t

THE LIBRARY

OF

THE UNIVERSITY OF CALIFORNIA

GIFT OF

Dr. William H. Ivie

THE GASOLINE AUTOMOBILE

Its Design and Construction

VOLUME II

Transmission, Running Gear and Control

By

P. M. HELDT Technical Editor of The Horseless Age

Second Edition

P. M. HELDT

Nyack, N. Y.

1917

Copyrighted by P. M. HELDT

1917 Previous Copyright, 1913.

GUT

"T vA

(ALL RIGHTS RESERVED)

PREFACE.

DURING the period that intervened between the original writ- ing of this volume and the present revision, a number of notable evolutions took place in the design of some of the component parts which are dealt with here. The most important of these was undoubtedly the introduction of the helical bevel gear drive. The adoption, of this drive confronted automobile engineers with new problems, chiefly in regard to bearing loads; these are discussed in some detail in the present edition and rules for the calculation of the bearing loads are given.

While the bevel-spur and the internal gear drive were both in use at the time the. first edition was prepared, only a single firm was prominently identified with each in the United States, so they were not deemed of sufficinent importance to warrant special treatment. Since then, however, the internal gear drive has made notable progress in this country and the bevel-spur drive has assumed some importance in England. At the same time addi- tional interest has been aroused in the four wheel drive for mili- tary and similar trucks, so it was decided to add a chapter cover- ing these three forms of final drive.

The advent of the high speed motor, together with a great in- crease in the use of unit power plants, resulting in the lengthening of propeller shafts, has compelled designers to give more atten- tion to the problem of critical speeds in shafts. Some matter on this subject has been incorporated in the Chapter on The Bevel Gear Drive and Rear Axle, the theory of critical speeds being explained and rules for their calculation given.

Another branch of automobile engineering in which great com- mercial development has taken place during the past four years is that relating to the worm drive. The chapter devoted to this subject has been largely rewritten and brought up to date. Minor additions and changes have been made throughout the book, and a number of typographical and other errors that occurred in the first edition have been corrected. For pointing out such errors the author wishes to thank some of his readers.

M80S109

PREFACE.

It may appear that in the chapters on the Sliding Change Gear and on Rear Axles, the annular ball bearing receives more atten- tion than is warranted by the scale of its present day use. Owing — at least in part — to the interruption of imports of ball bearings from Europe, roller bearings now predominate largely in auto- mobile construction. The problems of mounting, however, are very much the same as with ball bearings, and numerous examples of mounting roller bearings are given in the plates at the end of the book as well as in 'the text illustrations. In the most expen- sive cars the annular ball bearing still retains a prominent place, and it was, therefore, not deemed necessary to rewrite this part of the work.

As nearly all of the old plates had to be discarded it was de- cided to incorporate the plates in the book itself. Chassis views are shown for the most part in half tone, so the line cuts show only chassis components and these can be presented on a suffi- ciently large scale on a 5j4x8j4 sheet.

THE AUTHOR.

LIST OF CHAPTERS

CHAPTER I. GENERAL LAYOUT OF CARS 3

CHAPTER II. FRICTION CLUTCHES 13

CHAPTER III. SLIDING CHANGE SPEED GEARS 70

CHAPTER IV. PLANETARY CHANGE SPEED GEARS 125

CHAPTER V. FRICTION Disc DRIVE 147

CHAPTER VI. UNIVERSAL JOINTS 160

CHAPTER VII. DIFFERENTIAL GEARS 180

CHAPTER VIII. UNIT POWER PLANTS, TRANSMISSION AXLES 193

CHAPTER IX. BEVEL GEAR DRIVE AND REAR AXLE ; 203

CHAPTER X. THE WORM GEAR DRIVE 293

CHAPTER XI. THE CHAIN DRIVE 323

CHAPTER XII. BEVEL-SPUR GEAR, INTERNAL GEAR AND FOUR WHEEL DRIVES. 341

CHAPTER XIII. BRAKES 357

CHAPTER XIV. FRONT AXLES 386

CHAPTER XV. STEERING GEARS 411

CHAPTER XVI. CONTROL 441

CHAPTER XVII.

FRAMES 471

CHAPTER XVIII. SPRINGS . 497

CHAPTER XIX.

ROAD WHEELS 528

APPENDIX 543

PLATES , 571

LIST OF PLATES.

CHANGE SPEED GEAR OF THE PACKARD TWELVE 571

DRY Disc CLUTCH OF CHALMERS 6-30 572

BROWN-LIPE DRY Disc CLUTCH ON CUNNINGHAM CAR 573

MARMON CONE CLUTCH 574

SIMPLEX LUBRICATED Disc CLUTCH 575

MUNCIE CLUTCH AND CHANGE GEAR 576

CASE CLUTCH AND CHANGE GEAR 577

BORG & BECK PLATE CLUTCH AND COVERT CHANGE GEAR 578

TIMKEN TRUCK FRONT AXLE (7200 LBS. MAX. LOAD) 579

"AMERICAN" PLEASURE CAR REAR AXLE 580

TIMKEN PLEASURE CAR REAR AXLE 581

TIMKEN WORM DRIVE TRUCK REAR AXLE 582

TORBENSEN INTERNAL GEAR DRIVE TRUCK AXLE 583

Two FRENCH INTERNAL GEAR TRUCK DRIVES 584

'AMERICAN" PLEASURE CAR FRONT AXLE 585

FRANKLIN STEERING GEAR 586

BENZ STEERING GEAR 587

PEERLESS TRUCK STEERING GEAR 588

SPICER PROPELLER SHAFT ASSEMBLY 589

FRANKLIN THROTTLE CONTROL ASSEMBLY 589

PLAN VIEW OF LIPPARD- STEWART 1000 LB. TRUCK CHASSIS 590 SIDE ELEVATION OF LIPPARD-STEWART 1000 LB. TRUCK

CHASSIS 591

WINTON SPARK AND THROTTLE CONTROL (ABOVE) AND

CLUTCH AND BRAKE CONTROL (BELOW) 592

PLAN VIEW OF INTERSTATE FOUR CYLINDER CHASSIS 593

LEXINGTON-HOWARD Six CYLINDER CHASSIS 594

PLAN VIEW OF LEXINGTON-HOWARD CHASSIS 595

CADILLAC EIGHT CYLINDER CHASSIS MODEL 55 596

PLAN VIEW OF CADILLAC EIGHT CYLINDER CHASSIS 597

PACKARD FIVE TON TRUCK CHASSIS 598

PLAN VIEW OF PACKARD FIVE TON TRUCK CHASSIS 599

PLAN VIEW OF STUDEBAKER FOUR CYLINDER CHASSIS 600

PLAN VIEW OF AUBURN FOUR CYLINDER CHASSIS 601

PLAN VIEW OF HUDSON SUPER- Six CHASSIS.. 602

CHAPTER I.

GENERAL STRUCTURE OF THE CAR.

Location of Motor — In the first attempts to build road vehicles propelled by gasoline motors the general lines of horse vehicles were followed. The latter were then regarded as the highest type of vehicular design, and any departure from their lines was thought to be undesirable, as it offended the eye. This made it necessary to place the power plant under the body, and considerable difficulty was often experi- enced in getting it into this cramped space. It was thought essential to conceal the mechanical part of the vehicle as much as possible, because what people wanted was a self- moving carriage and not a road locomotive or a machine akin thereto. Before long, however, some bold spirit stood up for the idea that the pov/cr plant deserved such a location on the vehicle that it could be designed without regard to the space available in the body, and that when it required attention it zould be reached quickly and without disturbing the passen- gers. The precedent then set has since been generally fol- lowed, and with very few exceptions the motor is now located at the front of the car under a bonnet. It is hardly neces- sary to add that the public's conception of what a motor ve- hicle should look like has greatly changed since then.

Spring Suspension of Power Plant — The early automo- biles built in this country, almost without exception, had reach rods or perches extending between the front and rear axles, the object of which was to free the body springs of the driv- ing thrust. Some designers then placed the power plant on these reaches, so as to simplify the problem of transmission to the wheels. It was soon recognized, however, that, even though pneumatic tires were used, the vibration was so strong that it was practically impossible to keep the motor intact. Moreover, the hammering effect of the heavy unsprung weight

3

4 GENERAL STRUCTURE OF CAR.

on the axles, wheels and tires greatly reduced the life of these parts. The principle was thus established that as much as possible of the weight of the car should be supported on springs, and above all the more delicate parts, such as the motor.

Number of Wheels — The great majority of all automobiles have four wheels. This is the minimum number which in- sures stability under all reasonable conditions. Howevei cars have been and are being built with as few as three and as many as eight wheels. The smaller number of wheels is used to reduce the manufacturing cost of small vehicles, while the larger numbers, above four, are used either to keep the load per wheel inside a certain maximum (as required by the road laws in some countries) or to insure greater comfort of riding. However, certainly more than 99 per cent, of all auto- mobiles (not including motorcycles) are of the four wheeled type, and this construction may be considered standard.

Steering and Driving — With the number of wheels decided upon, the question arises as to how many and which shall be used for steering, and how many and which shall be used for driving. With a four wheeled vehicle it is possible to steer with either the front wheels, the rear wheels or all four wheels, and to propel the vehicle by either one front wheel, one rear wheel, both front wheels, both rear wheels or all four wheels. In this connection it must be borne in mind that the effectiveness of both steering and driving depends upon the adherence — the resistance to slippage — between the wheels and the ground, which in turn depends upon the weight carried by the wheels. As far as steering is con- cerned, at least two wheels have to be used for it in a four wheeled vehicle, and if one-third or more of the total load is carried on these wheels, then the requirement of positive steering is met in a satisfactory degree. As regards the choice between the front and rear wheels for steering pur- poses, the front wheels possess one important advantage over the rear wheels, and tha* is that, if a car stands alongside of a curb or other barrier, and it is desired to drive away from it, with rear steering this can only be done by back- ing up, because in order to cause the car to turn away from it in driving forward, the steering wheels would have to be turned toward the curb and would run into it. Rear steer- ing was used fcr many years on electric cabs in New York City, but the disadvantage mentioned is greatly against it,

GENERAL STRUCTURE OF CAR. 5

and has been one of the points that led to its abandon- ment. The only advantage of four wheel steering would be that with a certain maximum deflection or "lock" of the steering wheels, a car with four wheel steering could turn in a much smaller radius than one with two wheel steering. Four wheel steering, however, would be subject to the disad- vantage of rear wheel steering referred to in the foregoing, and the further disadvantage of the complication involved in combined driving and steering wheels, which more than offset its slight advantage, and it is therefore never used.

As regards the number of driving wheels, it would greatly simplify the problem of transmitting the power from the motor to its point of application if only a single wheel was used for driving. The simplification which results from this arrangement, as compared with that in which two wheels are used for driving, is one of the main considerations which lead to the selection of three wheeled construction in certain in- stances. However, in order that a vehicle may have plenty of traction or road adherence under all conditions, even on steep grades with greasy road surface, at least 50 per cent, of the total weight to be propelled must be carried on the driving wheel or wheels. Besides, in the ordinary four wheeled vehicle, if power was applied to one wheel — in other words, at one side only — it would tend to cause the car to slew or skid easily and affect the steering unfavorably. Driv- ing through at least two wheels is therefore considered essen- tial to successful operation. As to whether the front or the rear wheels should be driven, one thing that is largely deter- mining in this matter is that the front wheels are used for steering, and it involves considerable mechanical complica- tion to use the same wheels for both driving and steering. Moreover, if the motor is located at the front end of the car it can more easily be placed in driving connection with the rear wheels than with the front wheels. It must be remem- bered that the motor is carried upon a spring supported frame, and therefore constantly changes its position with rela- tion'to the axles; this relative change in position must be allowed for by some form of flexible connection, and this can be done more easily if the motor is at a considerable distance horizontally from the axle to which it is connected in driv- ing relation. There are several real advantages in front driv- ing. Owing to the fact that the propelling force acts at a tangent to the circumference of the driving wheels, if the

6 GENERAL STRUCTURE OF CAR.

front wheels are drivers, and they drop into a mud puddle, for instance, they tend to climb out of it, as it were, while with rear drive the combined effect of the forward thrust of the rear wheels and the weight on the front wheels may force the latter deeper into the mud. Another advantage of front driving is that with it there is much less tendency to skid than with rear driving. The problem of driving through the steering wheels can, of course, be solved, but it involves the use of two universal joints, preferably of a type which insures uniform transmission of motion irrespective of the angle be- tween the connected shafts, which must be so placed that the point of intersection of the two connected shafts lies in the centre line of the steering knuckle pin.

Four Wheel Drive — Driving on all four wheels has been employed to some extent, particularly on army wagons, which may under conditions have to travel off the roads. Four wheel driving makes the whole weight of the vehicle and load available for traction purposes, which is an advantage when the streets are covered with ice or snow, or for some other reason are exceedingly slippery. This system of driving would become more important if steel tires should ever come into common use for commercial vehicles, since the adherence between steel and the different road surfaces is very much less than that between rubber and these road surfaces. Where rub- ber tires are used sufficient traction is obtained under all nor- mal conditions by so arranging the design that from one-half to two-thirds of the weight of the car and load is always car- ried on the driving wheels, while under abnormal conditions such traction devices as tire chains or steel studded tire covers are resorted to.

Thus, while the front drive and four wheel drive are being exploited to some extent, at least 99 per cent, of all automo- biles built are steered by their front wheels and driven by their rear wheels.

Differential Gear — If both driving wheels were positively connected to the single source of motive power, they could not rotate at unequal speeds, as is required in turning corners. If the wheels had to drive the car forward only, the problem could be solved by driving them through ratchet clutches, but since they must drive the car backward as well as forward, it is necessary to incorporate a so-called differential gear in the drive through which the driving torque is always equally divided between the two driving wheels, in driving both for-

GENERAL STRUCTURE OF CAR 7

ward and backward, and which allows the two wheels to turn at different speeds as required by the course followed, or by any slight difference in their diameters. With the four wheel drive it is necessary to use three differential gears, one be- tween the front and rear axles and one between the two wheels on each axle.

Friction Clutch— Owing to the fact that the gasoline motor, unlike steam and electric motors, does not start from a stand- still with full torque, but must be started either by means of a hand crank or some starting device which generally produces only sufficient torque to just turn the motor over against the compression, it is necessary to disconnect the motor from the driving parts of the vehicle for starting it, and after the motor has attained speed, to connect it to the vehicle again. For this purpose a device must be used which will allow of a certain amount of slippage, until the motor speed has been reduced and the vehicle speed increased to such a point that the two correspond. This is accomplished by means of a friction clutch, which is always placed close to the engine and gen- erally built together with the flywheel — except in those cars provided with frictional means of power transmission, such as belts, friction pulleys and friction discs, which latter devices serve the dual purpose of changing the gear ratio between the engine crankshaft and the road wheels and of disconnecting the former from the latter.

Change Speed Gear — With any but the very lightest of gasoline motor vehicles it is necessary to provide means for connecting the motor to the driving wheels in several different ratios. The gasoline motor differs from other light motors in that when running at its speed of maximum economy or its speed of maximum output, it produces nearly the maximum torque of which it is capable. The motor, of course, must be so geared that under normal conditions of operation — that is, when the car is traveling over a level road at a good speed — it runs at about its speed of maximum economy, and it is then impossible for the motor to provide the propelling effort re- quire'd in climbing steep hills or in passing through deep sand, through the same gear reduction. It is, of course, understood that when two shafts or other rotating machine parts are con- nected together in driving relation, the torques of the two bear to each other the inverse ratio of their respective speeds. Thus, by providing a hill climbing gear giving a speed reduction, say, four times as great as the normal speed reduction, the

8 GENERAL STRUCTURE OF CAR

driving effort at the road wheel rims can be quadrupled for hill climbing. But since the hill climbing or low gear gives a com- paratively low vehicle speed, it is customary in all but the lightest vehicles to provide either one or two intermediate gears, for use on moderate hills, on soft or uneven roads, etc. The change gear mechanism, therefore, provides either three or four forward gear ratios as a general thing, and also one re- verse gear ratio. In this country gear boxes with three forward speeds are considerably more common, while in Europe the four speed gear is the most popular, the difference being no doubt due to the fact that we employ relatively more powerful motors.

Single and Double Reduction — There is now a tendency to use a stroke of about 5 inches in motors of all sizes. Pleas- ure car motors make about 1800 revolutions at normal speed or at their speed of maximum output. For pleasure cars it is customary to use wheels of 30 to 36 inches diameter. If we assume that the wheels are 36 inches in diameter and that the car is to be geared to make 45 miles per hour at normal engine speed, then the wheels must turn at

45 X 5,280 X 12

i=~ — 2 --- -- = 420 r. f. m.

60 X 36X3.14

and the gear reduction ratio from the engine to the road wheels must be 1800 to 420, or about 4.25 to 1. This ratio can easily be obtained by means of a single reduction gear of the helical bevel type.

Now let us take the case of a heavy truck which has wheels of, say, 40 inches diameter and is to be geared to make 15 miles per hour at 1,200 revolutions per minute of the motor. The driving wheels must then turn at

o4o 3,14

Hence the gear reduction ratio must be 1,200 to 140, or 8.5 to i. This cannot be obtained in a practical way by a single bevel or spur gear set or a chain and sprocket gear, for the reason that the outside diameter of the driven gear or sprocket on the rear wheels or axle is limited, since the car must clear the road by a certain amount. This reduction can be obtained by means of a worm and worm wheel, but if either a bevel gear or chain drive is used, a double reduction is necessary. It is cus- tomary in such cases to employ a first reduction by bevel gears to a jackshaft and a second reduction by chain to the rear wheels, although occasionally the two reductions are obtained

GENERAL STRUCTURE OF CAR. 9

by means of one spur gear set and one bevel gear set, both con- tained in a housing on the rear axle. All pleasure cars employ a single reduction for normal speed operation, obtained by either a set of bevel gears, a chain and sprocket wheels or a worm and worm wheel. Commercial vehicles of the lighter type with pneumatic tires are geared the same, while the heavier commercial vehicles have either a single worm gear reduction or a double reduction by bevel gears and chains, by bevel gears and spur gears, or by bevel gears and internal gears.

Dead and Live Axles — The driving wheels may either,, be mounted upon bearings on the rear axle and driven through chains or spur gears, in which case the axle is called a dead axle, or they may be fixed upon the ends of driving shafts extending through the rear axle housing, or be placed in driving connec- tion with such shafts through driving dogs or positive clutches, in which case the axle is a live axle. Dead axles are used on a good many heavy commercial cars and live axles on nearly all pleasure cars. When a dead axle is used the rear wheels are driven from a countershaft — except in the very few cases where two motors are used — and the differential gear is mounted on the countershaft. With live axles it is customary to mount the differential gear at or near the middle of the axle, though some- times it is mounted on the propeller shaft through which the power is transmitted to the driving axle. Live axles may be driven through a chain and sprockets or through bevel, worm or spur gears. Only a single driving connection is required in the case of a live axle, while in the case of a dead axle there must be provided an individual drive to each of the driving wheels. Though the chain drive is applicable to live axles, and was at one time extensively used on low priced pleasure cars, it is now, as a rule, used only in connection with dead axles, in the form of the so-called side chains. Nearly all live axles of pleasure cars are driven through a set of bevel gears.

Frames— One of the rules which have been established in automobile design is that the vehicle body should be as inde pendent as possible of the mechanical part, so that it can be removed without disturbing any of the mechanical parts, motor, transmission and control members are, therefore, carried upon a substantially rectangular frame made of pressed steel, rolled steel or laminated wood, which is supported upon the axles through the so-called body springs. The motor sets upon this frame in front; in the conventional type of pleasure car chassis the motor space is walled in by the radiator in front

10 GENERAL STRUCTURE OF CAR.

and the dashboard at the rear, and the motor is covered by a sheet metal bonnet. This same arrangement is also used to some extent in heavy commercial vehicles, but for this class of vehicles there are two alternate arrangements, viz., having the driver's seat on top of the motor or at the side of the motor. In fact, there may be said to be still one more alternative, since the motor may be under the seat proper or under the footboard of the driver's seat. These latter arrangements make that por- tion of the length of the frame which would otherwise be oc- cupied by the driver's seat available for loading space.

