Table of Contents

A. MISCELLANEOUS BELT AND CHAIN DRIVES

1.0 INTRODUCTION 468

2.0 BELT DRIVES 468

2.12.22.3

V-Belt Design Data Design Data for Neoprene and Buna Belts Design Data for Polyurethane Belts

469476477

3.0 CHAIN DRIVES 478

3.13.23.33.43.53.6

Design Data for No. 25 Single Strand Roller Chain Drive Selection Sprocket Selection Recommendations Selection of Center Distances Chain Drive Selections for Very Slow Speeds Lubrication

478480481481481482

4.0 TENSIONING MECHANISMS 484

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A. MISCELLANEOUS BELT AND CHAIN DRIVES1.0 INTRODUCTION

Since it is our specialty to service the users of small commercial quality drive components, we will limit ourselves to design dataup to 1 H.P. capacity. This should prove most useful to the designer of business machines, computer equipment, instrumentationand small automatic machinery of all kinds. This write up is divided Into the following categories:

Belt Drives* • Flat Belts • Synchronous Belts • V-Belts • Variable Speed Pulleys • Round Belts

Chain Drives • Roller 1. Standard Size 2. Miniature • Ladder • Bead

Tensioning Mechanisms

2.0 BELT DRIVES

(a) Flat BeltsFlat belts find considerable usage in applications requiring small pulley diameters, high belt surface speeds, low noise levels, lowweight and inertia. They cannot be used where absolute synchronization between pulleys must be maintained because they relyon friction for their proper functioning, All flat belts except the toothed type mentioned in the next data section are subject tocreepage because of the relative motion between the pulley surface and the adjacent belt surface which is under loaddeformation from the combined tension and flexural stresses. Flat belts must be kept under tension to function and therefore require tensioning devices. Fractional H.P. belts aremanufactured two ways. 1. Woven and 2. Film. Woven belts are usually made from Nylon or Dacron fibers and impregnated withrubber. Sometimes they are made from fiberglass for high temperature applications. They can operate at pulley speeds up to140,000 R.P.M. and pulley diameters as small as 3/8 inch. Film type belts are made from Mylar. Annular rings are die cut from a basic Mylar sheet and subsequently formed into anendless belt. Belt thickness can be as small as .0005 inches, thus enabling operation over pulley diameters as small as .050inches for tens of millions of cycles. Creep and pitch Pine variations are also small because of the thickness of the belt. Film typebelts are finding considerable application in miniature and airborne tape recording equipment.

(b) Synchronous BeltsSynchronous belts, commonly known as timing belts, are basically flat belts with a series of evenly spaced teeth on the insidecircumference, thereby combining the advantages of the flat belt with the_____________*Reprinted with permission from Uniroyal Industrial Products Manual 181 and Browning Manufacturing Div. of Emerson Electric Co. Cat. No. 6.

468

positive grip features of chains and gears. There is no slippage or creep as with plain flat belts. Required belt tension is low, therefore producing very small bearing loads.Synchronous belts will not stretch and require no lubrication. Speed is transmitted uniformly because there is no chordal rise andfall of the pitch line as with roller chain. The tooth profile on timing belt pulleys is an involute curve similar to a spur gear. Unlikethe spur gear however, the outside diameter of a timing pulley is manufactured smaller than its pitch diameter, thus creating animaginary pitch diameter which is larger than the pulley itself. This is illustrated in Figure 1. Backlash between pulley and beltteeth is negligible. Synchronous timing belts are also available in double sided designs which offer an infinite number of new design possibilities oncomputer equipment, business machines, office equipment, textile machines and similar light duty applications. Belts with drivingteeth on both sides make it possible to change the direction of rotation of one or more synchronized pulleys with only one belt.The inside and outside teeth are identical as to size and pitch and operate on standard diameter pulleys. Since this group of belts is extremely popular and their design and application is widespread, a separate section is devoted tothem in this handbook.

