 2.00 6.00 5.00

EM 42 1.10 3.88 9.50 1.15 1.50 8.25

28


GRADE WEAR TOUGHNESS RED

RESISTANCE HARDNESS

M 2 50 70 40

M 2, HIGH C 50 + 70 - 40

M 35 55 35 60

M 42 70 40 80

S 390 90 60 85S 590 70 70 75

S 690 75 65 60

S 790 70 70 60

ASP 2023 65 80 50

ASP 2030 80 70 80

ASP 2060 95 55 85

EM 2 50 70 + 40

EM 35 60 60 + 75

EM 42 70 40 + 80

N.B. ALL FIGURES FROM STEEL SUPPLIERS

29


30

ALL TOOLSCARBIDE

31

ALL TOOLSMATERIAL / COATINGS

0

2 0

4 0

6 0

8 0

1 0 0

C u t t i n g S p e e d T o o l L i f e

M 2 - U n c o a t e d M 2 - C o a t e d

A S P 3 0 - U n c o a t e d A S P 3 0 - C o a t e d

C a r b i d e - C o a t e d

32

HOBBINGBASIC RACK

33

HOBBINGCAD OUTPUT

34

HOBBINGCAD OUTPUT

35

HOBBINGCAD OUTPUT

36

HOBBINGCAD OUTPUT

37

HOBBINGQUALITY

38

HOBBINGQUALITY

39

HOBBINGTESTING

40

HOBBINGTESTING

41

HOBBINGTESTING -

MULTIPLE THREAD HOBS

42

HOBBINGOPTIMISATION

THE ‘MODERN’ HOB

SMALL DIAMETEREXTENDED LENGTH

MULTI-THREADMULTI-GASH

HIGH QUALITY SUBSTRATECERAMIC COATINGS

REVERSE / CONVENTIONAL METHOD OF HOBBING

43

HOBBINGOPTIMISATION

44

HOBBINGOPTIMISATION

45

HOBBINGOPTIMISATION

46

HOBBINGOPTIMISATION

47

SHAPING

48

SHAPINGGSC DESIGN FEATURES

NUMBER OF TEETHOUTSIDE DIAMETER

LENGTH / WIDTHPROOF BAND

BORE / SHANKTYPE OF SHARPENINGSUBSTRATE MATERIAL

COATINGBASIC RACK

QUALITY

49


50

SHAPINGGSC DESIGN

FEATURES

51

SHAPINGHELICAL GUIDES

52


53

SHAPINGSHARPENING

54

SHAPINGSHARPENING

55

SHAPING

SUBSTRATE MATERIALS

COATINGS

56

SHAPINGTHROUGH-GRINDING

57

SHAPINGTHROUGH-GRINDING

58

SHAPINGBASIC RACK

59

SHAPINGCAD OUTPUT

60

SHAPINGCAD OUTPUT

61

SHAPINGCAD OUTPUT

62

SHAPINGCAD OUTPUT

63

SHAPINGCAD OUTPUT

64

SHAPINGCAD OUTPUT

65

SHAPINGCAD OUTPUT

66

SHAPINGCAD OUTPUT

67

SHAPINGQUALITY

68

SHAPINGQUALITY

69

SHAPINGOPTIMISATION

70

SHAPINGOPTIMISATION

71

SHAVING

72

SHAVINGSHAVER DESIGN FEATURES

NUMBER OF TEETHHELIX ANGLE

OUTSIDE DIAMETERWIDTH

SERRATIONSCLEARANCE GROOVESSUBSTRATE MATERIAL

COATINGDESIGN LIFE

QUALITY

TYPES:-CONVENTIONAL, DIAGONAL, UNDERPASS, PLUNGE

73

SHAVINGOPERATION

74

SHAVINGOPERATION

75

SHAVINGSERRATIONS

76

SHAVING

SUBSTRATE MATERIALS

SURFACE TREATMENTS

77

SHAVINGCAD OUTPUT

78

SHAVINGCAD OUTPUT

79

SHAVINGCAD OUTPUT

80

SHAVINGDESIGN GRAPH

81

SHAVINGDESIGN

SHAVING CUTTERTOOTH

SERRATIONS/TIP WIDTH

OK

WRONG

82

SHAVINGCAD OUTPUT

83

SHAVINGCAD OUTPUT

84

SHAVINGCAD OUTPUT

85

SHAVINGREGRIND INFORMATION

86

SHAVINGQUALITY

87

SHAVINGQUALITY

88

SHAVINGOPTIMISATION

89

SHAVINGOPTIMISATION

90

SHAVINGOPTIMISATION

91

1

PART TWOCHOICE The various methods we might expect

to use in gear production will be discussed.