In the conventional type of vehicle the space on the frame back of the dashboard is occupied by the body. It is one of the rules of design that no part of the mechanism back of the dash- board, except the control members, should project above the top plane of the frame. Formerly half elliptic springs were used almost exclusively and were often placed directly under- neath the frame side members, whose top edge was then made straight from end to end. Now, however, three-quarter elliptic and even full elliptic springs are widely used at the rear on pleasure cars, and in order to preserve a comparatively low centre of gravity, the frame has a drop directly in front of the point of attachment of the rear springs, or the springs, even if semi-elliptic, are placed outside the frame and there is a so- called "kick-up" in the frame directly over the rear axle, so it will not strike the latter when the springs are fully compressed. The frame side members are generally swept in at the front end in order to allow of a greater limiting deflection of the steer- ing wheels.

The so-called reach or perch has been entirely done away with. The rear axle transmits its driving thrust to the frame either through radius rods or the rear springs and the frame transmits driving thrust to the front axle through the springs.

Tread and Wheel Base — The distance between the centre lines of ground contact of the wheels on opposite sides is known as the track or tread. This distance is generally 56 inches in pleasure cars and the lighter commercial vehicles, and 62 inches or more in heavy trucks. The National Automobile Chamber of Commerce has adopted a standard of 56 inches for the tread of pleasure cars, but there is no standard for truck treads. Practically all light horse vehicles used in the northern part of the country have a track of 56^ inches, which is measured from and to the outside of the tires at the point of contact with the road, so an automobile with the standard 56 inch tread ^will run in ruts made by the wheels of horse vehicles. In the South a tread of 60 inches is much used on horse vehicles,

GENERAL STRUCTURE OF CAR. 11

and since many of the roads there are deeply rutted a large part of the year, several automobile manufacturers have been furnishing their cars with a 60 inch tread to customers in that part of the country, but the practice has been discontinued.

The distance between the centre of road contact of the front and rear wheels, respectively, is known as the wheelbase. This, of course, is the same as the centre distance between the axes of the front and rear wheel spindles. The wheelbase differs widely in different types and sizes of machines. A long wheelbase makes for a more comfortable riding car and also tends to prevent skidding. On the other hand, a long car can- not be handled so well in crowded streets, since it cannot turn in such a short radius. Besides, a long wheelbase car is necessarily comparatively heavy, since the frame must be made of larger cross section in order to support the same weight, as well as be made longer. In passenger vehicle practice there is a fair -degree of uniformity with respect to wheelbases, the latter ranging between the following limits for different types of cars.

Four cylinder runabouts and roadsters, 30 horse power and under, 90-105 inches.

Four cylinder runabouts and roadsters over 30 and not over 40 horse power, 105-115 inches.

Four cylinder taxicabs, 4-5 passengers, 96-100 inches.

Four cylinder touring cars, 30 horse power and under, 100- 115 inches.

Four cylinder touring cars over 30 and not over 40 horse power, 110-120 inches.

Four cylinder touring cars over 40 horse power, 120-130 inches.

Six cylinder cars, about 10 inches longer than four cylinder ones of the same class.

CHAPTER II.

FRICTION CLUTCHES.

The friction clutch, as already pointed out, serves the purpose of connecting the motor, after it has been started running, with the driving gear of the car, in such a way that the car may be gradually accelerated and the motor at the same time pulled down in speed, until the speeds of the two correspond, thus pre- venting shock and jar.

In motor cars employing a single friction clutch which serves to connect the engine to the driving wheels through all of the different gear reductions, the clutch is normally held in engage- ment by a spring or springs, and when it is desired to discon- nect the engine in order to stop the car, or to change the gear, the clutch is first disengaged by compressing its spring by means ot a pedal, then the gear is disengaged or changed, and finally the clutch is let in again. In other cars, where a clutch serves for a single gear reduction only, it is normally disengaged, and is engaged by pressure exerted on a hand or foot lever, the mechanism transmitting the pressure to the frictional surface of the clutch being self locking in the engaged position.

There are quite a number of different types of clutches, all more or less extensively used, viz.:

Conical clutches.

Multiple disc clutches.

Dry plate clutches.

Band clutches.

Coil clutches.

Expanding segment clutches.

Multiple disc and dry plate clutches are identical as tar as their general principle is concerned, but they differ 'n respect to detail of design. Dry plate clutches are in very extensive use on American cars, as are conical clutches. The latter are par- ticularly suited to cars of relatively low power. Lubricated disc clutches also are quite popular, especially in Europe. The other three types mentioned have been used more or less

12

FRICTION CLUTCHES.

13

in the past, but arc now seldom met with, the practice of assembling cars from parts built by specialists having tended toward the standardization of types. The different types of clutches will be taken up in succession.

Cone Clutch — Conical clutches may again be divided into three sub-classes, viz., the direct cone, the inverted cone and the double cone clutch. The direct cone is the oldest and most popular of these types. As shown in Fig. i, with this type the flywheel is bored out to form the female cone, into which the male cone is forced by the pressure of a coiled spring concentric

FIG. i. — DIAGRAM OF DIRECT FIG. 2. — DIAGRAM OF INVERTED CONE CLUTCH. CONE CLUTCH.

with its hub. In the inverted cone clutch (Fig. 2) the female cone is formed by a cast iron or steel ring bolted to the rim of the flywheel, into which the male cone enters from the flywheel or engine end. The inverted type of cone clutch was originally adopted in order to make it possible to place the change gear box nearer the engine, under the floor boards of the driver's seat, since the clutch spring is placed between the flywheel and the clutch cone, instead of to the rear of the latter. The double cone clutch is a combination of a direct and an inverted cone clutch, and is particularly suited where great powers have to be trans- mitted.

14 FRICTION CLUTCHES.

Clutch Calculations— In calculating any part of the trans- mission we will assume that the mean effective pressure in the engine cylinder multiplied by the mechanical efficiency (T? p ) is 80 pounds per square inch at low engine speeds and 65 pounds per square inch at the speed of maximum output. These figures are fairly representative though a little low for some engines. Now let & = bore of cylinder in inches.

/= length of stroke in inches. n — number of cylinders. p = mean effective pressure. P = mean total pressure on one piston. T- torque in pounds-feet.

Then the energy developed during one revolution of the crankshaft is

n TT I

E = ~ £2/^~I foot-pounds.

If there is a torque T on the engine shaft, or a turning effort of T pounds at a radius of I foot, the energy transmitted during one revolution is

E=2ir Tfoot pounds. Hence

and

nib2 p T - Ig2 pounds-feet ..................................... (i)

A diagram of a cone clutch is shown in Fig. 3. The spring pressure P forces the male cone against the female cone, pro- ducing a normal pressure N at their contact surface. According to the principle of the parallelogram of forces

P _ ^ — sin a,

hence

P — Nsin a. ................................................ (2)

where a is the angle of the clutch cone. The adherence or frictional force F between the clutch cones is equal to the normal pressure multiplied by the coefficient of friction /

F=Nf

Angle of Cone — Cone clutches faced with leather or asbestos fabric are given an angle of cone of from 10 to 13 degrees, but the most common angles are 12 and i2l/2 degrees. . With metal to metal clutches an angle of 10 degrees can be used without risk of trouble, but such a small angle in a leather

FRICTION CLUTCHES 15

faced clutch is liable to cause it to stick. From equation (2) it will be seen that the spring pressure P required to produce a certain normal pressure decreases as the angle of the cone decreases, hence there is an advantage in using as small an angle as practical. However, a cone with a relatively large angle is less given to "fierceness" in action, i. e., sudden gripping.

Coefficients of Friction — The coefficient of friction between leather and cast iron varies greatly according to the condition of the cast iron surface and the state of lubrication. James Angelino in experiments made with a piece of old clutch leather found the coefficient of friction to vary from f = o.i$ to f =

FIG. 3.— COMPOSITION OF CONE CLUTCH FORCES. Kent gives the coefficient of leather on greasy metals as 0.23 In the calculations it, therefore, is the best plan to figure on a coefficient of friction of 0.2, since the leather is generally boiled in tallow or soaked in castor oil prior to being applied to the clutch, and hence is always somewhat greasy. Cast iron on cast iron cone clutches, lubricated, have been used to some extent, and in their case the friction coefficient is com- paratively low, not exceeding 0.07, depending upon the nature of the lubricant. Asbestos fabric is also used to some extent as a facing for clutch cones. It possesses the advantage that it is not affected by high temperatures. The coefficient of friction is also somewhat greater than that of leather.

16 FRICTION CLUTCHES.

Diameter of Cone — It is desirable to make the diameter of the cone small, for the following reason: The sliding gears or jaw clutches of the change gear practically never run at equal peripheral speeds just previous to being meshed, but from the moment they become meshed they must run at the same speed. This means that at the moment of engagement one of the connected parts must suddenly change its speed, and this results in a clash or hammer blow at the point of engage- ment. Now, one of these parts is mechanically or positively connected to the driving wheels of the car, and therefore can- not quickly change its speed. The other consists of the clutch cone and of a train of gears, and the resistance to a change in the speed of .these parts is proportional to the sum of their polar inertias, of which the inertia of the clutch cone is by far the greatest. The inertia of a revolving body is proportional to its weight, and to the square of its radius of gyration, which latter, in the case of a clutch cone, varies substantially as the mean outside diameter. Hence, the force of the clash increases and decreases substantially as the square of the mean outside diameter or radius of the cone. On the other hand, the radius must be made large enough to keep down the unit normal pressure (which determines the wear of the clutch facing) and the spring pressure required to transmit the torque of the motor, since a clutch with a very stiff spring is "harsh" and difficult to operate. As a general rule, the flywheel would be designed first and the clutch made of a corresponding diameter.

Unit Normal Pressure — In conical clutches lined with leather, asbestos fabric or similar material the unit normal pressure generally ranges around 12 pounds per square inch. However, in some of the largest cone clutches it is nearly 20 pounds per square inch, and yet satisfactory service is ob- tained. Cone clutches are used mainly for the smaller engines and multiple disc clutches for larger powers, and if the former are used on engines of 60 horse power and over it is necessary to employ large unit pressures, because the available diameter is not much greater than is used in clutches for smaller powers, and unduly wide clutch faces are also out of the question. While the clutches work satisfactorily under the higher pres- sures, it is natural to expect them to wear out quicker, and wherever the space available permits it is best to keep the unit pressure down to 12 pounds. In metal-to-metal cone clutches the unit normal pressure must be several times that used with leather faced clutches in order to transmit the same power.

FRICTION CLUTCHES

17

The foregoing figures are based upon the normal pressure re- quired to hold the clutch from slipping when fully engaged. The actual normal pressures are some- what greater because the clutch spring must be made stronger than required to produce this normal pressure under conditions of rest. Suppose that the normal pressure* N is just insufficient to produce the necessary driving torque and the clutch slips. The normal pressure can only be in- creased by forcing the cone further into the flywheel, and this necessitates overcoming the resistance to .motion of the leather over the cast iron sur- face in a direction normal to that of slippage.

Referring to Fig. 4, let N repre- sent the effective normal pressure between the clutch friction surfaces, and P the spring pressure necessary to produce this normal pressure. Let F represent the frictional force in the direction of a generatrix of the cone; Pi, its component parallel to the clutch axis, and 0 the so-called friction angle (tan 0 = /), then

P = N sin a,

F = N tan <i> and

Pi —F cos a== N tan <j> cos <*.

The total spring pressure necessary to cause the clutch to engage firmly without slipping is

P2 = P + Pi = N (sin a + cos a tan 0) (3)

In applying this equation it is permissible to use for tan <i> a con- siderably smaller value than the normal coefficient of friction between leather and cast iron. This is due to the fact that when one body moves frictionally over another in a given direction, it requires but an insignificant effort to start it moving at right angles to its original direction of motion. That is, the coeffi- cient of friction encountered in any given direction is virtually reduced by motion in a direction at right angles thereto. An illustration of this principle is furnished by the fact that when a

FIG. 4. — COMPOSITION OF

CLUTCH FORCES DURING

ENGAGEMENT.

18 FRICTION CLUTCHES.

mechanic wants to force a tight fitting collar over a shaft he will twist it angularly back and forth on the shaft, whereby the effort required to move it in an axial direction is greatly reduced. We may assume that the coefficient of friction in this case is one-fourth of its normal value, or 0.05. Hence, the general equation for the spring force required to engage a leather-faced cone clutch becomes P = N (sin a + 0.05 cos a) ............ *. ..................... (4)

The frictional force at the mean circumference of the cone is

TX 12

— pounds. ^m

The area of the cone face is

2 ir rm w,

and since there is to be a normal pressure of 12 pounds per square inch, the total normal pressure is

12 X 2 7T rm TV = 24 7T rm "W.

This multiplied by the friction coefficient 0.2 gives the total frictional force —

O.2 X 24 7T rm IV = 4.8 7T rm tV.

Equating this to the expression for the frictional force found above, we have

TX 12

— - - = 4.8 TT rm iv,

rm

and

•"-srr^n ............................................. (5)

from which equation the necessary width of face may be found.

It is also possible to derive an equation for the necessary spring pressure in terms of the fundamental clutch data. The normal pressure

N = 12 X 2 if rm TV = 24 TT rm w.

Substituting the value of w, found above,

and substituting this value of N in equation (4) we have

xr *r»

p _ - (sz-n a _f. 0 .05 cos a) ............................... (6)

' m

In order to facilitate the determination of the necessary face width and spring pressure, according to equations (5) and (6), Chart I has been drawn. From this chart can be found the low speed torque of four and six cylinder motors of any cyl- inder dimensions within the usual range of automobile practice, as well as the necessary width of clutch face and of the clutch

FRICTION CLUTCHES.

19

Spring Pressure For Six Cylinder JYIofor

Lbs. 300 450 JOO 750 900 W50

f"~ Spring Pressure for Four Cylinder Motor

Lbs. 200 JOO 400 500 600 700

Face 5 Width For 4" Six Cylinders

CHART i.— GIVING Low SPEED TORQUE OF FOUR AND Six CYLINDER

MOTORS AND WIDTH OF FACE AND SPRING PRESSURE

REQUIRED FOR A LEATHER-FACED CONE CLUTCH

TO TRANSMIT THIS TORQUE.

20 FRICTION CLUTCHES.

spring pressure required with different mean radii of clutch and angles of cone. The method of using the chart is indicated in diagram.

Constructional Details — Since the inertia of the clutch must be as small as possible, the clutch cone is generally cast of aluminum, though of late pressed steel clutches have come into quite extensive use, mainly abroad. In the case of aluminum cones the rim is generally made of a mean thickness of one-quarter inch, tapering from the edges toward the joint with the web, which latter should preferably be at the middle of the rim. In order to obtain the necessary strength in the web with the least amount of material the latter, instead of being made radial, is inclined considerably toward its axis, so the material will work partly under compression. The dimensions of the web or spokes are largely a matter of foundry limitations. For the smaller powers a plain web is used, tapering from about three-sixteenths inch near the rim to one-quarter inch where it joins to the steel centre, which is lightened by large holes being formed in it. Some designers, however, prefer to leave the rim solid, as it keeps out dust. When spokes are used they are often of cross- shaped section or ribbed, so as to provide additional lateral strength in the cone and also to support the rim more rigidly.

Clutch leather is generally treated before being applied to the clutch by being either boiled in tallow or soaked in castor oil, the excess oil or grease being removed by passing the leather through rolls. The leather must be cut to form a sector of an annular ring of an inside radius o-a and an outside radius o-b (Fig. 3). The* length of the inner edge of the annular sector must evidently be 2 TT r, Now, the radius

r

o-a = — : — ' stna.

and the circumference of a circle of radius o-a therefore is

2 TT r

sin a

Hence the angle 0 to which the leather should be cut can bt found from the proportion

2 irr

— — : 360 degrees = 2 v r : <f>

<f> = sin a X 360 degrees.

Therefore, in laying out the pattern of the leather (Fig. 5), strike two concentric circles of radii

and o.fr -- — : — 4- -uj

stria '

FRICTION CLUTCHES.

21

where w is the width of the face of the clutch. Then from the annular ring thus formed cut out a sector subtending an angle sin a X 360 degrees at the centre.

Some designers form a small radial flange on the edge of the rim at its bigger end which will retain the facing, and thus take some of the stress off the retaining means and off the facing itself. When leather facing is used it is retained by means of copper rivets whose heads are countersunk beneath the surface of the leather and whose ends on the inside of the clutch rim are hammered over. Usually two rows of one-eighth inch rivets are used, spaced about an inch apart. After the leather is riveted

FIG. 5.— PATTERN FOR CLUTCH LEATHER.

to the cone it is accurately turned off in a lathe. An improved method of holding the leather in place consists in the use of six or eight T bolts, and the provision of depressions in the rim of the cone parallel with generatrices of the latter, for the re- ception of the heads of the T bolts. This method of securing the facing, which is particularly applicable to asbestos fabric, (which does not lend itself well to riveting) is illustrated in Fig. 6. Provisions for Smooth Engagement — Cone clutches have a tendency to grip with a jerk, especially in case the car is oper-

22

FRICTION CLUTCHES.

FIG. 6. — CLUTCH LEATHER FASTENED BY T-BOLTS.

ated by a novice driver or the clutch operating linkage is such that the driver must exert a very strong pressure on the clutch pedal. In order to overcome this tendency, which is detrimental to the whole car, various devices are resorted to, all based on the principle that a portion of one of the engaging surfaces is raised by spring force above its normal height, and thus that portion alone first contacts with the opposing surface. The plan most commonly followed consists (Fig. 7) in turning a shallow circumferential groove on the outside of the aluminum cone near its large end, in which are placed a number of equally spaced flat steel springs which are fastened to the cone by one rivet each, or to a screw secured in the rim of the cone. These steel springs are of such form that they slightly lift the leather when the clutch is disengaged, so that certain portions of the leather come in contact with the flywheel rim first. These "auxiliary" springs are fully extended when the clutch surfaces first engage each other, and the pressure of contact therefore starts from nothing.

A similar device, comprising coiled instead of flat springs, is illustrated in Fig. 8. It consists of a small shell cast integral with or. riveted to the clutch rim from the inside, which contains a coiled spring and a plunger pressed outward thereby. The head

FIG. 7. — FLAT SPRING UNDEP CLUTCH FACING.

FRICTION CLUTCHES.

23

of the plunger, which presses against the clutch facing from underneath, may be either fillister shaped or in the form of a crossbar extending underneath the leather the entire width of the clutch face.

Where either a male or female cone of steel is used it is pos- sible to cut slits in it length- wise and circumferentially, as shown in Fig. 9, or at an angle to the edge, and then bend the flaps so formed slightly outward or inward, as the case may be. This prac- tice is or has been followed by Renault, Cadillac, Pullman and others.

FIG. 8. — SPRING PLUNGER UNDER CLUTCH FACING.

Cork inserts are used with" leather-faced cone cluches by a number of manufacturers, mainly with the object of making the engagement more gradual. The properties of these corks will be discussed in connection with plate clutches, in which they are more extensively used. Corks used in leather-faced cone clutches vary in diameter from five-eighths to one inch and cover from 5 to 30 per cent, of- the surface of the cone. When the area presented by the corks does not exceed 10 per cent, of the total frictional area, they do not materially affect the co- efficient of friction, but some advan- tage is gained in this respect when from 20 to 30 per cent, of the surface is made up by the corks. The fric- tion is then somewhat greater than that between leather and cast iron, and consequently the spring pressure can be reduced.

Multiple Springs— A few makers use three clutch springs placed at equal angular distances and about midway between the clutch shaft and the rim. One advantage of this arrangement is that the clutch is more easily adjusted, owing to the greater accessibility of the adjusting means. A typical design of this kind is shown in Fig. 10. A three armed spider is placed on the tailshaft just behind the web of the flywheel, whose arms carry studs or spring bolts extending backward parallel with

FIG. 9. — SLOTTED CLUTCH FEMALE CONE.