(C) V-BeltsV-belts find frequent application where synchronization between shafts is not important. V-belts are easily installed and removed,quiet in operation, low in maintenance, and provide shock absorption between drive and driven shafts. Normally V-belt drives operate best at belt speeds between 1500 to 6500 feet per minute. Ideal speed (peak capacity) isapproximately 4500 feet per minute. The maximum satisfactory speed ratio is approximately 7 to 1. V-belt drives operate at90−98% efficiency.

The limitations of V-belts are the following: a. Improper belt tension can reduce service life. b. Belt life at increased temperatures (above 180ºF) is significantly shortened. c. Centrifugal force prevents the use of belts above 10,000 feet per minute.Vbelts are made with a fiberglass reinforced neoprene core with a jacket of fabric impregnated neoprene which protects theinterior and provides a wear-resistant surface for the belt. To facilitate interchangeability and ensure uniformity, manufacturershave developed industry standards for the various types of belts. Most of these standards were set by the AMA (RubberManufacturers Association) and the MPTA (Mechanical Power Transmission Association). A properly designed drive should take into consideration the following factors: 1. Horsepower, speed, and characteristics of the prime mover. 2. Required driven speed, and horsepower or torque characteristics of the driven unit. 3. Approximate center distance, including minimum and maximum allowable. 4. Space limitations on various driven components. 5. Type of operation (intermittent, steady) and number of cycles or hours of service. 6. Method of adjustment for tensioning and takeup. 7. Service conditions−heat, abrasive atmosphere, moisture, oil, chemical, etc. 8. Necessary shaft diameters, lengths, and keyways.

2.1 V-BELT DESIGN DATAOverload Service FactorsLoad and operating characteristics of both the driving and driven units must be considered thoroughly in the selection of FHPDrives. It is essential that all drives be designed for maximum load conditions to be encountered. Most drives will at some time be overloaded, perhaps only momentarily, and it is good practice to have predetermined drivecapacity to handle this overload. For good design and satisfactory drive life all drives must be selected with careful considerationof two fundamental Conditions:

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1. The motor must have greater capacity than the driven unit. 2. The drive must have greater capacity than the motor. The following are suggested overload service factors for various types of driven units:

TABLE 1Driven Unit Service Factors

Driven Unit

OverloadFactor

Driven Unit

OverloadFactor

Agitators Liquid Semi-LiquidCompressor Centrifugal ReciprocatingConveyors and Elevators Package, Oven BeltFans and Blowers Centrifugal, Calculating ExhaustersFood Machinery Slicers Grinders, MixersGenerators Farm Lighting, Exciters

1.21.4 1.21.4 1.21.4 1.21.4 1.21.4 1.2

Heating and Ventilating Fans, Oil Burners StokersLaundry Machinery Dryers, Ironers washersMachine Tools Home Workshop and WoodworkingPumps Centrifugal ReciprocatingRefrigeration Centrifugal ReciprocatingWorm Gear Speed Reducers, Input Side

1.22.4 1.21.4 1.4 1.21.4 1.21.41.0

In addition to the Overload Service Factors determined by the driven unit of a drive, there are certain types of driving unitshaving starting or operating characteristics which call for increased capacity. If the driving unit is one of the following types, add 0.4 to the Driven Unit Service Factor: Frequent Starting and Stopping. High Starting Torque Motors. Gas Engines. Across the Line Starting. Multiply the combined Overload Service Factors by the horsepower rating of the driving unit to get the required normal ratingof the drive. Example: 3/4 H.P. Gas Engine driving Generator. Suggested Overload Factor 1.2 plus .4 or 1.6. Normal rating of drive is 3/4 x 1.6 or 1.2 H.P.

Drive Selection Charts for FHP Belts Always try to use the largest pulleys possible in any drive unit. Small pulleys are less efficient and greatly reduce belt lifebecause of slippage and extreme flexing of the belt. Where small pulleys are a must and speed is high, select the lightest belt 2L.Table 2 shows the recommended minimum pulley diameters as shown in NEMA standards. It is a result of concern for bearingand shaft loading rather than belt flex life.