SELECTION How we go about deciding which method we should use.

TOOL MATERIALS What should our gear tool be made of?

TOOL COATINGS Should we consider surface treatments?

SPEEDS AND FEEDS It is important to use the tool efficiently.

QUALITY How to achieve the accuracy required.

GEAR TOOL Ensure you always get the best fromMAINTENANCE your tool.

2

CHOICE

There are a number of ways to produce gears, and the final choice of tool will rest on the equipment available to you and the design of the part. Other influential factors will be the number of gears to be produced and the quality to be

attained.

MOST COMMON PRODUCTION METHODS

•HOBBING•SHAPING•PLANING •MILLING•CBN GRINDING•FORM ROLLING

3

CHOICE

FINISHING APPLICATIONS

•SHAVING•GRINDING•HONING

The production methods mentioned can be further split into two groups, as follows

Gear Generation & Gear Forming.

4

5

SELECTION FACTORS

•Volume of gears to be produced

•Final quality of gear

•Final hardness/hardness when cut

•Machines available

•Basic method to be used

•Is sub-contract required or necessary

6

GSC TYPES

7

OPTIONS•Conventional/Standard hob. Often available from stock

•Small diameter long length. These are usually multi gash design in special PM HSS and coated, many are designed as multi start type. (for use on high speed CNC hobbing machines) ••Larger diameter longer length. In this case the hob diameter is calculated to operate at the optimum surface speed to suit the material being cut and close to the top RPM of the hobbing machine. These can also be multi-start for further improvements in productivity (for use with slower conventional hobbing machines)

Other types of hobs which are used far less often in modern factories and are as follows:

•Inserted Blade. There are very few manufacturers still producing these as the ‘Modern Hob’ out performs them in most volume applications.

8

OPTIONS

Solid Carbide. Recent results have been encouraging but very careful consideration must be taken when choosing these. Firstly a

specially designed machine is required; most CNC machines are not suitable, as the high temperature swarf is required to be removed

directly from the machine. Carbide hobs are brittle and good housekeeping is essential. Even hob sharpening can be a problem as quite often two separate angles are required on the cutting face.

•Carbide Skiving Hobs. Special purpose tools for hard hobbing of larger type hardened gears.

•Throw away Hobs. These are similar to the modern hob but have so many gashes that there is no opportunity to re-sharpen them.

They appear to be an expensive option.

9

HSS STRUCTURE

10

PM MANUFACTURE

11

K94 ASP2060

92

90

88

86

84

82

80

78

76

74

72 T15 ASP2052

70 REX T15 REX T76

68 M15

66

64 CPM M4

62 M4 T6

60 ASP2023 ASP2030

58 M3

56 T4

54 T1 M7

52 M42

50 M2 M35

48

We 25 30 35 40 45 50 55 60 65

K= Wear ResistanceWe = Hot hardness

It should be noted that the materials indicated in bold type are

the preferred powder metallurgy grades and these should be used

wherever coatings are to be added. PM HSS has a better structure than conventional grades of HSS – it is

tougher, performs better and grinds more easily.

It is important to consider each material's toughness in addition to

the above information.For example, ASP 2060 has a

higher hot hardness coupled to wear resistance but it also has a poor toughness that can result in

tools chipping under certain circumstances.

12

S690 &CPM M4

S390 &ASP 2052

ASP 2060

ASP 2030

ASP 2023

M35

M2

0 5 10 15 20 25 30 35 40 45 50 55 60

TOUGHNESS COMPARISON OF HSS GRADES

13

MATERIAL REQUIREMENTS FOR MODERN GEAR TOOLS

HIGH STRENGTH Obtained from CARBON + ALLOYS

GOOD TOUGHNESS Obtained from the STRUCTURE OF POWDERMETALURGY HSS

HIGH WEAR RESISTANCE High Vanadium gives HARD CARBIDES

GOOD HOT HARDNESS Obtained from high COBALT

GOOD COATING SUBSTRATE Even grain structure with hardness and strength,again from POWDER METALURGY HSS

14

INFLUENCE OF ELEMENTS IN HSS

CARBON (C) ESSENTIAL FOR MARTENSITETRANSFORMATION - THE

ULTIMATE HARDNESS OFHSS, ALSO THE AMOUNTAND TYPE OF CARBIDE.