FRICTION CLUTCHES.

the tail shaft, through holes in the web of the clutch cone. The por- tions of the three spring bolts extending through the clutch cone are sur- rounded by coiled springs, whose rearward ends bear against washers supported by adjusting nuts. The spring thrust is taken up on a ball thrust bearing carried on the tail shaft. Construc- tions similar to the one here shown are used by several English manu- facturers.

Clutch Centre — The clutch may be regarded as composed of three main parts, viz., the cone with its web or spokes, the supporting bearing, and a spring housing or hollow shaft by which the power is transmitted to the change gear. Generally these three parts are made separate, though sometimes the cone is formed integral with the bearing. The clutch bearing is oper- ating only when the clutch is disengaged, and evidently carries very little load. It therefore may be of relatively small diameter and free fitting. The bearing is practically always a plain one, and generally the non-fluid oil in the clutch spring housing is depended upon for its lubrication, it being drilled with several large oil holes and cut with deep oil grooves, but some makers in addition provide a pressure grease cup on the outside of the clutch which can be screwed down at intervals, the grease being forced through a drill hole directly to the bearing surface.

The clutch spring generally surrounds the bearing, its forward end resting against a flange thereon and its rear end against a ball thrust bearing on the end of the tailshaft or on a cap screw screwed into the end of that shaft. This thrust bearing works only when the clutch is disengaged, whereas when it is engaged both ends of the spring press against parts rotating in unison and incapable of moving further apart. In other words, the

FIG. io.— MULTIPLE SPRING CLUTCH.

FRICTION CLUTCHES. 25

spring pressure is then self-contained. This is contrary to con- ditions in the earlier cone clutches, in which the clutch spring took purchase on a shoulder on the transmission shaft, thus creating end thrust in both the transmission shaft and the crank- shaft

It is quite desirable to keep down the length of the tailshaft, so the change gear box may be located underneath the floor boards of the driver's seat, and enough space should be allowed between the rear end of the tailshaft and the forward end of the transmission driving shaft, so the clutch can be removed from the car without removing either the engine or the gear box. If the web of the clutch cone is inclined backward, for the purpose of increasing its strength, the flange for connecting it to the clutch centre usually comes at a considerable distance from the fly- wheel flange, and it is therefore advantageous to make the bear- ing of a form similar to a cake mold, as shown in Fig. n, so its forward end will come within a short distance of the flywheel, making allowance only for the wear of the clutch leather.

In designing the clutch centre, attention must be paid to the exigencies of assembling. The clutch spring housing covers the spring and extends beyond the end of the tail shaft, hence the spring must be put in place and adjusted before the housing is put in place. Some makers bolt the web of the cone and the flange of the bearing together by, say, three bolts, and pass three intermediate bolts through the web of the cone and the flanges of both the bearing and the clutch housing. This admits of assem- bling the cone with the bearing, then placing them on the engine tail shaft, putting the clutch spring and its retaining nut in place, and finally bolting the clutch spring housing to the cone and bearing. Others place the web of the cone between the flange of the bearing and the flange of the clutch spring housing, and pass all of the retaining bolts through all three connected parts. This makes it necessary to assemble these parts after the clutch spring is in place, which, of course, can be done only with a spoked cone. Still other makers connect the clutch housing with the clutch bearing by means of radial bolts or set screws.

Spring Thrust Bearing— When the clutch is disengaged and at rest its spring bears with one end against a rotating part (tailshaft spring rest) and with its other against a stationary part, and to prevent undue wear and friction under these condi- tions the spring usually exerts its pressure through a ball thrust bearing at the rear end. In fact, if no ball thrust bearing were provided, the friction between the spring and its support would

FRICTION CLUTCHES.

FRICTION CLUTCHES. 27

likely be great enough to cause the clutch cone to keep on spin- ning. The ball thrust bearing may be passed over the end of the tailshaft and held in place by means of a castellated nut, or this bearing may be supported by means of a cap screw screwed into the end of the tail shaft. With either arrangement the spring pressure may be adjusted; with the latter it can be ad- justed through a considerable range, and, besides, the tail shaft will be shorter, so the change gear can be brought closer to the engine. In any case it is necessary to provide a lock for the adjustment, and with a cap screw carrying the ball thrust bearing this lock usually assumes the form depicted in. Fig. n. The cap screw is drilled through its centre and slightly tapered out and split at its outwardly threaded end, to receive a small screw, with a correspondingly tapered head. By means of this inner screw and its nut the split end of the cap screw can be expanded and the screw thus securely locked in place.

Clutch Springs— The springs which hold the clutch in engage- ment are generally helical or coiled springs made of either round or square steel wire. Formulae for the safe load and the deflec- tion of round steel wire coiled springs were given in Vol. I in the chapter on Valves and Valve Gear. The corresponding formulas for square steel wire springs are

= 0.471^- n P D3

D = mean diameter of coil.

W = maximum safe load in pounds.

r = compression of spring.

d = side of cross section of wire.

n = number of coils in spring.

5" = maximum safe fibre stress of material.

£ = torsional modulus of elasticity.

P = load in pounds.

Occasionally, in order to save space in a longitudinal direction, so-called volute springs, made of flat metal, as shown in Fig. 12, are used.

Pressed Steel Cones — Pressed steel cones are very attractive to the designer, owing to their light weight and their low cost when made in large numbers. Some trouble is said to have been encountered with these cones owing to insufficient rigidity and consequent shattering, and it has been recommended to press* the cone with radial ribs to overcome this difficulty. One English

28 FRICTION CLUTCHES.

manufacturer makes the web of his pressed steel cone in the form of a zone of a sphere, evidently with the same object. Either J4 inch or 3/16 inch stock is used. A recent develop- ment in the line of clutches is a pressed steel clutch with a leather facing secured to the driving cone. This should reduce the moment of inertia of the driven cone in such a degree as. to eliminate all obj ection to the cone clutch on this score.

The several designs of clutches here shown are particularly simple. A great deal of ingenuity has been applied by designers in working out the details of clutch centres, and much variety is to be found in the designs extant. With cone clutches of the inverted type it is not easy to provide adjusting means for the clutch spring, and none is generally provided.

Shifting Collar — To disengage a cone clutch the driven cone must be withdrawn from the driving cone against the pressure of the clutch spring. This necessitates a sliding connection between the clutch pedal shaft, which usually extends across the vehicle frame directly above the clutch housing, and this housing. The latter is usually provided with a circumferential groove, in which is located a sliding collar. If both flanges of the groove are integral with the housing, the shifting collar, of course, has to be made in halves in order to get it into tht groove, the halves being bolted together. However, generally only one flange of the groove is integral, so the shifting collar can be slipped over the housing from one end.

When the clutch is withdrawn the entire pressure of the clutch spring is taken up on one face of the shifting collar, and to obviate the necessity of constant attention to the lubrication of this collar a ball thrust bearing is generally placed in the groove to one side of the shifting collar, so as to take the thrust of the spring. This, of course, necessitates the use of one removable flange, in order to get the ball thrust bearing into place.

A typical shifting collar design is shown in Fig. 14. The collar itself is made of brass and provided with two radial pins, with which engage the free ends of the forked clutch shifting lever. These lever ends are formed with oblong holes for the pins to pass through, to make allowance for the fact that they move in an arc of a circle, while the shifting collar is constrained to move in a straight line. A grease cup is usually screwed into either one or both of the shifting collar trunnions.

In some cases the shifting collar is made in the form of a cir- cular disc, and the forked shifting lever is made cam-shaped and bears against one face of the disc. Pressure has to be transmitted

FRICTION CLUTCHES

29

FIG. 12.— VOLUTE CLUTCH SPRING.

FIG. 13.— PRESSED STEEL CONE CLUTCH.

30 FRICTION CLUTCHES,

from the clutch pedal to the clutch housing in one direction only, and if the clutch shifter fork is held against the shifting collar by means of a spring no groove for the collar is necessary.

Clutch Brakes — By lightening the cone, and especially by reducing its diameter, it has been endeavored to reduce the shocks due to clashing of the gears, but there have also been efforts in other directions to insure the possibility of smooth meshing. The clashing, of course, is due to unequal pitch velocities of the two

FIG. 14. — CLUTCH SHIFTING COLLAR.

gears meshed. If the speed of one of the gears can be increased or reduced previous to meshing until it corresponds to that of the other, then the gears can be meshed without shock or jar.

Suppose that a car is ascending a hill and it becomes necessary to change to a lower gear. It is evident that if the second gear, say, is disengaged, and an attempt is made to immediately engage the first gear, the driven wheel of the latter will run too fast or the driving pinion too slow to permit of easy meshing:. The

FRICTION CLUTCHES.

31

driver has no control over the driven gears, except through the use of the car brake, and it would be inadvisable to apply that while ascending a hill. However, as the car is on an up-grade, its speed and that of the driven gear decrease rapidly when the motor is disconnected, and the gears can be readily meshed after a short interval of time. In changing down on the level the driver speeds up the pinion of the first gear by allowing the clutch to partially engage momentarily. If the driver is skilled in han- dling the clutch and gears, he will be able to shift the gears when the two to be engaged are running at about' the same pitch line velocity.

FIG. 15. — CLUTCH BRAKE.

Thus in changing down the gears automatically approach the condition of equal pitch line velocity, or the condition of easy mesh. Not so in changing up. The driven gear of the pair to be meshed is now running too slowly and the driving gear too fast. The latter can only be reduced to the proper speed by applying a brake to the clutch. Many of the larger cars are now equipped with such clutch brakes, which act automatically when the driver completely pulls out the clutch. One design of such a brake is shown in Fig. 15. The clutch housing is formed with a flange B, against which bears a fibre block A, carried on an arm on the clutch pedal shaft, when the clutch pedal is fully depressed. Another type of clutch brake is illustrated in Fig. 16. This is a

32

FRICTION CLUTCHES.

clutch of the plate type, and the power is transmitted from the clutch shaft through a pinion and an internal gear, which latter is formed integral with a shaft coupling made in halves. This coupling is provided with an annular friction ring B, which when the clutch is fully withdrawn presses against a corresponding disc A, secured to the clutch shifting ring, which latter, of course, does not rotate.

There is quite a variety of designs of clutch brakes, the under- lying principle of all of them being that a part rotating with the clutch is brought into contact with a non-rotary part when the clutch pedal is fully depressed, and the friction engendered between the two parts causes the speed of the clutch to be reduced.

1

FIG. 16. — Disc CLUTCH BRAKE.

Multiple Disc Clutches— Disc and plate clutches are based on the same principle, but constitute in a sense opposite extremes in design. A disc clutch consists of two sets of annular .discs, one set of driving discs and one set of driven discs. These are placed together in alternate order, each driving disc being located between two driven discs. As generally used on automobiles, the driving discs are provided with key slots on their outer circum- ference into which fit keys on the inside of a drum shaped hous- ing secured to the flywheel, and the driven discs are provided with lugs or key slots on their inner circumference, which place them in driving connection with a drum secured upon the driven shaft. Generally there is one more driving disc than there are

FRICTION CLUTCHES.

33

driven discs, so that the two end discs are of the same kind. The drum carrying the driven discs has a radial flange at one end which forms a stop for the discs in respect to axial motion, and against the disc at the other end presses a compressing spider or presser, against which the clutch spring exerts its pressure.

Types of Disc Clutches — Multiple disc clutches operating in oil are of three different types of design, the differences depend- ing upon the manner in which the pressure of the clutch spring is transmitted to the flange or back stop of the discs on the clutch drum. Some of these clutches employ three clutch springs, the

FIG. 17.— MULTIPLE SPRING TYPE OF Disc CLUTCH.

same as some cone clutches, and a design of this type is shown in the sketch Fig. 17. An outer drum A is secured to the flywheel and is provided with a number of equally spaced keys on its inner circumference. With these keys engage the driving discs, which are shown sectioned. Between adjacent driving discs are located the driven discs, shown in black. The latter are carried on the inner drum B, which is provided with keyways for the lugs formed on the inner circumference of the driven discs. From the web of the clutch drum B extend three lateral spring bolts which carry the clutch springs C pressing against the disc compressing spider D. Drum B is keyed to clutch shaft E, which is connected with the driving shaft of the change gear, and the

34

FRICTION CLUTCHES.

disc compressing spider D is provided with a hub surround- ing shaft E and a groove for the clutch releasing collar, or merely a flange.

A multiple disc clutch with a single clutch spring surrounding the clutch shaft is illustrated in Fig. 18. The arrangement of the outer drum, driving and driven discs and inner drum is the same as in Fig. 17. In this case one end of the clutch spring bears against an inward flange on the hub of the disc compres- sing spider D, and the other against a collar on the clutch shaft E. The latter has the inner drum B securely keyed to it and held against endwise motion by a nut. Hence, the pressure of

FIG. 18. — SPRING PRESSURE TRANSMITTED THROUGH SHAFT.

the clutch spring is transmitted to the forward end plate or stop P of the discs through the clutch shaft E and the clutch drum B.

In Fig. 19 is shown a design of multiple disc clutch in which the spring pressure is transmitted to the stop P of the discs through the clutch housing A. The most forward disc bears against a stop ring P secured to the flywheel and against the rear- most disc presses the compression plate D in the usual way. This disc or spider D is acted upon by the coil spring C which rests against the flange of the casing A. Figs. 17, 18 and 19 are sketches only, not showing all of the necessary details of these clutches.

The spring forces the separate ftiscs together and causes the

FRICTION CLUTCHES.

35

driven discs to rotate in unison with the driving discs, provided the resistance to the motion of the driven discs is not greater than the adherence between the driving and driven discs. It will read- ily be seen that the pressure betwen any two discs is equal to the pressure of the spring, and the adherence or resistance to slipping at any contact surface is equal to the product of the spring pressure by the coefficient of friction. But if there is slip- page on one contact surface there must be slippage on all of them, and since the pressure on any contact surface is the same as on any other, the total resistance to slippage is equal to the prod- uct of the resistance to slippage at one surface by the num- ber of contact surfaces, which latter is equal to one less than

FIG. 19. — SPRING PRESSURE TRANSMITTED THROUGH CASE.

the number of discs. In a multiple disc clutch the frictional sur- face can be made much greater than in a cone clutch, and the frictional force per unit surface can be made smaller.

Calculation of Disc Clutches — In Fig. 20 is shown one disc of a multiple disc or plate clutch. In this figure dr is the width of an extremely narrow annular ring of radius r. Sup- pose that the unit pressure on the surface of this disc is p pounds per square inch. The area of the annular ring of width dr is

A = 2 if r dr and the normal pressure on it is

36

FRICTION CLUTCHES.

N = 2 T r dr p. This causes a frictional force

2 TT r dr p f, where / is the coefficient of friction, and a torque

r TT r2 dr p f

12 6

Now, in order to find the torque which the friction over the entire surface of the disc will produce we have to integrate the above expression between the limits r0 (outside radius) and n (inside radius)

Sdrpf

FIG. 20.

•*

= — p f (n3 — n*) pounds-feet, 18

(9)

n 6

Equation (9) is useful in the case of clutches whose discs have a very small inside radius. In the original type of this clutch the discs were often mounted directly upon the driven shaft, and the inside radius of the clutch disc was less than one-quarter the outside radius. However, in modern automobile clutches the inside radius is generally more than three-fourths the outside radius, and the so-called discs are really in the form of narrow annular rings. There are two main reasons for making the ele-

FRICTION CLUTCHES. 37

ments of the clutch ring-shaped rather than disc-shaped. The first is that the wear of the disc increases with the distance from the centre of rotation, owing to the fact that the speed of slippage increases with the distance from the axis. Hence, if there is a great proportional difference between the outside and inside radii, the rates of wear near the inner and outer edges will be greatly different. The result will be that as the outer portion of the disc becomes thinner than the inner portion, the pressure over its sur- face will become unevenly distributed, the unit pressure being greater near the inner edge than near the outer edge, and conse- quently the clutch will transmit less power than originally with the same spring pressure.

The other reason is that the resistance to lateral motion of the discs depends directly upon the pressure between the driven discs and their keys or keyway walls, which is less the greater the inner radius of the discs. When the clutch is disengaged the discs are not positively pulled apart, but are supposed to be either jarred apart by the vibration or to be forced apart by auxiliary springs, and especially in the former case is it desirable that the resistance to their lateral motion be as little as possible, as there is then less danger of dragging.

In the case of clutch discs or rings whose inner radius is more than two-thirds of the outer radius it is permissible to consider the engaging pressure (and hence the frictional force) concen- trated at a distance from the axis of rotation equal to the arith- metical mean between the outer and inner radii (rm ) . The fric- tional force at any contact surface then is P f, the aggregate fric- tion force (n — i) P f, and the moment of the frictional force or torque.

In any given problem of design the torque to be transmitted is a fixed quantity, but the limiting torque of the clutch is the product of four variables, viz., the mean radius of the discs, the number of contact surfaces, the spring pressure and the friction coefficient. Since these factors are independently variable, it is not surprising that practice in disc clutch design is not in the least uniform. The tendency is rather toward small mean radii and a very considerable number of discs, since the inertia increases as the square of the radius and directly as the number of discs, whereas the capacity of the clutch increases directly with both the radius and the number of discs. The coefficient of fric-

38 FRICTION CLUTCHES.

tion, of course, can be changed only by changing the material of the discs or the lubricant.

Material of Discs — In the type of disc clutch which has been used the longest in automobile practice, both sets of discs are metallic and run in oil. The discs are generally made from saw steel, about 3s inch thick, stamped rings of any desired diameter, with driving lugs or key slots, as desired, being fur- nished by several saw steel manufacturers. Some manufacturers believe that steel on bronze gives better wear and make one set of the discs of the latter material. Saw steel is a. very -suitable material, being hardened and more uniform in thickness than ordinary sheet steel. Sheet copper is also used together with sheet steel. Whatever material is used, the greatest care must be exercised to get the thickness as nearly uniform as possible and to give the surfaces a smooth finish.

Laws of Friction — The coefficient of friction between metals with lubrication varies widely according to conditions. Some of the laws of friction which have a bearing on the value of the coefficient of friction in disc clutches may be stated as follows: The coefficient of friction between two metallic surfaces separated by a film of lubricant is much greater when the surfaces are at rest relative to each other than when there is sliding motion between them. The friction does not depend so much upon the material of the discs as upon the lubricant. When the discs are stationary the coefficient of friction increases with the specific pressure. On the contrary, when there is sliding motion between the surfaces the coefficient of friction decreases as the specific pressure increases (up to a certain limit which, however, is far beyond the pressure used in disc clutches). The coefficient of friction also varies with the speed; it seems to be a minimum at loo to 150 feet per minute, increasing as the speed is increased or diminished, and approaching the static friction coefficient at very low speeds.

From the above it will be seen that it is difficult to assign a definite value to the coefficient of friction f for use in the calcu- lation of friction clutches. However, the author believes it to be on the safe side to use a coefficient ^ = 0.04 for steel on steel, phosphor bronze or copper with lubrication.

Disc Clutch Data— Denoting the mean radius of the discs by rm and the number of f rictional surfaces by na , the equation for the limiting torque of a disc clutch may be written

FRICTION CLUTCHES. 39

Now, even if the material of the discs is settled, so that / is a fixed quantity, there remain three independent variables, and the desired torque, therefore, can be obtained in many different ways. In this connection it is to be remembered that if we increase the mean radius rm we increase the inertia of the clutch, even if we correspondingly decrease the number of discs so as to retain the same limiting torque. On the other hand, if we increase either the number of discs or the spring pressure we increase the work which must be done by the operator in disen- gaging the clutch, because the spring must be compressed an amount proportional to the number of discs, in order that there may be sufficient clearance between adjacent discs, and the work done in compressing the spring is measured by the product of the clutch spring pressure by the distance of the compres- sion of the spring during the process of declutching. The foot has only a small range of comfortable motion, and the pressure which can be exerted by it is also limited. It is evident that the product Pna is a measure of the work to be done in dis- engaging a clutch, and it has been found that this product should not exceed 12,000 if clutch operation is not to be irksome. The friction force per unit of contact surface varies from 0.6 pound per square inch to 2 pounds, the average value being i pound. In clutches of this type (metal-to-metal-in-oil) the average ratio of the inside to the outside radius is five-sixths.