TABLE 2

Belttype

Size of belt(in.)

Minimum pitch diameterRecommended Absolute

2L3L4L

1/4x1/83/8x7/321/2x5/16

1.01.52.5

1.01.51.8

TABLE 3 RECOMMENDED MINIMUM PULLEYDIAMETERS (INCHES) FOR ELECTRIC MOTORS

Motorhorsepower

Motor rpm575 695 870 1160 1750 3450

1/23/41

2 1/233

2 1/22 1/2

3

2 1/22 1/22 1/2

−−2 1/22 1/2

−−−−

2 1/4

−−−−−−

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Correction Factor for Loss in Arc of ContactThe rated horsepower tables in the following pages are based on a 1 to 1 ratio with 180º of belt wrap around each pulley. Whenpulleys of different diameters are used there is a loss in tractive effort on the smaller pulley. The loss in arc of contact from 180ºfor different drives can be determined in the following manner:

Loss in Arc of Contact (in degrees) =

Where D = large pulley P.D. d = small pulley PD. C = center distance between shafts (inches)

The correction factors for loss in Arc of Contact in degrees on small pulleys are:

TABLE 4

Loss in Arcof Contact

CorrectionFactor

Loss in Arcof Contact

Correction Factor

0º5101520 2530354045

1.00.99.98.96.95

.93

.92

.89

.89

.87

50º55606570 75808690

.86

.84

.83

.81

.79

.76

.74

.71

.69

Calculate Belt LengthThe belt length can be calculated by means of the formula given previously:

L = 2C + 1.57(D + d) +

When pulley pitch diameters are used, resultant is belt pitch length. When pulley outside diameters are used, resultant is beltoutside circumference. The chart shown below is a rapid means of determining outside belt length and is accurate enough for most applications.

Belt Size StandardsAllowance must be made for manufacturing tolerances in belt width and length. This tolerance may be summarized by length,regardless of cross-section, as follows:

TABLE 5

NominalOutside Length

OutsideLength Variation

14"- 20"21 - 6061 - 7980 -100

+1/8", −3/8+1/4, − 5/8

+5/16,−11/16+5/8, − 7/8

Allowance for Installation and Take-UpThe center distance obtained from the Nomogram in Figure 1 or the belt length formula is that for a normal length belt underproper operating tension. Provision must be made in the drive for moving the pulley centers closer to permit mounting of thebelts without forcing them over the pulleys. Also, the pulley centers must provide for adjustment beyond the nominal centerdistance to compensate

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for belt stretch and wear. In addition, the manufacturing tolerance of the belts themselves must be considered. These variousfactors can be consolidated into terms of overall adjustment as follows:

TABLE 6

Belt Size Range Minimum Recommended Adjustment

6R12 − 3L110 − 3L2403L250 − 3L3703L380 − 3L6104L170 − 4L3704L380 − 4L6004L610 − 4L7904L800 − 4L1000

1 - 1/8"1 - 1/41 - 1/21 - 1/41 - 3/42 - 1/42 - 3/4

The driving unit should be installed at approximately the mid-point of this recommended adjustment, thus allowing formounting of belts and subsequent take-up. The application of proper tension is of extreme importance as insufficient tension canshorten belt life, cause slippage and even cause belt breakage. Excess tension can cause bearing wear and possible failure. Proper tension can usually be determined by pressing on the belt after installation. It should feel alive and springy. Keeping thebelt at correct tension by periodic inspection will insure longer life and better service.