CHROMIUM (CR) HARDENABILITY –WEAR RESISTANCE

TUNGSTEN (W) CARBIDES _ WEAR RESISTANCESECONDARY HARDNESS

MOLYBDENUM (MO) 1% MO = 2% W

VANADIUM (V) STABLE HARD CARBIDES

COBALT (CO) HOT HARDNESS – IMPROVES TEMPERRESISTANCE

15

COMPARATIVE HSS CHEMICAL ANALYSIS

C Cr Mo W Co VM2 0.9 4.2 5.0 6.4 Nil 1.9

M3-2 1.2 4.1 5.0 6.2 Nil 3.0

M4 1.3 4.2 4.5 5.5 Nil 4.0M35 0.9 4.2 5.0 6.4 4.8 1.8

M42 1.1 3.8 9.4 1.5 8.0 1.2

T1 0.75 4.1 Nil 18.0 Nil 1.1

T4 0.75 4.1 Nil 18.0 5.0 1.1

PM 23 GRADES 1.3 4.1 5.0 6.4 Nil 3.1PM 30 GRADES 1.3 4.2 5.0 6.4 8.5 3.1

S390 GRADES 1.6 4.8 2.0 10.5 8.0 5.0

PM 60 GRADES 2.3 4.2 7.0 6.5 10.5 6.5

Items shaded yellow These are the important powder metalurgy HSSgrades.

Items shaded red Have cobalt for hot- hardness.

Items shaded blue These have vanadium content for hard carbides,giving wear resistance.

16

COATING COMPARISONTiN TiCN TiAlN TiAlN

‘FUTURA’ ‘X.TREME’

Hardness Vickers 2300 3000 3000 3500

Max. Operating Temp. 600 C 400 C 800 C 800 C

Colour yellow/ blue/ violet/ purple/Gold grey grey grey

TiN Coating is used as the standard coating

TiCN Coating is used more for abrasive materials and higher performance

TiAlN ‘FUTURA’ is ideal for high temperature conditions, and is one of thecoatings being used in dry hobbing trials. It is a multi layer type coating.

TiAlN ‘X.TREME’ is recommended for skive hobbing (carbide tools). This is amono coating.

17

HSS STRUCTURE

18

CUTTING SPEEDS FOR HSS GEAR TOOLS

19

FREQUENCY CURVE

20

WEAR PATTERNS

21


1. Very High Surface Speed

2. High Chip Loads

3. Increased Tool Life

4. Environmentally Friendly

5. Reduced Cost Per Part

6. Reduced Tool Cost Per Part

7. Higher Productivity than Carbide

22


Comparison Test Data on Liebherr LC153 Hobbing Machine

Gear Data

2.54 Mod 20 deg. PA. 56 Teeth 14.133 deg. HA151.76mm Outside Diameter 25mm Face widthMaterial SAE 8620 (225 BHN 21 HRC)

Hob Material Solid Carbide Acealloy 240

Coating TiN Balinit Futura

Hob Diameter 80mm 75mm

Hob Tooth Length 162mm 180mm

Number of Gashes 19 16

Number of Starts 2 2

Cutting Speed 280 M/M 240M/M

Hob RPM 1114 1019

Feed Per Rev. 2.25mm 3.10mm

Chip Thickness 0.131mm 0.180mm

Cutting Time 35Sec . 24.9Sec.

Parts Per Hour 86 118

Parts Per Regrind 774 941

Tool Life 4.0M 5.0M

23

QUALITY

Gear quality will depend upon a number of factors

• The type of process used

• The quality of the tool being used

• How well the tool has been set up

• How well the tool has been maintained

• The condition of the machine tool

24

QUALITY

25

26

27

28

29

30

31

RE-SHARPENING

• THE CORRECT SHARPENING ANGLES• RUN OUT• QUALITY OF GRINDING• SURFACE FINISH ON THE CUTTING FACE

WITH GEAR HOBS THERE ARE EVEN MORE FACTORS TO CONSIDER

• GASH RADIALITY• GASH TO GASH SPACING• GASH LEAD

32

RE-SHARPENING

33

RE-SHARPENING

34

RE-SHARPENING

35

RE-SHARPENING

36

37

38

IN CONCLUSIONTool maintenance is a very important part of production. High cost savings canbe achieved when tool grinding is correct.

Tools sharpened to their original condition will:

• Produce gears to the correct profile.

• Maintain their maximum wear resistance.

• Accept coating after sharpening.

• Reduce scrap levels.

• Save time in setting (as the setter is not producing poor quality from in builtrun-out).

Documents

BGA Training Notes TN1-1bga.org.uk/wp-content/uploads/2017/03/BGA-mechanical... · 2018. 8. 30. · gear teeth in contact. Alternatively as we think of involutes as corresponding