If we made Pn, = 12,000 for all clutches, then the small clutch would be as hard to operate as a large one, which is not exactly desirable. Besides, the mean radius rm would increase in direct proportion to the torque to be transmitted; it should increase with the torque, but not in direct proportion. It may well increase as the square root of the torque, and the following equation gives a good value:

3-4

We may therefore recapitulate the rules for multiple disc clutch design as follows : n

- =s in average practice. fo

Friction force = I pound per square inch. Coefficient of friction / = o.O4. The area of one friction surface is

TT (r02 — n2) square inches. and the frictional force between adjacent discs, expressed in

40 FRICTION CLUTCHES.

pounds, is the same. The total frictional force at the mean radius of the discs is

TX i2?

hence the number of friction surfaces required is 12 T rm 12 T

ir(r02 — n2) — TT rm (r02 — and the number of discs required

The spring force required is

P= -v up ^ -i (I2)

Both of the above equations can be materially simplified if the ratio of the inner to the outer radius is fixed.

For n= P =

n _ 21.2 T

^ ~™

n — A ,o-f

r\

•£=»

If we assume an inner radius equal to five-sixths the outer radius and substitute in the equations the value of the torque of a -four cylinder, 4x5 inch motor we find that the mean radius of the discs should be 3.4 inches, the outer radius 3.71 inches — say 3.75 inches — and the inner radius 3.12 inches — say 3^ inches. The number of discs figures out to 35 and the spring pressure to 337 pounds. An uneven number of discs is generally employed.

When the spring exerts its pressure through the clutch shaft or through spring bolts secured into the web or spokes of the inner drum, it is well to have one more driven disc, whereas when the spring exerts its pressure through the clutch housing it is best to have one more driving disc. In either of these cases if the

FRICTION CLUTCHES.

j^ 4" 4±" 5" ~3p

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7"orque in J^ou nets- feet

CHART II.— GIVING NUMBER OF Discs AND SPRING PRESSURE RE- QUIRED IN MULTIPLE-DlSC-IN-OlL CLUTCHES.

42

FRICTION CLUTCHES.

clutch is slipping there will be no relative rotary motion between the two parts against which the clutch spring bears, hence no ball thrust bearing will be required to take up the thrust of the spring.

Number of discs and spring pressures required in metal-to- metal multiple disc clutches can be readily found from Chart II, after the torque of the motor has been obtained from Chart I. Chart II is based on a unit frictional force of i pound per square inch and a coefficient of friction of 0.04. If a very light clutch is desired the number of discs found from the chart can be re- duced, and the spring pressure found increased in proportion.

Methods of Releasing Discs— In order to insure positive separation of the discs when the spring pressure is removed, and thus prevent dragging of the clutch, it is necessary to provide

FIG. 21. — METHODS OF SEPARATING Discs.

alternate discs with tongues sprung to one side, as shown in Fig. 21 at A, or some similar means. Two such tongues on each driving disc, one opposite the other, are sufficient. An alternate method of insuring positive separation is illustrated in the same figure at B, and consists in providing the driving discs with radial lugs on the outside circumference into which are riveted buttons whose heads are slightly thicker than the driven discs, so that the lugs are slightly sprung when the discs are forced to- gether by the clutch spring. The driving discs may be provided with four lugs, at quarters, and rivets inserted into two of these lugs, located oppositely. The discs may then be assembled in such a manner that the riveted lugs of adjacent driving discs are at quarters. Where separating springs of this kind are used, the clutch spring must be made sufficiently strong to overcome

FRICTION CLUTCHES.

43

the force of these springs and still give enough frictional force between the discs.

Constructional Details—Multiple disc clutches, the same as other types, are generally combined with the flywheel, but occa- sionally they are enclosed in a special compartment of the change gear case, which can be done without difficulty, since these clutches can be made of a relatively small diameter. When thus enclosed in the gear box or when used in a unit power plant, there is no need to specially enclose the clutch. But in other cases an oil-tight housing must be provided. This housing is sometimes made of one-eighth inch pressed steel in a single piece, with a radial flange at its open end for bolting to the flywheel web and a hub portion either formed integral or riveted

FIG. 22. — METHODS OF DRIVING Discs.

to it which takes the adjusting bushing for the spring if the spring pressure is transmitted through the housing, and forms an oil-tight joint with the hub of the disc compressing spider. This housing may also be made of two castings — a cylindrical shell and an end plate. Some designers even provide a stuffing box in the hub of the clutch housing to insure oil tightness.

If the clutch has no special housing it may be driven from the flywheel by radially extending driving pins secured into the web of the latter (A, Fig. 22). If a housing is used the driving is done either through keys riveted to the cylindrical shell (B, Fig. 22), or through bolts which hold both the shell and the end plate to the flywheel (C, Fig. 22). The key slots on the outside of the

44

FRICTION CLUTCHES.

driving discs are cut either in the full ring, or the rings are formed with driving lugs which have key slots or driving pin holes cut in them. The latter form of construction leads to a saving in weight, but necessitates a somewhat more expensive die for stamping out the discs. In any case, there must be a liberal clearance between the inner surface of the driving keys and the outer edge of the driven discs and between the inner edge of the driving discs and the surface of the inner drum so there will be no dragging owing to contact at these surfaces after slight wear.

Practice as to the number of driving pins or keys and driving lugs on the driven discs varies greatly. Some designers pro- vide as many as ten or twelve large size keys, which seems to be more than necessary. The number and size of keys do nol affect the freedom of lateral motion of the discs, but, of course affect the wear of keys and key slots, but clutches with only three one-half inch driving pins with an aggre- gate maximum pressure of three hundred pounds on them are known to give good results.

It is generally considered that one-hundredth of an inch is the minimum clear- ance between discs which will insure freedom from dragging, and in the de- sign of the housing and the inner drum allowance must be made for end

motion of at least — inch. In practice the allowance made

TOO

varies from i/ioo to 1/64 inch per friction surface. However, one well known manufacturer of multiple disc clutches allow? only from 1/125 to 1/175 inch.

The inner drum or the shaft to which it is secured is usually supported upon or in a radial ball bearing. The reason for the

FIG. 23. — INNER DRUM.

FRICTION CLUTCHES.

45

use of a ball bearing at this point is that the bearing, if plain, would be hard to lubricate effectively except through the engine tailshaft, and, besides, the friction of this bearing tends to pro- duce dragging, and the tendency to drag is already the weak point of the multiple disc clutch. Usually the radial bearing is carried upon a short tailshaft, and its outer race is forced into a counterbore in the drum, but in some constructions the bearing is carried upon the end of the clutch shaft and its outer race rests in the bore of the flywheel web. The drum (Fig. 23) is preferably made of a steel or malleable iron casting and milled with from four to twelve key slots in which engage the key lugs formed on the driven discs. The end plate which forms the stop for the discs is made separate from the drum and is secured to its rim by machine screws, or else passed over the drum against a small flange turned thereon.

The rim of the drum should be made sufficiently longer than the combined thickness of the discs to allow the latter to separate completely without passing beyond the rear edge of the rim.

FIG. 24. — SKELETON FORM INNER DRUM AND PRESSER.

46

FRICTION CLUTCHES.

Owing to the fact that the rirn of the compressing spider must move Over the rim of the inner drum for a considerable distance, while at the same time the web ot this spider must be quite close to the web of the inner drum, so the clutch spring will not extend too far to the rear of the clutch proper, this compression spider usually has a rather awkward form and is quite heavy. This difficulty can be overcome by making the inner drum in skeleton form, as shown in Fig. 24, cutting away its rim between those portions where the keyways are, and making the compres- sion spider spoked, the spokes entering between the lateral projec- tions of the inner drum rim. Besides reducing the weight of the driven part of the clutch, this construction allows of a more compact housing.

Hele-S-haw Clutch — A special type of multiple disc clutch which is extensively used both in this country and abroad is the Hele-Shaw, which consists of alternate discs of steel and phos- phor bronze with V-groove corrugations whose walls form an angle of 35 degrees. Only the walls of the V-grooves come in frictional contact, and the remaining parts of the discs merely serve to help radiate the heat engendered during slippage. Oil holes are drilled through the inner walls of the grooves near the peak, so the oil can enter and escape freely. It is obvious that in this clutch there is a sort of wedge action, the same as

r

FIG. 25.— HELE-SHAW CLUTCH.

FRICTION CLUTCHES.

47

FIG. 26.— CLUTCH SPRINGS INSIDE SHAFT.

in a cone clutch, and much less spring pressure is required to produce a certain amount of frictional force than with a flat disc clutch of the same number of discs and the same mean diameter. On the other hand, the discs have to be moved laterally consid- erably farther to obtain the proper clearance between them, and the number of discs that can be used is therefore more limited. The Hele-Shaw clutch shown in Fig. 25 is provided with a clutch brake, as are most large size disc and plate clutches.

Springs Inside of Shaft — Generally the clutch spring sur- rounds the clutch shaft, as shown in Figs. 17, 18 and 19, but some designers prefer to place it inside the clutch shaft or the engine tailshaft. Two such designs are shown in Fig. 26. In the first of these (Panhard) the spring acts through a plug and a key which extends through a long diametral slot in the shaft, against the clutch compressing spider. In the second (Hudson "33") the clutch spring is located inside the rear bearing of the crank- shaft and presses through a steel washer, a collar on the clutch shaft, a ball thrust bearing and a screw collar against the hub of the inside clutch drum. It should be explained that in this clutch the usual order of things is reversed, the inner drum being moved in an axial direction in order to disengage the clutch, thus serving as "presser."

48 FRICTION CLUTCHES.

Lubrication of Discs — The surfaces of the discs should be covered with lubricant when there is slippage, but it is also de- sirable that all or at least most of the lubricant be squeezed out from between them when the full pressure of the spring is ap- plied, since the clutch will hold the better the less lubricant there is on the discs. In order to insure these conditions, some manu- facturers provide the discs with radial slots extending over half their width, as shown in Fig. 27, through which the oil may escape when the discs are pressed to- gether.

Whereas the weak point

of the ordinary cone FIG. 27. — CLUTCH Discs

clutch is its great inertia, WITH OIL SLOTS.

that of the multiple disc-

in-oil clutch is its tendency to drag if the oil in the clutch housing is not suitable for the purpose, or if too much is intro- duced. Most makers recommend a mixture of machine oil or gas engine oil with kerosene. It is obvious that the thinner the lubricant the better the clutch will hold, while the more viscous the lubricant the more gradually it will pick up its load.

Dry Plate Clutches— In order to overcome the dragging evil the dry plate clutch was introduced. In this one set of plates

FIG. 28. — CORK INSERT CLUTCH.

FRICTION CLUTCHES. 49

is either faced with asbestos fabric on both sides or else pro- vided with cork inserts. Both of these materials when in contact with steel have a much greater friction coefficient than steel or bronze on steel. Cork on steel is claimed to have a friction co- efficient of about 0.34 when not lubricated. The cork, of course, is quite compressible. It is customary to make the plugs of such size that when free they project about & inch above the sur- face of the metal plate. Hence when the discs are forced together the contact is1 at first between metal and cork only, and owing to the compressibility of the cork the engagement is very smooth. However, when the full pressure of the spring is applied to the friction surfaces the corks are compressed flush with the plate surface, and one of the surfaces is then part metal and part cork. This, of course, will reduce the effective friction coefficient some- what, depending upon the relative area of the corks and of the metal and upon the compression of the cork at the moment metal to metal contact is established. As a rule, the cork covers from 25 to 50 per cent, of the total area of the discs, though there are extreme cases in which either more or less than the above range is covered.

The majority of disc clutches with cork inserts are of the three plate type, the middle plate containing the corks, though occa- sionally cork inserts are also used in multiple disc clutches. More- over, it is not necessary to run the cork insert clutches dry. Lubrication will reduce wear of the corks, but, of course, it will also reduce the friction coefficient. A typical cork insert clutch is illustrated in Fig. 28.

Asbestos fabric is also used for facing clutch discs. This is a fabric composed very largely of asbestos fibre and containing some brass wire and cotton, which latter give the necessary tenacity, while the asbestos is used on account of its good fric- tional qualities and its resistance to heat. The asbestos fabric is secured to the metal discs by means of rivets passing through the metal and asbestos on opposite sides of it. The frictional force in asbestos-faced disc clutches varies from less than one pound to about four pounds per square inch. With lower friction per unit surface the life of the clutch will, of course, be greater. The friction coefficient of asbestos fabric on steel seems to be approximately 0.3, and for ordinary purposes a normal pres- sure of 10 pounds per square inch will give satisfactory results. This gives a frictional force of three pounds per square inch, and

50 FRICTION CLUTCHES.

the formulas for number of discs and spring pressure required become

4T

and

/>== I0 TT (r02 — n2) ......................................... (14)

From the data at hand it seems that these same equations are applicable to cork insert clutches in which the spring acts on the discs directly and in which the corks cover1 from 25 to 50 per cent, of the total surface.

It may here be pointed out that a clutch for a vehicle in which the gear has to be changed frequently and the clutch therefore slipped a great deal should logically be designed with a somewhat lower unit friction force than a clutch for a high powered touring car, for instance, the speed of which can be largely controlled by the throttle. A lower unit frictional force will result in less wear and less heating.

The asbestos is generally secured to the driving discs, so the driven member may have the least possible inertia, but in one design the asbestos rings are free between the two sets of metal discs. The latter are made about % inch thick to get sufficient bearing surface on the keys ; if lighter stock is to be used the edges may be flanged to get additional driving area. Fig. 29 shows the Packard dry disc clutch which comprises six driving and five driven discs.

Three Plate Clutches — Another method of obviating the dragging tendency of disc clutches is to use only three discs or plates, without lubricant. These discs are made of cast iron and bronze, or of cast iron and steel. Since there are only two friction surfaces, for moderately high powers it is necessary to use rather large discs and to multiply the pressure of the clutch spring by levers or toggle mechanisms. Fig. 30 shows a typical design of this kind in which the spring pressure is multiplied by a toggle mechanism. One of the three discs is a driving disc, and the other two are driven discs. The driving disc is driven from the flywheel through keys riveted to the inside of the flywheel rim. One of the driven discs, the one nearest the flywheel, is secured to the clutch shaft and is provided with four sets of laterally extending lugs on which bell cranks are fulcrumed. One arm of these bell cranks connects through a link with a sliding sleeve on the clutch shaft on which the clutch spring acts. The other arm of the bell crank is provided with a set screw, the point of which presses against the rearmost driven disc. This latter disc is provided with driving lugs which enter between the

FRICTION CLUTCHES

52

FRICTION CLUTCHES.

lugs on the other disc serving as a fulcrum for the bell crank. The set screws permit of making adjustment for wear of the discs. Separation of the discs is effected by means of small coiled springs inserted into drill holes in one of the driven discs and pressing against the other driven disc.

The multiplying factor of the toggle mechanism attains the infinite as the toggles assume a radial position. In practice, of

FIG. 30.— THREE PLATE TOGGLE TYPE CLUTCH.

course, the set screws must be so adjusted that this cannot hap- pen, as the toggle links would pass by the radial position and the clutch would disengage again. If the links make a small angle 0 with a radial line then the multiplying factor is equal to co- tangent 0. This may be readily seen by reference to Fig. 31, in

FRICTION CLUTCHES.

53

which A B represents a link of the toggle. Let P be the pres- sure of the spring and N the radial pressure exerted on the bell crank arm. Now let point B be moved the slightest distance under the force of the spring P, so that the angle B A C (0) decreases to 0 — d 0. Now we have

C B = A B sin 0 A C = A C cos 0 When 0 decreases to 0 — d <t>, A B sin 0 decreases by A B cos 0 d 0 and A B cos 0 in- creases by A B sin 0 d <f>. But the product of the force into the distance through which it works represents the work done, and this must be the same at both points A and B. Hence,

PX A £ cos <j> d <j> =

N X ABsin<t>d$

and

*L p

cos 0

FIG. 31.

sin 0 "

A plate clutch in which the spring pressure is multiplied by double armed levers is illustrated in Fig. 32. In this clutch there are two driving discs, one being constituted by the web of the flywheel, and one driven disc. The free driving disc is driven from the flywheel through a stud bolt passing through the fly- wheel web and an annular flange bolted to the flywheel rim. The stud bolt is provided with a collar against which the short arm of the double armed lever takes purchase. This lever is fulcrumed on lugs cast integral with the free driving plate, and its long arm extends radially inward and is pressed against by the sliding sleeve which contains the clutch spring and is formed with the groove or flange for the shipping collar.

Band Clutches— Band clutches are of two kinds, viz., con- tracting and expanding. A contracting band clutch consists of a drum and a metal band surrounding it, which may be lined with friction material. One end of the band is fixed to a spider or housing carried upon one of the connected shafts, and the other end can be displaced angularly with relation to the first so as to contract the band into frictional contact with the drum. Contracting band clutches are of three different types. The

54

FRICTION CLUTCHES.

first of these, shown in Fig. 33, comprises two bands of which each extends substantially half way around the circumference of the clutch drum. The bands are generally made from thin strip steel, and lined with leather. One end of each band is hinged to one arm of a two armed spider secured to the driven shaft or clutch shaft, and the other end to the short arm of a double armed lever fulcrumed on the arm of the spider, the

FIG. 32. — PLATE CLUTCH, LEVER TYPE.

inwardly extending arm of the lever being adapted to be moved outward from the clutch axis by a sliding cone or wedge under the pressure of the clutch spring. When the levers are thus moved around their fulcra the bands are drawn tight on the clutch drum, and driving connection is established.

FRICTION CLUTCHES.

55

56

FRICTION CLUTCHES.

Fig. 34 shows the Mercedes coil clutch, which may also be regarded as a form of band clutch. The band in this case con- sists of a coil of steel, one end of which is anchored to the hous- ing of the clutch and the other end of which is attached to one arm of a double armed lever whose fulcrum support is in the end wall of the housing. The long arm of this lever is acted upon by a sliding cone against which the clutch spring presses. When the sliding cone is forced under the lever arm the steel coil is contracted upon the clutch drum and grips the latter. The

FIG. 34. — 'MERCEDES COIL CLUTCH.

housing is formed integral with the flywheel and the drum is se- cured to the clutch shaft. This clutch is entirely enclosed and runs in oil.

Theory of Band Clutch — In Fig. 35 is shown a sketch of a band clutch in which d9 represents a small arc of con- tact of the band on the drum and 6 the angle or arc of contact between this section dO and the point of contact between band and drum nearest to the free end of the band. At the free end a pull Pi is exerted on the band. Owing to the friction between

FRICTION CLUTCHES.

57

the band and drum the pull on the band varies from point to point of its length. Let the pull at one side of the differential section d 0 be represented by P and that on the other side by P + d P, as indicated in the sketch. Also let the normal pressure between the band and drum on the section d 0 be represented by N and the f rictional force resulting therefrom by / N. When the. system is in equilibrium the forces in any direction are equal to zero. Hence, taking the forces in the horizontal plane, dd d8

(P + dP)cos- -fN — Pcos— = O

2 2

But the cosine of an infinitely small angle is equal to unity, hence dP = fN (15)

FIG. 35. — DIAGRAM OF BAND CLUTCH. Now, taking the forces in the vertical plane,

dO d6

N — (P + d P) sin -- P sin — = O

and since the sine of an infinitely small angle is equal to the arc, we may write $0 je

N — (P + d P) -- P — = O 2 2

do

The term d P — , a differential expression of the second

2

order, may be neglected, and we may write N = P d 0

58 FRICTION CLUTCHES.

or

AT

= 7* ; (16)

Dividing equation (15) by equation (16) we get dP

= f d e

dX Now the integral of a differential expression of the form

x is log x (the natural logarithm, whose base is 2.71828. Hence

log P = / 0 + C or

log P = / 0 4- log c (17)

Remembering that in all logarithmic systems the logarithm of the base is 1, we may write

log e* 0 = fO X 1 =fO Inserting this value of / 6 in equation (17) we have

log P = log ef 0 + log c and taking antilogs — P = c e< e (18)

To find the value of the constant c we make 0 equal to zero, in which case P equals the initial pull Pi applied to the free end of the band.