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ldlersWhen it is impossible to provide sufficient center-distance adjustment for belt length take-up, an adjustable idler should be used.The idler should contact the slack side of the drive. It should be sufficiently adjustable to permit installation of the belts withoutforcing and to permit take-up for tensioning. On most V-belt installations having sufficient arc of contact, the best type of idler is a grooved pulley placed against the insideof the drive. Its diameter should not be smaller than the minimum listed pulley diameter. When the diameters of both the driver and driven pulley are appreciably larger than the listed minimums, it is desirable thatthe diameter of the idler be at least as large as the diameter of the small pulley. A flat idler pulley is not generally advisable for V-belt drives. If it is used, there must be no crown on the face. A flat idlerpositioned inside the drive against the bottom of the belts must be no smaller than the minimum pulley diameter. If used againstthe top of the belts, the flat idler diameter should be at least as large as the diameter of the small pulley. The 6Z9 series productgroup in the SDP catalog provides adjustable drive tighteners and idlers.

To Calculate a Drive 1. Apply overload service factor in Table 1 2. Determine speed ratio and pulley diameters 3. Select belt sizes from Table 7, 8 and 9 4. Select belt length from Tables 10 and 11 or calculate from formula 5. Check power loss because of belt contact arc loss 6. Check belt speedExample−−A 1/3 H.P. Capacitor start A.C. motor is driving a reciprocating air compressor at 720 R.P.M. on approximately 12 inch centers.Maximum diameter of compressor pulley to be 8 inches pitch diameter. (a.) Overload factor is 1.8 (.4 for high starting torque motor plus 1.4 for reciprocating compressor). Normal rating of drive is .33 x 1.8 = .6 H.P.

(b.) Speed ratio =

Using the maximum allowable driven pulley pitch diameter, then the required motor pulley pitch

dia. is = 3.35

(C.) Select the desired pulley from the product section and note that a 3.35 P.D. pulley has a 3.5 outside diameter. Byinspection of Tables 5, 6 and 7 it is obvious that the 1/2 x 9/32 (4L, series) belt is the only one capable of handling the designrequirements. It is rated at 1.05 H.P. at 1750 R.P.M. (d.) The center distance between pulleys can frequently be found by referring to Tables 10 and 11. In this particular problem,the ratio of 2.4:1 is not listed and must therefore be calculated. The belt length is determined by the formula

L = 2C + 1.57(D + d) +

L = 24 + 1.57 (8.00 + 3.35) +

L = 42.25 inches pitch length

The nearest standard belt listed in the product pages is 42 inches pitch length. (e.) Remember that the above 1.06 H.P. rating is for a 1 to 1 ratio only. Calculate correction factor for loss of arc to make certainthat the choice of pulley diameters and centerdistance does not derate the 1.06 H.P. value below the drive requirements.

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Where D = Pitch diameter of large pulley d = Pitch diameter of small pulley C = center distance

Correction factor from Table 12 for 22º is .94therefore 1.06 x .94 = .99 H.P. which is still ample to carry the design load.(f.) Check that belt velocity does not exceed 5000 ft. per minBelt Velocity (feet/min.) = Pulley R.P.M. x Pitch Dia. x .262 = 3.35 x 1750 x .262 = 1550 feet/min., which is satisfactory

TABLE 7 HORSEPOWER RATINGS FOR 1/4 x 1/8 (2L)SECT. V-BELTS

TABLE 8 HORSEPOWER RATINGS FOR 3/8 x 1/4 (3L) SECTION V-BELTS

TABLE 9 HORSEPOWER RATINGS FOR 1/2 x 9/32 (4L) SECTION V-BELTS

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TABLE 10 TABLE 11 3L AND 4L CENTER DISTANCES

(d.) Variable Speed PulleysA variable speed pulley is actually a pulley with a variable pitch diameter. This variation can be achieved through manual orautomatic means depending on the design, by widening the space between the two pulley-halves. When it is used in conjunction with another fixed or variable pulley, an infinite number of speed ratios can be obtained within arange determined mainly by whether two or one variable pulleys are in the belt drive. Figure 2 illustrates a V-belt drive whereboth pulleys are variable, however one is spring loaded and the other is manually adjusted.