Pi = c e°. But any term with the exponent zero is equal to unity, hence

C = Ft

and inserting this value in equation (18) we have

P = Pi ef 0 (19)

This latter equation gives the pull on the band at any angle 6 from the point of contact between band and drum nearest the point of application of the initial pull. The total frictional force F between the band and drum is equal to the difference be- tween the initial pull and the pull at the point of contact between band and drum farthest from the point of application of the initial pull —

F = P, ef e — P, = P, (ef 0 — 1) and

P1 = — — (20)

efe—i

In using this equation the arc 0 must be expressed in radians. Values of the expression ef & — 1 for various values of / 9 may be found from Fig. 36.

FRICTION CLUTCHES.

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FIG. 36. — CURVE GIVING RATIO BETWEEN FRICTIONAL FORCE AND

PULL ON BAND.

Sample Calculation— Now let it be required to design a band clutch for a four cylinder 4x5 inch motor, which, as we have seen, develops a torque of 133 pounds-feet. Suppose we choose a drum 12 inches in diameter, then the frictional force required at the surface of the drum is

133 X I2 = 266 founds. 6

Let the band be made of steel and lined with leather, so we can figure on a coefficient of friction f = o.2. In the case of a clutch comprising a band extending all around the drum the arc of contact will be about 5.5 radians, and in the case of a clutch with two bands, each extending half way around the drum, the arc of contact of each will be about 2.5 radians. These figures are approximate and the correct arcs of contact would have to be determined from the drawings.

Let us take the case ot a single band brake. Inserting values in equation (20) we have

60 FRICTION CLUTCHES.

This is the pull which must be exerted on the free end of the band. The pull of the fixed end on its anchorage is equal to the pull on the free end plus the friction,

133 + 266 = 399 pounds,

and the band and its anchorage must be designed sufficiently strong to withstand this stress.

If the band is of uniform width the normal pressure at its contact surface varies from end to end, being least near the free end and most near the fixed end. From equation (16) it will be seen that N varies directly as the pull P on the band. We know that the f rictional force F = 266 pounds, and since the coefficient of friction is 0.2, the aggregate normal pressure is

OAA

Q 2 = 1,330 pounds.

Also, if we allow an average unit pressure of 18 pounds per square inch, then the frictional area required is

1,330 — jg— = 74 square inches,

and since the drum has a diameter of 12 inches, and consequently a circumference of 37.68 inches, it would have a width of

3; 53 = 2 inches (appr.)

The normal pressure, as already stated, will not be uniform but greater near the fixed end than near the free end in the propor- tion of 399 : 133 or 3 to 1. Hence the lining will wear faster near the fixed end.

Effect of Centrifugal Force — At high "speeds, like those em- ployed in automobile clutches, the centrifugal force on the band h:,s quite an effect on the friction between the band and the drum, and this is the cause of the chief difference between a contracting band clutch and an expanding band clutch. The above analysis with respect to the frictional force between band and drum at low speeds applies equally to both types of band clutches, but the centrifugal force tends to expand the band, and hence to decrease the frictional force of a contracting clutch, and to in- crease the frictional force of an expanding clutch.

Let w be the weight of a section of the band 1 inch in length. Then the weight of an element d & of the band is w r d Q and the centrifugal force on this element (see equation 31, Vol. 1) is

1.226 (iv rdB)1r — = 0.102 w if r* d e,

FRICTION CLUTCHES. 61

where w is the speed in revolutions per second and r the radius in inches. This force, which we will denote by Fc d 0 (Fc being the centrifugal force on a section of the band equal to one radian), in a contracting clutch acts in the same direction as force N. Hence we may write the equation of the forces in the vertical plane —

Transposing and contracting,

(P — Fc) and

But since Fc is constant,

d (P — Fc) =dP = f N (equation 15). Hence

Integrating both sides of the equation,

log (P — Fc) =f 0 + C = f 0+ log a and taking antilogs —

P— Fc-=ae*d

In order to determine the constant for this case, let 0 = o, then P = Pi, and

hence

and

j?= P pl = (Pl — Fc) e* 6 -{- Fc — PI

Multiplying out,

Transposing

and dividing by the coefficient of Pi,

Comparing equation (21) with equation (20) we see that the effect of the centrifugal force on the band of a contracting clutch is to increase the required pull on the free end of the band by an amount equal to the centrifugal force on a section of the band one radian in length. This might have been expected, since the total centrifugal force on the band is 2irFc, and if the band

62 FRICTION CLUTCHES.

moves radially outward under this force a distance x, then the free end of the band will be moved a distance 2 •* x. Hence the motion of the free end is 2 TT times greater than the radial mo- tion, and the force in the direction of motion of the free end 2 v times smaller than the radial (centrifugal) force.

Equation (21) is applicable to contracting band clutches at all speeds. A similar analysis may be applied to expanding band clutches, and the resulting equation for the initial pull required is the same as (21), except that the sign of the term Fc is reversed, the centrifugal force in this case adding to the normal pressure, instead of subtracting from it. Therefore, for expand- ing clutches —

F

P> = — Fc (22)

e'0-1

Returning to the examples of a band clutch for a motor devel- oping a torque -of 133 pounds-feet, let the band weigh 0.1 pound per inch of length ; then

Fc = 0.102 X 0.1 X 202 X 62 = 147 pounds the initial pull becomes

p* = 3?IT+ 147 = 28° Pounds, and the pull at the anchorage of the band is

280 + 266 = 546 pounds.

Expanding Band Clutches — Expanding band clutches of the type shown in Fig. 37 require comparatively little pressure to hold them in engagement at high speed, since the centrifugal force on the band presses it against the inside of the clutch drum. The advantage of this fact is doubtful, however, since the greatest torque is produced by the motor — and, conse- quently, the greatest frictional force required of the clutch — at low motor speed. If in this type of clutch the spring were to act against a sliding cone, which through a connecting linkage acted on the free end of the band, the latter at high speed would not be released from the drum when the sliding cone was withdrawn, owing to the fact that the centrifugal force on the band would then produce the necessary fric- tional force betwen band and cone to hold the load. Conse- quently, the operating mechanism must be so arranged that when the sliding sleeve is moved by pressing on the clutch pedal the band is positively released from the drum. The band rs made of band steel, faced with leather, and supported by a skeleton drum which can be cast of aluminum, or the band may be made of a ribbed iron casting yieldingly supported

FRICTION CLUTCHES.

63

by a bracket. The engaging pressure is furnished by a tension spring, whose one end is anchored to the web of the band supporting drum. In some designs a second spring must be provided to keep the sliding cone in contact with the operating lever, which spring may either surround the clutch shaft and press directly against the cone, or may be anchored to some part of the car frame and draw the clutch pedal in the direction corresponding to clutch engagement.

In another type of band clutch both ends of the band are free and the middle is anchored to a bracket on the driving shaft. In this case one-half of the band is drawn tighter on

FIG. 37.— EXPANDING BAND CLUTCH.

the drum by the friction between band and drum, and the other half is unwound, as it were. Hence the effects of the friction on the pull or tension in the halves of the band exactly neutralize each other and can be neglected in cal- culating the frictional force. Let P be the pull exerted on each free end of the band, and suppose that under this pres- sure the ends move together a distance x. Then, if the band is supposed to be of circular shape, both before and after

64

FRICTION CLUTCHES.

FRICTION CLUTCHES. 65

the contraction, the radius will be reduced by — • Since the

27T

ratio of circumferential to radial motion is 2 ^ the ratio of circumferential to radial pressure is — and the total normal

2 IT

pressure is 2-rrP, which when multiplied by the coefficient of friction gives the total frictional force.

Expanding Block Clutches — This type of clutch, which is widely used in stationary work, is rarely found in automobile practice. It consists of a drum and two or more blocks or segments which by means of toggles or right and left hand screws can be expanded against the rim of the drum. The blocks are in driving connection with a spider secured to the clutch shaft. The calculation of such a clutch is very simple. From the arrangement of the mechanism the multipli- cation of the spring pressure at the friction surface can be readily calculated and the frictional force is then equal to the product of the normal pressure by the friction coefficient. These blocks or segments are often faced with fibre or leather. though they may also have metallic surfaces. The Metallurgique clutch, a typical expanding segment clutch with right and left hand screw operating mechanism, is shown in Fig. 38. In the Mais truck clutch, the clutch surface, instead of being a cyl- indrical envelope, is corrugated, so as to increase the normal pressure on the frictional surface on the principle of a wedge.

Clutch Shaft Dimensions — The torsional strength of shafts is calculated by means of the formula M = 0.196 d3 S,

where M is the torsional moment in pounds-inches, d the diameter of the shaft .in inches, and 5" the safe torsional stress in pounds per square inch. 6" can be figured at 5,000 pounds per square inch for carbon steel and 7,000 pounds for nickel and chrome-nickel steel. The torsional moment of a four cylinder 4x5 inch engine would be

12 x 133 = 1,596 pounds-inches. Hence

0.196 d*x 5,000= 1,596 and

d=\/ - ? - = 1.18 — say i T3s inch,

r 0.196 X 5000 for carbon steel.

Of course, if the shaft is weakened in any way, as by being squared for a coupling, the diameter should be made propor- tionally heavier. The stress allowed in the shaft seems to be

66

FRICTION CLUTCHES.

very low, bat a high factor of safety is necessary, since, owing to changes in the coefficient of friction of the clutch facing and adjustment of the spring pressure, the torque transmitting ca- pacity of the clutch may be greatly increased and much greater torques than that of which the engine is capable continuously may be produced by "jamming in" the clutch while the engine is racing, thus withdrawing some of the energy stored up in the flywheel. All other parts of the clutch transmitting the torque of the motor should be calculated on the same basis, allowing a factor of safety of about 10.

In these calculations, as well as in the calculations of other transmission members, unless exceptions are specifically men-

FIG. 39. — BLOCK AND TRUNNION TYPE UNIVERSAL AND SLIP JOINT.

tioned, a torque based upon a brake mean effective pressure of 80 pounds per square inch is to be used. That is to say, the constants of all formulae to be given will be based on this engine torque, which may be found from Chart I.

Connection Between Clutch and Change Gear — In a cone clutch the torque of the motor is transmitted by the cone and the hollow shaft to which it is secured, and since the cone must move in an axial direction when it is engaged and disengaged, there must of necessity be a slip joint in the transmission line between the clutch and the change gear. The same applies to some other types of clutches, as, for instance, multiple disc clutches in which the inner drum serves also as the presser. Moreover, unless the change gear housing and engine crank case are rigidly secured together, it is very desirable that a double universal joint be interposed between clutch and change gear, so there may be no binding of the bearings of either member when the vehicle frame "weaves" or distorts in consequence of road shocks, and also so as to obviate the necessity of absolute align- ment in assembling. A favorite construction of universal and

FRICTION CLUTCHES.

67

slip joint in connection with cone clutches is the block and trun- nion type illustrated in Fig. 39. The shaft is forged with a transverse hub which is drilled to receive a trunnion. Over this trunnion are slipped two square blocks of steel, adapted to slide lengthwise in slots formed on the inside of the hollow shaft. These slots may be cut in a planer or shaper in a short length of hollow shaft which is flange-bolted to the adjacent trans- mission part, or the slots may be milled entirely through the wall of the hollow shaft, for a certain distance from the end, and a piece of steel tubing forced over the end of the shaft as far as the slots extend.

In calculating the necessary size of the blocks and trunnions a unit pressure of 1,200 pounds per square inch can be figured

FIG. 40.— INTERNAL AND SPUR GEAR TYPE OF UNIVERSAL AND SLIP JOINT.

on between the blocks and the walls of the slots in which they slide, and a unit pressure of 1,800 to 2,000 pounds per square inch between the trunnions and the blocks. In order to obtain the maximum bearing surface with a given outside diameter of hol- low shaft, the blocks are often beveled off on the outside and beveled out on the inside. These blocks are hardened and the hollow shafts case hardened, to reduce wear. To obviate rattling of the intermediate shaft against the ends of the hollow shaft, a spring is sometimes placed between one of the hollow shafts and the intermediate shaft, which takes up the end play. Another method of accomplishing the same result consists in using a standard form of universal joint at one end of the short inter- mediate shaft and a block and trunnion type of joint at the other. The block and trunnion type of joint must be packed in

68 FRICTION CLUTCHES.

grease, and to this end must be provided with a leather "boot, as shown in Fig. 39.

Another type of universal and sliding joint employed between clutch and change gear consists of spur and internal gears. A design of this kind is used on the Oldsmobile, and is illustrated in Fig. 40. The intermediate shaft is forged with flanges at both ends which are cut with spur teeth on their circumference. These teeth mesh with the teeth of internal gears bolted re- spectively to the clutch shaft and a coupling fixed to the change gear driving shaft. Since the two sets of gears do not run together it is not necessary that their teeth should be of any particular form, and substantially square teeth probably are the most advantageous. Leather discs bolted to the sides of the two gears respectively here take the place of the usual leather boots, and at the same time limit the endwise play of the intermediate shaft and thus prevent rattling.

Leather disc universals are also much used between the clutch and transmission. These are discussed in the chapter on Uni- versal joints.

End Thrust Due to Pedal Pressure. — Most modern auto- mobile clutches are so designed that when they are engaged the spring pressure is self-contained. However, when the clutch is disengaged the end thrust due to the pressure on the clutch pedal has to be taken up in some way. The clutch itself is not supported by any structural part, and this thrust may be trans- mitted either to the engine crankshaft or to the driving shaft of the change speed gear, whichever seems the most convenient and practical in any particular design. Another thing to be con- sidered is the possibility of dismounting the clutch without re- moving the engine or gear box — especially those clutches vith renewable wearing surfaces.

CHAPTER III.

SLIDING CHANGE SPEED GEARS.

Historical — Many different devices have been tried for changing the gear ratio between the motor and the driving wheels of an automobile, and the change gear was long thought to present the most difficult problem in automobile design. Daimler and Benz, the pioneers of the gasoline automobile, both used belts and stepped pulleys in their earliest designs. The Daimler motor was taken up in France by the firm of Panhard & Levassor, and after a few experiments with belts M. Levassor, the engineer of the concern, introduced the sliding pinion change speed gear in combination with the leather faced cone clutch. The idea of meshing toothed gears by shifting them axially was at first ridiculed as crude and unmechanical, but in the end the system, after having undergone a number of important refinements and modifications, proved more satisfactory on the whole than all others, and it is now in almost universal use.

Levassor's change gear is illustrated in Fig. 41. It con- sists of two parallel shafts mounted in bearings in an alumi- num gear box. The first of these shafts, known as the pri- mary shaft, is in driving connection with the clutch. This shaft is squared and carries a set of three toothed gears or pinions, whose common hub has a square hole broached through it to make a sliding fit with the square shaft. On the secondary shaft are carried three other toothed gears, each of such a diameter as to properly mesh with one of the gears on the primary shaft. The gears on both shafts are so spaced that by shifting the primary set corresponding gears on the two shafts can be brought in to mesh successively with- out interference from the other gears. Shifting of the sliding set is accomplished by means of a hand lever located con- venient to v the operator, and a suitable connecting linkage. The secondary shaft at its rear end carries a bevel pinion meshing with a bevel gear on a cross shaft or jackshaft, from

69

70

SLIDING CHANGE SPEED GEARS.

which the power is transmitted to the rear wheels by means of side chains.

One disadvantage of Levassor's gear set was that the power was transmitted through a pair of toothed gears — with consequent power loss, noise and wear — even at high car speeds, when there was absolutely no occasion for it, since the speed was not changed by the gearing. This objection was overcome in a change gear brought out some years later by Louis Renault, which differed from Levassor's in that the gears of the two shafts were rolled into mesh instead

FIG. 41. — SKETCH OF LEVASSOR'S SLIDING CHANGE SPEED GEAR

of being slid into mesh. The primary shaft of this gear set was in two parts, the forward or driving part, and the rear- ward or driven part, the latter being journaled at its forward end inside the former. The secondary shaft served as a countershaft through which the motion was transmitted for low and intermediate speed and for reversing. For high speed the two parts of the primary shaft were locked together by means of jaw clutches formed integral with gears on the two parts of the primary shaft, which could be slid into en- gagement. This gave the so-called direct drive, the power being carried directly through the gear set without being transmitted through the toothed gears. The direct drive fea- ture was soon also incorporated in the Levassor type of slid-

SLIDING CHANGE SPEED GEARS.

71

ing gear, as shown in Fig. 42. This gear, which is known as the three speed and reverse progressive sliding gear with direct drive on high, was used very extensively for many years, and is still being used to some extent, especially on commercial vehicles.

As the speed capabilities of automobiles increased it be- came customary to fit change gears giving iour forward gear changes and one reverse, so as to enable the operator to run the engine near its most advantageous speed under all road conditions. Now, a four speed gear constructed on either the original Levassor principle or the direct drive principle comes out exceedingly long, as may be seen from Fig. 43, which represents the non-direct type. Not only does this lead to a bulky and heavy gear box, but the shafts, being relatively

r

FIG. 42. — SLIDING GEAR WITH DIRECT DRIVE.

long, are likely to be insufficiently rigid and to spring and bend under the thrust on the gear teeth, the gear thus oper- ating noisily and inefficiently. The great length with this construction is mainly due to the fact that the gears on each of the shafts must be spaced relatively far apart so as to avoid interference. This difficulty was first overcome by Wilhelm Maybach, engineer of the Daimler Motor Company, of Cannstadt, Germany, who with a non-direct drive type of sliding gear used two sliding sets. This principle was later also applied to the direct drive type, and proved so popular that at present it is used on pleasure cars almost exclusively, and also largely on commercial vehicles, and not only for four speed gears but for three speed as well.

72

SLIDING CHANGE SPEED GEARS.

Three speed and reverse gears usually have two sliding sets and four speed and reverse gears three. The several sliding sets are operated by means of a single lever, convenient to the driver, which lever, in addition to its motion for shifting the gears, has a motion at right angles to the plane of the former motion, for picking up and dropping the different sliding sets. This type of change gear is known as the selective type of sliding gear. It has the advantage over the other, the pro- gressive type, that the driver may change directly from any one gear to any other without passing through intermediate ears, which is not possible with the progressive type of gear.

FIG. 43.— PROGRESSIVE TYPE FOUR SPEED AND REVERSE SLIDING GEAR.

A sketch of a four speed selective sliding gear is shown in Fig. 44. By comparing this figure with Fig. 43 the saving in length by the use of the selective principle becomes apparent. Gear Material— It is absolutely necessary to use high grade materials for the gears of sliding gear sets. Owing to the fact that driving and driven gears are often running at greatly different pitch line velocities when they are meshed, the teeth "clash" together with considerable force, and their ends would soon be battered up if they were made of soft metal. Hardening the gears involves considerable difficulty, because if they are hardened after they are finished they are very likely to warp on being quenched, and hence to run noisily, whereas if they are hardened before being finished they can be finished only by grinding.

SLIDING CHANGE SPEED GEARS. 73

The gears may be made of either ordinary low carbon steel (so- called case hardening steel), low carbon nickel or low carbon chrome vanadium steel, all of which steels are case hardened; or they may be made of high carbon chrome nickel or high carbon chrome vanadium steel, gears of these materials having been used both in the natural state and hardened by quenching. The last two materials have exceedingly high elastic limits when properly heat treated, but they are so difficult to forge and machine that gears made of them are very expensive. These materials are fairly hard in the natural state, and gears of them therefore can be used in that state; but such gears wear faster than case hardened gears, and since they are more ex- pensive they are now no longer used, except possibly in ex- ceptional cases. Gears of chrome nickel and chrome vana- dium steel with a carbon content of 0.45 per cent., hardened

FIG. 44. — SELECTIVE TYPE FOUR SPEED AND REVERSE SLIDING GEAR

through and through, are used on the higher grades of cars. When gears are carbonized for case hardening the carbon is allowed to penetrate to a depth of 3*2 inch. Following are the standard specifications and heat treatments of steels suitable for sliding gears that have been adopted by the Society of Automobile Engineers:

Specification No. 1020—0.20 per cent, carbon steel. The fol- lowing composition is desired :

Carbon 0.15% to 0.25% (0.20% desired)

Manganese 0.30% to 0.60% (0.45% desired)

Phosphorus not over 0.045%

Sulphur not over 0.05%

This steel forges and machines well and is particularly

74 SLIDING CHANGE SPEED GEARS.

suited for case hardening. It has an elastic limit of 35,000 pounds per square inch in the annealed state and as high as 70.000 pounds when cold rolled or cold drawn. For sliding gears this steel should be treated as follows: After forging, machining and cutting the teeth, carbonize at a temperature of between 1,600° and 1,750° Fahr., cool slowly in the carboniz- ing mixture, reheat to 1,550-1,625 ° Fahr., quench, reheat to 1, 400° -1, 450°, quench and draw in hot oil at a temperature of from 300° to 450° Fahr.

Specification No. 2320—3^ per cent, nickel steel. The fol- lowing composition is desired:

Carbon 0.15% to 0.25% (0.20% desired)

Manganese 0.50% to 0.80% (0.65% desired)

Phosphorus not over 0.04%

Sulphur not over 0.045%

Nickel 3.25% to 3.75% (3.50% desired)

The elastic limit of this material in an annealed condition is 45,000 pounds per square inch, with good reduction and elongation. When suitably heat treated the elastic limit may be brought up to 60,000 pounds, and even 70.000 pounds per square inch, with better reduction of area than in the annealed state. This material is carbonized and heat treated as fol- lows: After the gears are cut. carbonize at between 1,600° and 1,750° Fahr., cool slowly in the carbonizing material, reheat to 1,500°-1,550° Fahr., quench ; reheat to 1,300°-1,400° Fahr., quench ; reheat to 250-500° Fahr. and cool slowly. The last quenching operation must be conducted at the lowest temperature at which the material will harden, which will sometimes be as low as 1,300° Fahr.

Specification No. 3140. — 0.40 per cent, carbon, chrome nickel steel. The following composition is desired:

Carbon 0.35% to 0.45% (0.40% desired)

Manganese 0.50% to 0.80% (0.65% desired)

Phosphorus not over 0.04%

Sulphur not over 0.045%

Nickel 1.00% to 1.50% (1.25% desired)

Chromium 0.45% to 0.75% (0.60% desired)

This steel contains a sufficient amount of carbon to harden without being carbonized. Heat treatment produces an elas- tic limit as high as 200,000 pounds per square inch, with good reduction of area and elongation. The steel is difficult to forge and must be kept at a thoroughly plastic heat while being forged, and not hammered or worked after dropping to ordinary forging temperature, as cracking is liable to fol-

SLIDING CHANGE SPEED GEARS. 75

low. Since the temperature range within which forging is per- missible is small, the steel must be frequently reheated. The heat treatment is as follows: Heat to 1,500°-1,600° Fahr., quench; reheat to 1,450°-1,500° Fahr., quench; reheat to 600°- 1,200° Fahr. and cool slowly. This steel cannot be machined un- less thoroughly annealed. The desired Brinell hardness for gears is between 430 and 470, the corresponding Shore hardness between 75 and 85.

Specification No. 6120. — 0.20 carbon, chrome-vanadium steel. The following composition is desired:

Carbon 0.15% to 0.25% (0.20% desired)

Manganese 0.50% to 0.80% (0.65% desired)

Phosphorus not over 0.04%

Sulphur not over 0.04%

Chromium 0.70% to 1.10% (0.90% desired)

Vanadium not less than 0.12% (0.18% desired)

The treatment of the above steel is as follows : Carbonize at a temperature between 1,600° and 1,750° Fahr. ; cool slowly in the carbonizing mixture; reheat to 1,65.0°-1,750° Fahr., quench; reheat to 1,475°-1,550° Fahr., quench; reheat to 250°-550°, and cool slowly. The heating for the second quench should be con- ducted at the lowest temperature that will harden the carbonized

Specification No. 6145. — 0.45 per cent, carbon chrome-vanadium steel. The following composition is desired:

Carbon 0.40% to 0.50% (0.45% desired)

Manganese 0.50% to 0.80% (0.65% desired)

Phosphorus not over 0.04%

Sulphur not over 0.04%

Chromium 0.70% to 1.10% (0.90% desired)

Vanadium not less than 0.12% (0.18% desired)

This steel hardens without being carbonized and attains an elastic limit of as high as 200,000 Ibs. per square inch. The proper treatment for gears is as follows : Heat to 1,525°-1,600° Fahr.; hold at this temperature one-half hour to insure thor- ough heating; cool slowly; reheat to 1,650°-1,700° Fahr., quench; reheat to 350°-550° Fahr., and cool slowly.

For the gear shafts 0.45 per cent, carbon steel, 3^ .per cent, nickel (0.30 per cent, carbon) or 0.30 per cent, carbon chrome nickel steel is used.

Gear Reduction Ratios — With very few exceptions sliding pinion change gears pfbvide either three or four forward speeds, besides one reverse speed. Four speed gear sets are

76 SLIDING CHANGE SPEED GEARS.

fitted, as a rule, to the more expensive pleasure cars and to the larger sizes of commercial vehicles manufactured. It is cus- tomary to proportion the different gear reductions so they will substantially form a geometrical series. For instance, in a three speed gear the reduction ratio of the intermediate gears is generally about 1.8, and that of the low gears 3.2, which latter figure is substantially the square of 1.8. If the motor is relatively powerful in respect to the weight of the car and the speed to which it is geared on direct drive, then these reduction ratios of the gear set can be made somewhat smaller; in the opposite case they should preferably be somewhat greater.

In four speed gears the reduction ratio of the low gears (first speed set) varies from 3.25 to 4.25, being generally near 4. With a geometrical progression, calling the first speed

ratio r, the second speed ratio would be (ty \ and the third

speed ratio ^ r t There is a tendency, however, to make the reductions of the two intermediate gears a little smaller, the idea being that the speed shall not be too low while driving on the intermediate gears, but the first speed gear must be sufficiently low to provide ample driving torque for all emer- gencies. The general run of ratios falls within the following limits :

First speed 3.75—4.25

Second speed 2 --2.2

Third speed 1.4—1.6

Fourth speed Direct drive.

The reverse gear ratio is generally made somewhat greater than that of the low gear — as great as the design permits.

Arrangement of Gears — Referring to Figs. 42 and 44, it will be seen that in these gears (which represent the modern types) the driving part of the primary shaft carries a pinion which meshes with a gear on the secondary shaft. These two gears remain constantly in mesh, while the rest of the gears are shifted into mesh when it is desired to use them. It will be noticed that the gear on the secondary shaft has about twice the pitch diameter as the driving pinion on the primary shaft, hence the secondary shaft runs at all times at about one-half the speed of the engine. There is an alternate construction in which the constantly meshed set of gears is located at the rear end of the gear box, but this is subject

SLIDING CHANGE SPEED GEARS. 77

to the disadvantage that when the direct drive is in operation, which it is a very large proportion of the time the car is in use, the secondary shaft runs at substantially twice engine speed, and the pitch line velocity of the constantly meshed gears is practically twice as great. This arrangement is now nearly obsolete, and with it has passed the practice of en- tirely disconnecting the primary and secondary shafts from each other when engaging the direct drive.

Form of Gear Teeth— There are two forms of gear teeth in use, the i4l/2 degree involute and the stub tooth. The latter, which was specially created to meet automobile requirements, is used in the great majority of cases. The involute tooth, shown in Fig. 45 at A, is the standard form of tooth for machine cut gear- ing for ordinary purposes. Its general proportions are given in the Appendix to Volume I. The tooth contact surfaces make an angle of 14^ degrees with a radial plane through the axis of the gear. The stub tooth, illustrated in Fig. 45 at B, is not as high as an involute tooth of the same circular pitch, and has a greater contact angle (20 degrees). Rules for the general proportions of stub teeth were also given in the Appendix to Volume I.

Stub tooth gears are much stronger than involute tooth gears of the same circular pitch, and that is the reason they have sup- planted the latter. It is sometimes FIG. 45.— INVOLUTE 14^ DEGREES TOOTH urged against the AND STUB TOOTH. stub tooth gear

that the radial

thrust between centres of shafts, which is proportional to the tangent of the pressure angle, is somewhat greater with the stub tooth, but since the radial thrust is only a fraction of the whole gear load on the shafts, this objection is not a very serious one. Another special form of tooth, intended to have some of the same advantages as the stub tooth, is known as the "long addendum." While the total working height is tbe same as that of the standard involute tooth, seven-tenths of this height is above the pitch circle and only three-tenths below it in the pinion ; three-tenths above and seven-tenths below it in the gear. Calculation of Gears — In determining the necessary di- mensions of change speed gears it is advisable to calculate the engine torque on the basis of 65 pounds per square inch brake m. e. p., because the permissible stress in the gear teeth decreases

78 SLIDING CHANGE SPEED GEARS.

rapidly as the pitch line velocity increases, hence the torque at normal engine speed should be figured with. The dimensions of gears necessary to transmit a certain torque at a certain angular velocity are calculated by means of a formula given by Wilfred Lewis in a paper read before the Engineers' Club of Philadelphia in 1893. This formula reads

w = S p f y,

where w is the tangential force in pounds; 5", the stress in the material of the teeth, in pounds per square inch ; p, the circular pitch ; f, the face of the gear in inches, and y a constant depend- ing upon the form and number of teeth in the gear. The follow- ing table gives the values of y for 14^2 degree involute teeth for that range of tooth numbers which is likely to be used in auto- mobile work :

TABLE I— VALUES OF y FOR 14# DEGREE INVOLUTE TEETH.

12 teeth 0.067 21 teeth 0.092

13 " 0.070 23 " 0.094

14 " 0.072 25

15 " 0.075 27

16 " 0.077 30

17 " 0.080 34

18 " 0.083 38

19 " 0.087 43

20 " 0.090 50

0.097 0.100 0.102 0.104 0.107 0.110 0.112

The above formula may be rearranged so as to directly give the width of face required — w

f = (23)

Spy

With stub tooth gears, owing to the fact that the height of the tooth is not proportional to the circular pitch, the Lewis formula is not directly applicable, since the value of the constant y changes with the pitch of the gear as well as with the number of teeth. For this form of gearing the following simplified formula may be used:

w f = , (24)

S 2

where z is a constant depending upon the pitch and the number of teeth in the gear. The values of z for the three pitches and the numbers of teeth that are likely to be used in automobile change geans are given in the table on the following page.

Pitch Line Velocity and Allowable Stress — In three speed gears the pitch line velocity of the two gears that remain constantly in mesh (where these are located at the motor end) varies between 90 and 100 per cent, of the piston speed ; in other words, the pitch diameter of the constantly meshed pinion varies

SLIDING CHANGE SPEED GEARS. 79

TABLE II— CONSTANTS FOR STUB TOOTH GEARS.

No. of Teeth	5-7 Pitch.	6-8 Pitch.	7-9 Pitch.
14	0.078	0.061	0.051
'5	0.081	0.064	0.053
16	0.083	0.066	0.054
17	0.084	0.067	0.055
18	0.086	0.068	0.056
19	0.088	0.069	0.058
20	0.090	0.071	0.059
21	0.091	0.072	0.060
23	0.093	0.074	0.061
25	0.095	0.07S	0.062
27	0.098	0.077	0.064
30	0.100	0.079	0.066
34	0.104	0.082	0.068
38	0.108	0.085	0.071
43	0.111	0.088	0.073
50	0.116	0.091	0.075

between 57 and 64 per cent, of the length of piston stroke, the higher figure being more suitable for high powered motors. In four speed gears the pitch diameter of the constantly meshed pinion is made from 57 to 77 per cent, of the length of stroke. The average ratio between length of stroke and pitch diameter of the constantly meshed pinion is 0.6 in three speed gears, and 0.7 in four speed gears.

As to the allowable stress in the material of the teeth, this varies greatly with the pitch line velocity, and, of course, also depends directly upon the physical properties of the material used. Besides, it is logical that the stress in the constantly meshed pair of gears should be somewhat less than the stress in the gears pertaining only to one particular speed, since the constantly meshed pair works under load as much as the several other pairs collectively. The author has gone over the data of a great many sliding gear sets, and finds that the following stresses in gear teeth give good results in the intermittently meshed pairs of gears :

TABLE III— ALLOWABLE UNIT STRESS IN ALLOY STEEL GEAR TEETH, CASE HARDENED.

Pitch Line Velocity. Allowable Stress.

(Ft. P. M.) (Lbs. P. Sq. In.)

750 30,000

900 27,000

1050 24,000

1200 21,000

1350 18,000

1500 . 15,000

80 SLIDING CHANGE SPEED GEARS.

TABLE IV— ALLOWABLE UNIT STRESS IN CHROME NICKEL

AND CHROME VANADIUM STEEL GEAR TEETH,

HARDENED ALL THROUGH.

Pitch Line Velocity. Allowable Stress.

(Ft. P. M.) (Lbs. P. Sq. In.)

750 60,000

900 53,000

1050 47,000

1200 42,000

1350 38,000

1500 34,000

1650 30,000

1800 27,000

In the above two tables the pitch line velocity is based on a piston speed of 1,500 feet per minute.

For the constantly meshed pair of gears the stress in the teeth should be taken 15 per cent, less than for the intermittently meshed gears.

In calculating the face of the gear it is to be remembered that the engaging edges of the teeth have to be chamfered in order to insure positive meshing, and this chamfering necessarily somewhat reduces the effective width of the gear face. In pro- gressive sliding gears some of the gears are chamfered on both sides, while in selective sliding gears the gears are chamfered on one side only. The loss in the effective width of the face amounts to about & inch for each chamfer. Another thing that deserves consideration is that, after the gear shifting linkage has become somewhat worn, there is a possibility that when the gears are meshed by the operator they will not be accurately opposite each other, with the result that some of the face width will be ineffective, and it is well to also allow iV inch for inaccurate meshing 01 the sliding gears. This makes a total allowance, for chamfer and inaccurate meshing, of l/% inch for sliding gears chamfered on one side only and $s inch for sliding gears cham- fered on both sides. If it is desired to make the gears of carbon steel, case hardened, the stresses in the teeth must be taken somewhat lower than the allowable stresses in alloy steel case hardened, for the same pitch line velocity.

Application of Formula. — We will now calculate the dimen- sions of a change speed gear for a four cylinder 4x5 inch motor, the gear to be of the three speed selective type. The driving pinion would have a pitch diameter of

0.6 x 5 = 3 inches. We will use gears with 6-8 pitch teeth, hence the pinion will

SLIDING CHANGE SPEED GEARS. .81

have 18 teeth. We found that in three speed gears the low speed reduction is usually about 3.2, and it is customary to make the reduction ratio of the constantly meshed set of gears the same as that of the low gear set. Hence the reduction ratio of either set should be about

3.2 = 1.8 (approximately), and the number of teeth for the driven member of the con- stantly meshed set should be

1.8 X 18 = 32 (approximately).

The low gear set should have the same number of teeth as the constantly meshed set, and the intermediate gear set should both have an equal number of teeth, since the constantly meshed set gives the full reduction (1.8) desired for the intermediate speed. Since the sum of the numbers of teeth must be the same for each set, each gear of the intermediate speed set must have 18 + 32 - = 25 teeth.

2

The torque of the motor, on the basis of 65 pounds per square inch brake m. e. p. is (Equation 1) : 4X5X4X4X65 - — - = 108 pounds-feet

The pinion of the constantly meshed set has a pitch radius of 1^4 inches, hence the tangential force on the pitch circle is 108 X 12

= 864 pounds.

At 1,500 feet piston speed the pitch line velocity is 1.5 TT - X 1500 = 1413 ft. p. m.

5

We will assume that the gears are to be made from low carbon alloy steel and to be case hardened, and from Table III we see that at this pitch line velocity the permissible stress is

16,800 pounds — 15 per cent. = 14,300 pounds. From Table II we find the value of the constant z for an 18 tooth 6-8 pitch gear to be 0.068. Hence, according to equation (24), the necessary face width is

864

- = 0.888— say tt inch. 14,300 X 0.068

The tangential force on the pitch line of the intermediate gears is greater than that on the pitch line of the constantly meshed set in the proportion of the number of teeth of those members of the constantly meshed and the intermediate sets which are se- cured to the secondary shaft. In the present case the force is

82 SLIDING CHANGE SPEED GEARS.

32 864 X — = 1,106 pounds.

25

The pitch line velocity of this set at 1,500 feet piston speed per minute is 25

1,413 X — = 1,104 ft. p. m.

•j£

At this speed the allowable stress in the teeth (see Table III) is 23,000 pounds per square inch. The value of constant z for 25 teeth of 6-8 pitch is 0.075. Hence the effective width of the face should be

= 0.641 inch,

23,000 X 0.075 and the total width of face

0.641 + 0.125 = 0.766 inch — say it inch.

For the low gear set the pitch line pressure figures out to 1,536 pounds, and the pitch line velocity to 530 ft. p. m. From Table III we hnd the allowable stress in the teeth to be 29,000 pounds per square. inch, and the value of constant z for 18 teeth is 0.068. Hence the total width of face of the low gear should be

- - £§- - - + o. 125 = 0.905— soy \\ inch. 29,000X0.068

It will be seen that the widths of face of the three gears come out almost the same, and, as a matter of fact, in many three speed sliding gears all of the gears are made of the same face width. Some designers simplify their calculations by merely calculating the required width of face for the constantly meshed set and making all other gears of the same width of face.

In practically every case the sliding member of the low gear set serves also to give the reverse, hence the face width of the reverse pinions is fixed by the face width of the low speed gears

Pressure on Bearings — The earlier change gears of the slid- ing type were fitted with plain bearings, but anti-friction bearings present such important advantages that they are now almost invariably used in this part of a motor car, radial ball bearings being used in the majority of gear boxes, and roller and cup and cone ball bearings in some instances. The bearing; have considerable influence on the design of the case, and ;n order that the proper sizes may be selected the gear loads on them have to be accurately calculated.

In Fig. 46 is shown a diagram of a pair of gear teeth in mesh. We will assume the teeth to be of stub form and their contacting surfaces to make an angle of 20 degrees with the plane through the axes of the two shafts. The pressure be-

SLIDING CHANGE SPEED GEARS.

83

tween the two teeth, which is represented by the line A D is normal to the contact surface. On the other hand, the tan- gential load on the gear, which is represented by the line A C, is normal to the plane of the axes and, therefore, makes an angle of 20 degrees with the tooth pressure A D. In fact, the tooth pressure A D may be resolved into two components: one, A C, normal to the plane of the gear axes and tangential to the pitch circles, which causes the driven gear to turn, and the other, A B, in the plane of the gear axes, which tends to force the gear shafts apart.

FIG. 46.— COMPOSITION OF GEAR TOOTH REACTION. Let T be the torque transmitted by the driving gear and r its pitch radius, then the tangential force is

and the tooth pressure is

A D =

TX 12

r X cos 20

There is, however, another factor to be taken into account, namely, trie friction of the teeth as they move over each other.

84 SLIDING CHANGE SPEED GEARS.

When the teeth first come together their outer ends touch each other, and they partly slide and partly roll over each other until they are in full mesh. This frictional force is in the plane of the contact surface and is represented in the diagram by A E. The resultant of this frictional force and the normal pressure on the tooth surfaces is represented by A F. The friction angle D A F may be taken at 5 degrees, which will make the angle between the tangential force and the resultant of the tangential force, the radial bearing pressure and the frictional force on the teeth, 25 degrees. Neglecting the fact that D F is not quite in line with C D, we may write T X 12

A F =

(25)

r X cos 25°

Equation (25) gives the resultant reaction at the tooth surface of any pair of meshing gears, if T is made equal to the torque

FIG. 47.

of the driving member and r equal to its pitch radius. It is now to be shown what bearing pressure results from this tooth reaction.

In Fig. 47, A represents the shaft of the driving pinion which has a torque T impressed upon it at some point in front of the bearing. This shaft is provided with a lever arm B, representing a portion of the driving pinion, which lever presses against the end of another lever C, similarly mounted upon the secondary shaft. The contact surfaces of the two lever arms make an angle of 25 degrees with the plane of the axes of rotation, so that the pressure between them makes an angle of 25 degrees with a tangent to the circles described by the centres of the con- tact surfaces. Now, the reaction of lever C on lever B produces a moment P X r around the axis of primary gear shaft A. The principle that action and reaction are equal and opposite applies

SLIDING CHANGE SPEED GEARS.