The driver and driven shafts are on fixed centers. Turning the control knob while the drive is running brings together the two faces of the controlling pulley, thus forcing theV-belt to ride at a greater pitch diameter. This forces the V-belt deeper into the spring loaded pulley and reduces its pitchdiameter. The related increase and decrease in pulley diameters produces an exceptionally wide range of speed control. Correctspeed is held by spring tension. There is no need for locking the adjusting screw. Stop nuts are provided to limit total adjust

475

ment if desired. This type of variable speed drive can only be adjusted while running. For limited range applications, thespring-loaded pulley can be used with a fixed pulley and adjustable motor base. Spring loaded and manually controlled pulleys as well as adjustable motor bases are available in the Product Section of the SDPcatalog.

(e.) Round BeltsRound belts, also known as O-Rings, are used in numerous applications such as photographic projectors and slide changers,dictating recorders, tape recorders, electric typewriters. Round belts require no idlers or spring devices to keep tension. Roundbelts are well suited for serpentine drives, reverse bends, and 90º twists. Their inherent damping characteristics reduce noise andtheir elasticity effectively smooth out shock and vibration. The best materials for belt drives are Neoprene, Buna and polyurethanes. The product pages list round belts made from thesematerials, ranging from 1/16 to 5/16 cross-section diameter as well as pulleys specifically designed for these belts. Thepolyurethane belts listed in the catalog employ a unique design feature where the belt can be installed or removed without takingapart adjacent mechanism components which is sometimes unavoidable with endless solid belts.

2.2 Design Data for Neoprene and Buna Belts

1. Pulley diameters should be at least four times the cross-sectional diameter of the belt. For higher speeds (above 2500R.P.M.) the 4:1 minimum ratio must be increased proportionately because of hysteresis heat build up in the frequentflexing around the pulleys. Belt life is inversely proportional to the 5th power of the pulley diameter.2. Keep belt speeds below 4000 feet per minute.3. Center distances should not exceed 4 to 5 times the larger pulley diameter.4. Temperatures ranging from 0º to 150ºF will not affect serviceability.5. While the belt is at rest, initial tension (or sketch) is distributed uniformly throughout the length of the belt. As load isapplied, the tension in the tight strand of the belt increases, while tension in the loose strand decreases. Total tension inthe belt remains constant. Effective tension is the difference between the tension in the tight strand and the tension inthe loose strand. This effective tension is equal to the load being transmitted by the belt.Tension in the loose strand should be positive in order to prevent "looping" of the belt caused by centrifugal force. For thisreason, total tension must be somewhat greater than minimum calculated effective tension.6. Overall efficiency is approximately 80%. The following formulas and tables are useful in neoprene and Buna belt designs:

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NomenclatureAWCDdEeG

========

Cross-sectional area of 0-ring, sq. in.Cross-sectional diameter of belt in.Center distance, in.Diameter of larger pulley, in.Diameter of smaller pulley, in.Tension modulus, psiElongation, per centShaft torque, lb-in.

L1 = Free length of belt, in.L2 = Installed length of belt, in.nP = Speed of smaller pulley, rpm

Power transmitted by belt, hpTe = Effective tension, lb

= T1 - T2T1 = Tension of tight strand, lb.T2 = Tension of loose strand, lb.σ = Unit belt stress, psi

TABLE 12 O-RING RATINGS

CordDiameter (in.)

Max Rating per1000 fpm (hp)

Max EffectiveTension (lb)

0.0700.1030.1390.2100.275

1/501/201/101/51/3

0.751.53712

TABLE 13 TENSILE MODULI

Material Modulus E, to50% Elongation (psi)

Neoprene50 durometer

6070

50180400

Example:Determine a suitable O-Ring drive belt for a tape recorder with a 1/50 H.P.−−600 R.P.M. motor with a 1½" pulley driving a 3"pulley:

W = .199

Therefore, a standard 0.210 cross-section O-ring will be adequate to transmit the load when installed with a 10% stretch.