85

c3
D	[
LtJ.
ii- R,	D	[		m R2

to moments the same as it does to forces, and the reaction of the bearing on shaft A tends to turn lever B around the centre line of contact D, with the same torque, but in the opposite direction, as the contact pressure P tends to turn the arm around the axis of p

shaft^. Hence Pi represents the re- action of the bear- ing on shaft A and P2 the pressure of shaft A on the bearing.

Each of the gears i s supported o n two bearings, these bearings being on opposite sides of the gear FIG. 48.— DISTRIBUTION OF TOOTH PRESSURE respective- BETWEEN BEARINGS.

ly, and the bearing

pressure is distributed between them in a certain proportion which we shall investigate presently. The constantly meshed pinion in many gears is an exception to this rule, since it over- hangs its bearing support. From the above we see that the pressure on the bearings supporting any gear is equal to the resultant tooth reaction, and in direction parallel to it. Another thing to 'be observed is that the pressures on the shafts of two meshing gears due to the pressure between the teeth are equal but in opposite directions. This is easily seen, since the pressure of the driving gear teeth against the driven gear teeth is equal to the reaction of the driven gear teeth, but in the opposite di- rection.

Next it becomes necessary to determine the division of the bearing pressure due to the tooth reaction, between the two bearings supporting any gear. The shaft forms a beam sup- ported at both ends, with a concentrated load at the centre of the gear. Referring to Fig. 48, let Ri and R* be the reac- tions at the supports, or loads on the bearings; P the total bearing load due to one pair of gears; x, the distance of the centre of the gear from the centre of the left hand bearing and y the distance from the centre of the right hand bearing.

Then, taking moments around the centre plane of the gear

g<5 and

SLIDING CHANGE SPEED GEARS.

(P—

Except when the direct drive is being used, two pairs of gears are in mesh and transmitting power simultaneously, viz.,

FIG. 49. — CONSTANTLY MESHED AND INTERMEDIATE SPEED GEARS (SEEN FROM ENGINE END.)

the constantly meshed pair and one of the other pairs. However, the bearing pressures due to these two pairs of gears are not in the same direction, and therefore cannot be added together directly, but must be added by means of the parallelogram of forces. This may be seen from Fig. 49, which is a front view of the constantly meshed and intermediate speed pairs of gears. In this figure, Pi represents the reaction of the contsantly meshed gear C on the constantly meshed pinion A, and P2 the pressure of the intermediate pinion D on the intermediate speed gear B. The loads on the bearings of the primary shaft R are equal and parallel to Pi and P2, while the loads on the bearings

SLIDING CHANGE SPEED GEARS.

87

of the secondary shaft are equal and parallel to Pi and P2, but oppositely directed. All of these forces make an angle of 25 degrees with the vertical.

Therefore, in order to determine the total load on the different bearings of the gear set corresponding to any particular speed or gear, we first calculate the bearing load due to one pair of gears, then find the proportion of this on each bearing; next

FIG. 50. — LAYOUT OF GEARSET UNDER CALCULATION.

we determine the bearing load due to the other pair of gears, then find the proportion of this on each bearing and finally add the two loads on each bearing together by means of the parallelo- gram of forces, which can be done either graphically or trigo- nometrically.

We will now carry this calculation through for the gear set whose gear dimensions were calculated in the foregoing. This

88 SLIDING CHANGE SPEED GEARS.

gear with its bearings is laid out in Fig. 50. The tangential forces on the pitch circles we found to be :

864 pounds on the constantly meshed gears; 1,106 pounds on the intermediate gears; 1,536 pounds on the low speed gears,

and if we assume that the reverse pinion has 14 teeth, it is 1,975 pounds on the reverse gears. Since the bearing loads are equal to Tangential Force

cos 25 degrees

and the cosine of 25 degrees is 0.906, we have for the bearing loads due to these tangential forces :

953 pounds due to the constantly meshed gears ; 1,222 pounds due to the intermediate gears; 1,693 pounds due to the low speed gears; 2,180 pounds due to the reverse gears.

Now, assume the intermediate pair of gears to be in operation. The load on bearing I due to the tooth pressure of the con- stantly meshed gears is

7.469

953 X = 832 pounds.

8.563 That on bearing II due to this pressure is

953 — 832 = 121 pounds.

The load on bearing I due to the tooth pressure of the inter- mediate gears is

4.219

1,222 X = 602 pounds.

8.563 That on bearing II due to this pressure is

1,222 — 602 = 620 pounds.

Adding the two loads on each bearing graphically, as shown in Fig. 51, we find the loads on bearings I and II to be 642 and 550 pounds, respectively. The directions of these loads are as indicated by the arrows, the gear being looked at from the front. The load on bearing V due to the tooth pressure of the inter- mediate gears is

4.219

1,222 X = 708 pounds.

7.25

The load on bearing VI due to the tooth pressure on the inter- mediate gears is

1,222 — 708 = 514 pounds.

The load on bearing IV due to the tooth pressure on the inter- mediate gears is

SLIDING CHANGE SPEED GEARS.

2.969

708 X = 1,271 pounds.

1.656

The load on bearing III due to the tooth pressure on the inter- mediate gears is

1,271 — 708 = 563 pounds.

The load on bearing III is opposite in direction to the load on bearing IV.

Secondary Shaft Bearings

Primary Shaft Bearing*

FIG. 51. — BEARING LOADS FOR INTERMEDIATE GEAR OPERATION The load on bearing IV due to the tooth pressure on th3 con- stantly meshed gears is

953 X fffiff* *,&o pounds.

The load on bearing III due. to the tooth pressure on the con- stantly meshed gears is

1,580 — 953 = 627 pounds.

90 SLIDING CHANGE SPEED GEARS.

The loads on bearings III and IV while the intermediate gear is in operation are added together graphically in the right

Secondary Shaft Bearings » Primary Shaft Searing FIG. 52. — BEARING LOADS FOR Low GEAR OPERATION.

hand diagram in Fig 51, and the magnitude and direction of the load on bearing VI are also shown.

When the low gears are in mesh the bearing loads due to the tooth pressure on the constantly meshed pair of gears will be the same as when the intermediate gears are in mesh, which

SLIDING CHANGE SPEED GEARS.

91

loads we have already found. The load on bearing I due to the tooth pressure on the low speed gears is 3.219

1,693 X = 637 pounds.

8.563

The load on bearing II due to the tooth pressure on the low speed gears is

1,693 — 637 = 1,056 pounds.

Adding the two forces on each bearing graphically, as in Fig. 52, we find the loads on the secondary shaft bearings for low

FIG. 53. — MAGNITUDE AND DIRECTION OF TOOTH PRESSURE ON

REVERSE GEARS.

gear operation to be 645 pounds on bearing I and 981 pounds on bearing II.

The load on bearing V due to the tooth pressure on the low speed gears is

3.219

1,693 X = 753 pounds.

7.25

The load on bearing VI due to the tooth pressure on the low speed gears is

1,693 — 753 = 940 pounds.

The load on bearing IV due to the tooth pressure on the low speed gears is

2.969

753 X = 1,350 pounds.

1.656

92 SLIDING CHANGE SPEED GEARS.

The load on bearing III due to the tooth pressure on the low speed gears is

1,350 — 753 = 597 pounds.

The loads on the bearings of the primary shaft corresponding to low gear operation are added graphically in the right hand diagram in Fig. 52, and we find that the load on IV is 1,263 pounds and on III, 513 pounds.

The direction of the tooth pressures on the reverse gear and

Secondary <5haft 3ectring>s.

FIG. 54.— BEARING LOADS FOR REVERSE GEAR OPERATION. pinion may be found graphically from Fig. 53. It is seen that the pressure of the idler gear on the reverse gear makes an angle of 10^ degrees with the vertical, and the reaction of the idler gear teeth on the teeth of the reverse pinion makes an angle of 46l/2 degrees with the horizontal.

The load on bearing I due to the tooth pressure between the reverse pinion and idler is

1094

2,180 X - - = 278 pounds. 8.563

SLIDING CHANGE SPEED GEARS. 93

The load on bearing II due to the tooth pressure between the reverse pinion and idler is

2,180 — 278 = 1,902 pounds.

The load on bearing VI due to the tooth pressure between the reverse gear and idler is*

6.156

2,180 X = 1,851 pounds.

7.25

The load on bearing V due to the tooth pressure between the reverse gear and idler is

2,180 — 1,851 = 329 pounds.

The load on bearing IV due to the tooth pressure between the reverse gear and the idler is

2.969

329 X = 590 pounds.

1.656

The load on bearing III due to the tooth pressure between the reverse gear and the idler is

590 — 329 = 261 pounds.

Adding the two loads on each bearing graphically (see Fig. 54) we find the loads on bearings I and II to be 570 pounds and 1,789 pounds, respectively, and the loads on bearings III and IV, 627 pounds and 2,094, respectively.

The following table shows at a glance the load on each bearing for each speed:

Bearing. I. II. III. IV. V. VI.

Reverse 570 1789 627 2094 329 1851

Low gear 645 981 513 1263 753 940

Intermediate gear 642 550 519 1263 708 514

High gear

Bearing Load Due to Bevel Gears — Cars fitted with side chain drive have a bevel gear set enclosed in the rear portion of the change gear box, the bevel pinion being keyed to the rear end of the primary shaft. Of course, the tooth reaction of the bevel gears throws considerable load on bearing VI, and this must be taken into account. In very powerful cars the bevel pin- ion is sometimes located between ball bearings on opposite sides of it, but the more common arrangement is to have only a single large radial ball bearing directly back of the bevel pinion. We will assume that in the change gear under calculation the above arrangement is used and that the ratio of the bevel gear set is 3 to 1. We will further assume that the pinion has eighteen teeth of 6 pitch and the gear fifty-four. This makes the maximum pitch diameter of the pinion 3 inches and the pitch angle such

94

SLIDING CHANGE SPEED GEARS.

that its tangent is 0.333, viz., 18° 26'. If the bevel pinion has a face of \y% inches, then the mean pitch diameter is 3 — (1^ X sin 18° 26') = 3 — (!3/£ X 0.316) = 2.567 inches,

and the mean pitch radius, 1.283 inches. • Since the motor develops a torque of 108 pounds-feet, the tangential force on the gear teeth,

FIG. 55.— TOOTH REACTION IN BEVEL GEARS.

figured as though it was concentrated at the middle of the face length, is

103 X 12

— 1,010 pounds.

1.283

The tooth reaction makes an angle of 20 degrees with the tangential force, hence its value is

1,010

= 1,074 pounds.

0.94

Now, in a bevel gear the tooth reaction is not in a plane per- pendicular to the axis of the gear, and for this reason the bearing

SLIDING CHANGE SPEED GEARS. 95

pressure is not equal to the tooth reaction, as in the case of a spur gear. We have to resolve the tooth reaction into two com- ponents, one in a plane perpendicular to the gear axis, which is equal and parallel to the load of the shaft supporting bearings, and the other in a direction parallel to the gear axis, which is equal to the end thrust. This requires three successive steps.

In Fig. 55, A B represents the normal pressure on the tooth contact surfaces. We first resolve this into a component A C in a vertical plane perpendicular to the gear axis, and a component C B in a horizontal plane through the axis of the gear and at right angles to the element of the gear tooth surface on which the tooth pressure comes. A D represents this latter component both in direction and magnitude.

AD = CB = AC tan 20° = T tan 20°.

The latter may be resolved again into a component A E perpen- dicular to the gear axis and a component D E parallel to the gear axis.

A E = A D cos B = T tan 20° cos 6

D E = A D sin Q — T tan 20° sin 6 ......................... (26

D E represents the end thrust of the bevel pinion which is usually taken up on the radial ball bearing, though some designers provide a special thrust bearing, or use a combined radial and thrust bearing at this point. This equation is general in its nature, applying to all 14^/2 degree involute gears; while for stub tooth bevel gears tan 25° should be substituted for tan 20°.

The radial bearing load is equal to the resultant of A C and A E which is

an 20° co 3 0)2 ..................................... (27

In our example 7 =1,010 pounds. The tangent of 20° is equal to 0.364, the cosine of 0 (18° 25') is 0.949 and the sine of 0, 0,316. Substituting these values in equations (26) and (27) we find the end thrust to be

1,010X0.364x0.316=116.2 pounds, and the radial bearing load

vijOio2 + (1,010 X 0.364 X o-949)2= 1,051 pounds.

The arrow heads in Fig. 55 indicate the direction of the reac- tion of the bevel gear teeth on the bevel pinion teeth and of its components, and the resultant radial bearing pressure is in the direction of A F, which in this case makes an angle of 22>y2 de- grees with the vertical.

Like the constantly meshed pinion, the bevel pinion overhangs its bearing. From the centre of the rear ball bearing to the centre

96 SLIDING CHANGE SPEED GEARS.

cf the bevel pinion would be about i% inches, and since the dis- tance between centres of the two bearings of the bevel pinion shaft is 7^4 inches, we have for the load on bearing VI due to the tooth reaction on the bevel pinion :

M

1,051 X ^T= 1,232 pounds,

and the load on bearing V due to the tooth reaction on the bevel pinion,

1232 — 1051 = 181 pounds.

When the direct drive is employed these are the only loads on bearings V and VI, but when either of the lower gears or the reverse is in mesh the loads on bearings V and VI due to the bevel pinion tooth pressure are multiplied by the reduction fac- tor of the particular gear, and there is in addition the load due to the reduction gears on bearings V and VI which must be combined with the loads due to the bevel gears by means of the parallelogram of forces. For bearing VI this is done in Fig. 56, the values of the loads on VI shown in Figs. 51, 52 and 54 being used, and the value of the load due to the bevel gears represented in Fig. 55, multiplied by the reduction factor of the particular gear combination. It will be seen that the bearing loads due to the bevel and spur gears respectively partly neutralize each other, and that with a gear of this kind the load on the rear bearing of the primary shaft is greatest when the low gear is in operation. The tooth pressure of the bevel gears has little influence on the load on bearing V and its effect may be neglected.

Sizes of Bearings — Manufacturers of ball bearings issue tables of load capacities with the aid of which the proper size of bearing for each point can be determined. These load capacities are the loads the bearing will stand under continuous running at normal speed. Now, it will be seen from the table of bearing loads above given that the loads on all the bearings except 7 and V are a maximum when the reverse gear is in operation, and these maxi- mum loads in most instances are far greater than the loads corre- sponding to the other gear combinations. It will be remembered that the bearing loads were calculated on the basis of full engine power, and it practically never happens that the engine works at full load while the reverse gear is being used. The reverse gear is made extremely low for the sake of safety in backing, and not because an unusually large torque is needed. Hence the calcu- lated bearing loads for the reverse gear never obtain in practice, and they may be neglected when selecting the proper size of bear-

SLIDING CHANGE SPEED GEARS.

97

FIG. 56. — LOADS ON PRIMARY SHAFT

REAR BEARING (VI) WHEN

CARRYING A BEVEL PINION.

ings, though it is well to make sure that the calcu- lated load on bearing II does not exceed the rated load by more than 100 per cent.

Various constructional and operative considera- tions often influence the choice of bearing sizes. Thus, although there is a very considerable differ- ence between the maxi- mum loads on 7 and //, these bearings are often chosen of the same size ; for one reason, because it simplifies the boring of the bearing holes in the gear case, since the holes at op- posite ends can be bored in one operation. Another reason is to be found in the advantage there is in reducing the number of different parts in a car, due to the fact that a smaller stock of repair parts will suffice. When it is thus decided to use the same size of bearing at both ends of the sec- ondary shaft the size of bearing selected should have a rated load capacity intermediate between the maximum, loads on the two bearings for forward running. Thus in our ex- ample the loads are 550, 642, 645 and 981 pounds, and the No. 306 bearing would probably be selected which has a rated capacity of 860 pounds. To give a general rule, the bearings should be selected to have

98 SLIDING CHANGE SPEED GEARS.

a rated load capacity of from 75 to 125 per cent, of the calcu- lated maximum gear loads due to other than the reverse gear, depending upon the general quality of construction.

Intermediate Bearings — In the construction Fig. 50 the most heavily loaded bearing is IV, which is due to the fact that the constantly meshed pinion overhangs this bearing. Although the primary driving shaft is supported in two bearings, the load due to the tooth pressure is not divided between these bear- ings, as might possibly be supposed. The gear overhangs the bearings and the load on bearing IV from the constantly meshed gears alone is equal to the tooth pressure on the constantly meshed pinion plus the load on bearing ///. The load on bear- ing IV resulting from that on bearing V is also nearly twice the latter. The conditions are somewhat more favorable when a plain bearing is used at V, extending a considerable distance into the primary driving shaft, so that the middle of its length lies substantially in the plane of bearing IV, in which case the load on V is transferred directly to IV. In the case of unit power plants and designs of clutches requiring no slip joint in the clutch shaft, it is advantageous to use only a single bearing on the primary driving shaft, as the load on the bearing will then be less than that on IV in Fig. 50.

In large gear boxes the constantly meshed pinion is some- times supported in two bearings, as shown in Fig. 57, one on either side, the inside bearing being carried on a pedestal or in a partition wall in the case. The loads are then divided be- tween the two bearings in the inverse proportion of the centre distances. Bearing / may also be placed inside the constantly meshed gear, causing the latter to overhang, an arrangement that naturally suggests itself when the constantly meshed pinion is carried in two bearings. It increases the load on bearing I and reduces that on bearing //, so their maximum loads will be about equal, which may be considered an advantage if both are to be made of the same size. However, this construction is rare.

Truck Change Gears. — In change gears designed for motor trucks the unit stresses are kept lower, for the reason that trucks are operated a great deal of the time in congested thorough- fares where it is necessary to do much driving on the lower gears. Besides, a little extra weight does not count for so much in a truck as in a high speed pleasure car. For this same reason chrome nickel or other high tensile steels are seldom, if ever, used for the gears and pinions of truck transmissions. With

SLIDING CHANGE SPEED GEARS. 99

carbon steel and low carbon alloy steel, case hardened, the fol- lowing unit stresses may be allowed in the gears \/ -

Pitch Line Allowable Velocity. Stress.

(Ft. p. m.) (Lbs. p. sq. in.)

500 20,000

600 18,000

700 16,000

800 14,000

900 12,000

1000 10,000

The bearings of commercial change gears should also be of

FIG. 57. — CONSTANTLY MESHED PINION WITH BEARINGS ON BOTH SIDES.

somewhat more liberal size than those in pleasure car gears, for the same reason.

Shaft Dimensions — One of the chief requirements in a change gear box is quiet operation, and this necessitates rigid shafts. The sizes of the shafts are, therefore, more dependent upon the maximum permissible flexure than upon the torque to be transmitted. The tooth pressure on the gears located midway between bearings creates an appreciable flexure of the shafts, and the pairs of gears located near the bearings also create some flexure, but this may be neglected. The shafts should be made of such a diameter that the maximum flexure due to any pair of gears is not more than 0.003 to 0.005 inch. In Chapter XI of Volume I is given a formula for the flexure of shafts supported

100 SLIDING CHANGE SPEED GEARS.

at their ends and carrying a concentrated load between hear- ings, viz., .

where P is the load on the shaft in pounds ; /, the length of the shaft between the centres of bearings, in inches ; d, the diameter of the shaft in inches, and x the ratio of the distance of the load from the farthest support to the distance between supports.

Applying this equation to the secondary shaft of the gear box calculated in the foregoing, in which the flexure is evidently a maximum when the low gear is in operation, we have P = 1,693 pounds I = 8.563 inches

2 x* + 2 x* — 4 x* = 0.11

If we decide to allow a maximum flexure of 0.005 inch, then 1,693 X 8.563

X a11

8,800,000 X and

. 1,693 X 8.563 X 0.11 , „ , */.* • ,

d =^ 8,800,000X0.005 = L28 ~ say 1 5/16 wch'

In some designs of change gears the secondary shaft is made of somewhat greater diameter in the middle than at the ends, •with the object of securing the most rigid shaft with the least material.

The primary shaft, since it has substantially the same span between the supports and is subjected to the same loads similarly located, should be made of practically the same diameter as the secondary shaft ; or, rather, it should have a cross section equiva- lent to that of the secondary shaft with respect to bending stresses.

Reverse Gear Arrangement — Various arrangements of gears for obtaining the reverse motion are in use. The most common is that already illustrated in Fig. 50, in which the secondary shaft carries a reverse pinion sufficiently smaller than the low speed pinion to allow the low speed gear to clear it when shifted opposite it. This reverse pinion meshes with a reverse idler on a special shaft mounted parallel with the primary and secondary shafts, usually in the lower part of the gear box.

A somewhat different arrangement is shown in Fig. 58, in •which A is a pinion of double width serving for both the low gear and the reverse; B is the low speed and reverse gear and

SLIDING CHANGE SPEED GEARS.

101

FIG. 58. — REVERSE GEAR WITH Two IDLERS.

Ri R2 are reversing idler gears on a special short shaft. Sliding gear B is shown in the position corresponding to the reverse motion. By sliding it to the left until it meshes with A the low forward speed is obtained. One advantage possessed by the ar- rangement Fig. 58 over that of Fig. 50 is that with the former there is less strain on bearing II (at the rear end of the second- ary shaft) than with the latter when the reverse gear is oper- ating.

The two types of reverse gear so far shown are used in three speed selective and in progressive type gears. In four speed gears the reversing idlers may be arranged slidably (see Fig. 59),

FIG. 59.— REVERSE GEAR WITH SLIDING IDLERS.

102 SLIDING CHANGE SPEED GEARS.

and by means of a separate sliding bar slid into mesh with both the low speed pinion and gear while the latter are out of mesh. To obtain the low speed forward, gear B is shifted to the right into mesh with pinion A. On the other hand, when it is desired to back up, gear B is placed in the neutral position (which it occupies in the illustration) and reversing pinions Ri and Rz are slid to the left into mesh with A and B respectively, as shown.

Direct Drive Clutch — There are two types of direct drive clutches in common use, viz., the jaw type, illustrated in Fig. 60, and the spur and internal gear type, shown in Fig. 61. The former type consists of jaws formed on the adjacent faces of the constantly meshed pinion and the intermediate speed gear respectively. Usually each part has four such jaws, equal in size, and subtending at the axis of the shaft an angle slightly smaller than that subtended by the space between them. The outer edges of the jaws are chamfered to facilitate engagement. The radial width of these jaws is usually made about one- quarter the shaft diameter and the length the same.

Where the spur and internal gear type of clutch is employed the constantly meshed pinion often serves as the spur member, and the intermediate speed gear is cut with internal gear teeth, in addition to its regular spur teeth, to serve as the other mem- ber. It is somewhat difficult to cut these internal gear teeth. The job can be done by counterboring the rim of the spur gear and then planing the teeth, but it is a much preferable plan to use a form of mongrel teeth made by drilling holes into a solid gear blank from the side and then chambering the blank out so as to cut away half of the stock between the holes (see Fig. 61).

Front Bearing of Sliding Gear Shaft— Notwithstanding the difficulty of keeping such a bearing effectively lubricated, a plain bearing is often used at the forward end of the squared or fluted shaft, on which the gears slide. This construction renders non- fluid oil unsuitable as a gear box lubricant. With a fluted shaft the journal would be made about three-quarters the diameter of the shaft proper so as to give a substantial shoulder, and about three diameters long. As in the case of the engine tailshaft, large oil holes and grooves are necessary, and the scheme of lubri- cation should be carefully worked out.

Instead of a plain bearing, a cylindrical roller bearing consist- ing of long, thin rollers is sometimes used, extending into the counterbore of the shaft, the same as the plain bearing. How- ever, a more common construction is to use either a single or a double row non-adjustable ball bearing, as illustrated in Fig. 61.

SLIDING CHANGE SPEED GEARS.

103

FIG. 60. — DIRECT DRIVE JAW CLUTCH.

Some designers use a specially large constant mesh pinion in order to be able to accommodate a ball bearing of sufficient capacity, obtaining the required reduction ratios by using very small intermediate, low speed and reverse pinions on the second- ary shaft. The light series of ball bearings is naturally best adapted for this purpose, since it has the least radial depth for a given load capacity. However, double row bearings seem to be preferred for this point, since it is difficult to find room for a bearing of ample capacity.

Sliding Gear Shaft — As already pointed out, in the earlier sliding change gears the sliding pinions were slid on squared shafts. These are still used to a slight extent, but have for the

FIG. 61. — DIRECT DRIVE SPUR AND INTERNAL GEAR CLUTCH.

104

SLIDING CHANGE SPEED GEARS.

most part been replaced with splined or integral key shafts. The two types of shafts are shown in cross section in Fig. 62. So- called squared shafts are not absolutely square, but have rounded corners. They are made from round shafts by milling four flats on them to such a depth that the distance between opposite flats is 0.8 the diameter across the corners, or the diameter of the original shaft. Denoting the side of the square formed by the flats by h, the torsional strength of such a shaft is about 0.21 h3S pounds-inches, h being given in inches. The flats are often fin- ished by grinding, and if the shaft is to carry long sleeves sup- porting the sliding gears, they are sometimes cut with wavy oil grooves so that oil may flow to parts of the shaft that are never exposed by the sliding members. Some makers bore the hole in the gear to a slightly greater diameter than the side of the

•OJ8<f-

\

FIG. 62. — SECTIONS OF SQUARED AND SPLINED SHAFTS.

squared shaft, so that when the hole is broached out, from two- thirds to three-fourths of its side will be a plane surface and the rest cylindrical. (See Fig. 63.) This facilitates the broach- ing, tends to obviate gripping of the sliding members an^ does not appreciably reduce the effective bearing surface, because the pressure is localized near one edge of the flat.

As compared with the squared shaft, the splined shaft pos- sesses the advantage that it takes the torsional load perpendicu- larly on the sides of the splines, whereas in a squared shaft most of this load comes close to one edge of the flats, with the result that in the latter the unit pressure may become very high and the lubricant may in consequence be squeezed out, which is not likely to occur with a splined shaft.

SLIDING CHANGE SPEED GEARS.

105

In American practice, splined gear shafts are made with four splines for small and moderate sized gear boxes, while in large gear boxes six splines are used. European practice tends to a more general use of six splines. Uneven numbers of splines have also been used, but they are subject to the disadvantage that they make it very difficult to caliper the diameters of the shaft accurately. The ratio of the bottom diameter of a splined shaft to the top diameter or diameter over the splines is gen- erally about 0.8, and the width of the splines is made about one- quarter the bottom diameter, or 0.2 times the outside diameter. (For S. A. E. standard splined fittings see Appendix.)

Practice varies as to the manner of locating the gears. Some

FIG. 63. — BROACHED SLIDING GEAR WITH PART OF FLAT RELIEVED.

FIG. 64. — FLANGE BOLTED GEARS ON SECONDARY SHAFT.

manufacturers grind the outside of the shaft— that is, the top surfaces of the keys, and let the gear ride on these surfaces, using the broached hole in the gear. Others grind out the hole in the gear (after the latter has been hardened) true with the pitch circle or the bottom circle, and let the gear ride on the bottom surface of the splined shaft. Both methods involve cer- tain difficulties, and it is hard to say which is the better of the two, everything considered.

Proportions of Gears — The rims of gears below the tooth annulus are made of a thickness varying from 0.5 to 0.6 the circular pitch, and the webs about the same. Since teeth of 6 and 6-8 pitch are used almost exclusively in sliding gears, whose cir-

106

SLIDING CHANGE SPEED GEARS.

cular pitch is' 0.52 inch, both rim and webs are generally made Y&, inch thick. When the web is located to come flush with one side of the rim, the latter may taper from */i to 5/16 inch in width, but it is undoubtedly preferable to have the web central. In this connection it is worth remembering that substantial rims and webs and liberal fillets tend to quiet operation, and the general tendency seems to be toward a slight increase in the thickness of the sections. The smaller pinions, of course, are made solid, and only the larger gears are webbed. As regards the secondary shaft gears, in American practice they are gener- ally secured to the shaft by means of Woodruff keys, while European designers, as a rule, flange-bolt the gears to the shaft

FIG. 65.— SECONDARY SHAFT ASSEMBLED WITH GEARS AND BEARINGS.

or to a sleeve keyed to the shaft. Frequently the gears for the two intermediate speeds are bolted to the same flange, as shown in Fig. 64. One of the reasons for flange-bolting the gears is that they are then of very simple form and are not so likely to distort in hardening. To insure concentricity the web of the gear is bored out to fit accurately over an enlarge- ment of the shaft. The gears may also be riveted to the flanges. The gears on the secondary shaft must be accurately and securely fixed in position longitudinally, and this is generally accomplished by turning the shaft with a collar near its middle against which a gear is forced from either end, and using tubu- lar spacers between these inner and the outer gears on the shaft, as shown in Fig. 65.

SLIDING CHANGE SPEED GEARS.

107

Instead of keying the gears on the shaft and supporting the latter in antifriction bearings in the housing, the entire set of secondary gears may be made in a single forging, which re- volves on a stud secured in the housing, as illustrated in Fig. 66. Bronze bearing bushings are forced into the hub of the gear set from both ends. This construction is made possible by modern methods of gear planing. It is obvious that a sec- ondary gear set so arranged may be made quite rigid, and as the journal diameter is small, the frictional loss should be low. If the gear case has a separate end plate the shaft may

FIG. 66. — SECONDARY GEAR ASSEMBLY ON STATIONARY SHAFT.

even be dispensed with, the gear set then being forged with journals at both ends which have a bearing in the housing.

Manufacture of Gears — Blanks for the pinions and gears of sliding gear sets are made either from bar stock or from drop forgings, the larger blanks being generally -forged on account of the saving in machine work. Before any work is done upon the blanks they should be annealed to remove the forging strains, and thus obviate undue distortion during the subsequent heat treatment.

It is not intended to go extensively into the question of gear cutting in this volume, because it is an involved subject and has

108

SLIDING CHANGE SPEED GEARS.

been ably treated in special works. Suffice it to say that gear teeth are either milled by means of formed cutters, or planed with ordinary cutters, which by means of templates or other devices are moved so as to produce the proper shape of tooth. In all gear cutting there are two operations, the rough cutting or stocking and the finish cutting. Only very little stock should be left for the latter operation, so that there may be very little

FIG. 67.— FORCING GEARS ONTO SECONDARY SHAFT.

strain on the cutting tool, and thus the highest degree of ac- curacy attained.

After the teeth are finish-cut, the ends from- which the gears are to be meshed have to be chamfered. This may be done by means of a milling machine attachment, as illustrated in Fig. 68. The attachment is clamped to the table of the milling ma- chine, and the chamfering tool is held in the spindle of the latter.

SLIDING CHANGE SPEED GEARS.

109

The attachment comprises a work spindle on which the gear to be chamfered is mounted, which is alternately fed toward and away from the revolving cutter by means of a cam driven through gearing from the main shaft of the attachment. On a secondary shaft is mounted a worm of the same pitch as that of the gear to be chamfered and in which it is meshed. This secondary shaft is driven through gears from the main shaft. The main shaft is driven by belt from an overhead countershaft, which is entirely independent of the milling machine counter- shaft. As the main shaft revolves the worm, meshing with the

FIG. 68. — "LONG ARM" TOOTH CHAMFERING ATTACHMENT.

gear to be chamfered, turns it, and at the proper intervals the cam mechanism feeds it toward and away from the V-shaped revolving cutter. The gear to be chamfered is thus automati- cally indexed.

The contour of the chamfering may be changed by using spe- cial cams, or special cutters, or both. The profile at the end of the tooth may be changed by swiveling the attachment on the

110 SLIDING CHANGE SPEED GEARS.

table. The end of the tooth may thus be left at right angles with the axis of the gear or at any desired angle.

The next operation in the manufacture of the gears is to harden or case-harden them. In case-hardened gears, if it is desired that any portion of the surfaces should remain soft, this can easily be accomplished by leaving about 1/32 inch extra stock on these surfaces and removing it after the gear is carbonized and before it is quenched. This practice also tends to prevent undue distortion of the gear during the quenching. Another process designed to accomplish the same purpose, and which is undoubt- edly less expensive, consists in copper-plating the gears just before the finishing cut is taken and the ends are chamfered. The result is that when the gears are carbonized after these machining operations only those portions of the gear from which the copper shell is removed will take up carbon from the pack and will become hardened on being quenched. Gears thus treated are so little distorted by the quenching that they can readily be corrected to the desired degree of accuracy.

Every effort must be made in the manufacture of gears to get every part as nearly true as possible. It would not seem to matter much whether or not the sides of the gear blanks are turned absolutely true. This, however, is quite essential, for the reason that gears are generally cut in "gangs," a considerable number of them being forced over the mandrel and the milling cutter, etc., then being fed through the whole set in one opera- tion. Now, if the sides of the blanks are not absolutely parallel there is a tendency to distort the mandrel when the nut is turned up, and thus to produce irregularity in the teeth.

For the grinding of the hole after the teeth are cut, as re- ferred to in the foregoing in connection with splined shafts, a special fixture is required for holding the gears. This consists of a face plate with Several studs driven into it parallel with its axis and at such a distance therefrom that they fit accurately between the teeth of the gear at the pitch circle. These locate the gear concentrically with the grinder spindle, and it may then be held in position by means of a couple of clamping plates and bolts. The fixture serves also as a rough gauge for indicating the accuracy of the gear cutting operation. If the teeth have been cut too deep, the gear will be loose in the fixture, whereas if they have not been cut deep enough it will not enter between the studs.

Tester for Gears — A more delicate gauge or gear tester is made as follows (Fig. 69) : A vertical shaft A is fixed to a

SLIDING CHANGE SPEED GEARS.

Ill

base and provided with a bushing over which fits the gear to be tested. An eccentric stud B is mounted on the base in such a po- sition that when the line between its centres is perpendicular to the line between the axis of its top portion and that of the fixed stud, the distance between the latter two axes is the exact dis- tance between the axes of the gear shafts. An indicating hand or pointer C secured to the eccentric stud then points to zero. The pointer moves over a double scale, and therefore shows exactly how much the gear is either too small or too large.

Unless the teeth are finished by grinding after hardening — a process that is seldom applied at present — some allowance must

FIG. 69. — GEAR TESTER.

be made for swelling or distortion during the hardening process, by either cutting each of the gears 0.005 to 0.010 inch small on the pitch diameters, or else placing the two shafts that much farther apart than the calculated distance.

Sliders— The individual sliding members in a gear set are operated by means of sliding bars, ^ to ^ mcn m diameter, and arranged parallel with the gear shafts, which carry forks that fit into grooves formed in the projecting hubs of the gears. Two such sliding bars are provided in all three speed gears, and three in some four speed gears. Generally the sliding bars are placed

112

SLIDING CHANGE SPEED GEARS.

side by side, but sometimes they are arranged concentrically. The sliders are located inside the gear box near one of the side walls thereof, and have their bearings in the end walls. In order to insure accurate meshing of the gears, as well as to lock them out of mesh, a locking arrangement similar to that illustrated in Fig. 70 must be provided. It consists of a spring pressed plunger or ball which enters V slots in the sliding bar, corre- sponding to the neutral position of the sliding set and the two or more positions of engagement, respectively. These locking dogs will hold the slider in the neutral position when it is dis- connected from the operating lever and enable the driver to find the correct meshing position when it is connected thereto. While this method of locking the sliders is not positive, it is sufficiently dependable for all practical purposes. In most designs of selec-

FIG. 70. — LOCKING DOG FOR GEAR SLIDER.

tive gear the operation of picking up one slider with the shifting lever entails the automatic and positive locking of the other sliders.

Mounting of Bearings — If the gear case is made of aluminum and anti-friction bearings are used, the latter are generally mounted in bronze bushings, instead of directly in the casing. This practice was introduced because the aluminum was con- sidered too soft, and it was thought necessary to distribute the pressure over a greater surface than that of the bearings alone. With the improvements which have been made in aluminum al- loys in recent years this is no longer absolutely necessary, but the practice is still adhered to by some designers. The bushings are provided with outward radial flanges so as to be held secure- ly against endwise motion.

SLIDING CHANGE SPEED GEARS.

113

The inner races of radial ball bearings should always be forced onto the shaft under moderate pressure, and should be securely clamped between a substantial shoulder on the shaft and a nut which is locked by some approved means. Of the outer races on a single shaft not more than one should be firmly secured in a lengthwise direction, as otherwise there is danger of subjecting the bearings to undue end thrust.

Taking up the bearings on the secondary shaft first, the inner races are secured to the shaft as above described. Of the outer races one may be clamped between an inward flange on the bushing and the bearing end cap, as shown in Fig. 71A, and the other one made a sliding or "suction" fit in the casing or bush-

FIG. 71.— MOUNTING FOR SECONDARY SHAFT BEARINGS.

ing and left free to move endwise. An alternate arrangement consists in leaving both outer races free endwise and taking up the end thrust on hardened thrust buttons fitted into the shaft ends and the bearing caps, respectively. Set screws with rounded points may be screwed through the centres of the caps to take the place of the buttons therein as shown at B in Fig. 71.

The rule that the inner races must be firmly clamped between a shoulder and a nut or spacer applies to all bearings. Like- wise, if there are two or more bearings on one shaft, the outer races of all but one of them should be free endwise, and if a thrust bearing is used in addition to radial bearings, the outer races of all the latter should be free. In some cases the for-

114

SLIDING CHANGE SPEED GEARS.

ward bearing on the primary shaft is subjected to the end thrust of the clutch spring, and should then be provided with a ball thrust bearing. This is generally placed between the two radial bearings. However, the necessity of firmly clamping both of the inner races on the shaft and allowing the outer races some end- wise motion should not be lost sight of in this case. Fig. 72 shows two ways in which these requirements can be met. At A is shown the Alco design, which employs a single thrust bearing. The design shown at B is taken from a paper read by F. G. Barrett before the Institute of Automobile Engineers, London, on February 14, 1912. With the latter design the thrust bear- ings can be properly adjusted and the adjustments locked before these bearings are placed on the shaft.

Geared-up Fourth Speed — The greatest transmission effi- ciency and the most silent operation are obtained with the direct

A B

FIG. 72.— MOUNTINGS FOR PRIMARY SHAFT BEARINGS.

drive, and the designer, therefore, should strive to so propor- tion his gear reduction that the car can be driven on direct drive under all normal conditions. This means that there should be a relatively large reduction between the gear box and rear wheels. However, in many types of cars very high maximum speeds are desired, which conflicts with the requirement of a high reduction ratio in the final drive. These conflicting require- ments led to the construction of four speed gears in which the direct drive is the third speed, and the fourth is a geared-up speed, 25 to 30 per cent, higher than the direct drive. Fig. 73 shows the lay-out of the Winton change gear, with indirect fourth speed. The geared-up speed is obtained by placing on the second- ary shaft near its rear end a gear with a larger pitch diameter

SLIDING CHANGE SPEED GEARS.

115

than the constantly meshed gear, adapted to be meshed with a sliding pinion on the driven primary shaft of a smaller pitch diameter than the constantly meshed pinion. In a gear of this type it is advantageous to keep the reduction ratio of the con- stantly meshed pair of gears low, as otherwise the pitch line velocity of the high speed gears will be very high and their operation is likely to be attended by considerable noise.

Gear Cases — The gear cases of nearly all pleasure cars are cast of aluminum alloy of the same composition as that used

FIG. 73. — LAYOUT OF WINTON CHANGE GEAR WITH GEARED- UP FOURTH SPEED.

for the engine crankcase. However, manganese bronze is also used for that part of the case which supports the shafts and on which the greater part of the strain comes. The gear boxes of many motor trucks, especially those of European design, are made of cast steel, and cast iron cases are also in use. There are two common arrangements of the shafts in a gear

116

SLIDING CHANGE SPEED GEARS.

box. Either the secondary shaft is located directly underneath the primary shaft or the two shafts are located in a horizontal plane. There is, of course, a third possible arrangement, where the plane of the shafts is neither