2.3 Design Data for Polyurethane BeltsPolyurethane belting, listed under ROUNTHANE BELTS in the product catalog is a tabular extrusion designed so that the user cancut it to any desired length from a coil. The belt ends are then manually connected with a supplied aluminum insert that is shortenough so as not to produce a slap or thump against the minimum pulley size recommended.

DIMENSIONAL STABILITY−−Rounthane is exceptionally strong, and will maintain its original tension without idlers or springs.COEFFICIENT OF FRICTION−−Rounthane provides an excellent "pull" has a coefficient of friction from 0.5 to 0.7.FLEXIBILITY−−Rounthane's inherent flexibility permits its use in unusual configurations and limited spaces. It will perform well ondual or variable speed drives, reverse bends, radial twists and serpentines.CHEMICAL RESISTANCE−−Rounthane has an extremely high resistance to water absorption and is

477

virtually unaffected by weather, sunlight, oils, fuels, and most chemicals, as well as abrasive conditions.CLEANLINESS−−Rounthane will not attract lint or dust, will not absorb fats or greases, is easily cleaned with soap and water.TEMPERATURES−−Rounthane provides power or motion in a temperature range from − 40ºF to + 150ºF.LIMITATIONS−−Rounthane is designed primarily for fractional horsepower drives, conveyors and motion systems. It should not beused where slipping is required. Cross-drive applications should be individually evaluated. Belt hooks are not recommended.TENSIONS−−Belt tensions are acceptable up to8% for light duty Rounthane and up to 4.5% for heavy duty Rounthane. For bestresults, tensions should be experimented with on specific applications. This is easily done because Rounthane can be shortenedand refastened in seconds.INSTALLATION DATA−−Pulley diameters should be at least 8 times belt cross-section diameter. A smaller diameter will tend tothrow the belt away from the groove, particularly when tension is low. Under certain loading conditions, belt speed loss may varyup to 13% and efficiency may be reduced up to 20%.

3.0 CHAIN DRIVES

Chain drives are often used where positive synchronization between shafts and transmission of substantial torque is required.This section will be devoted to chain drives under 1 horsepower.

(a) Roller Chain 1. Standard size −− Roller chains used for fractional H.P. applications are almost always the ¼ pitch #25 single strand type. Theonly disadvantage inherent in roller chain is a phenomenon known as "chordal action" which produces a slight pulsation in theoutput sprocket and becomes less pronounced as the number of sprocket teeth are increased. Generally, the elasticity of thechain, the clearances among parts, and the oil films will damp much of the chordal action in chain drives with large sprockets.Figure 3 illustrates "chordal action" for a 9 tooth and 4 tooth sprocket. Notice the difference in chordal rise.

3.1 Design Data for No. 25 Single Strand Roller ChainThe power transmission capacity of a Roller Chain Drive is a function of the speed and number of teeth on the smaller sprocket,and, other factors being equal, is in proportion to the pin-bushing bearing area of the chain. Determine Class of Service.

Example:Determine a suitable 0-Ring drive belt for a tape recorder with a 1/50 H.P.−−600 R.P.M. motor with a 1½" pulley driving a 3"pulley:

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Therefore, a Standard 0.210 cross-section 0-ring will be adequate to transmit the load when installed with a 10% stretch.

TABLE 14 OVERLOAD SERVICE FACTORS

The Overload Service Factors in Table 14 are based on an 8 to 10 hour operating day and the driving unit one of the following types: a. Squirrel Cage Motors−−Normal Torque b. D.C. Motors c. Single Phase MotorsIf the driving unit is one of the following types, add 0.2 to the foregoing factors: a. Squirrel Cage Motors−−High Torque b. Wound Rotor (Slip Ring) Motors c. Synchronous Motors−−Normal Torque d. Line Shafts and Clutch StartingIf the driving unit is one of the following types, add 0.4 to the foregoing factors: a. Synchronous Motors−−High Torque b. Gas, Diesel or Steam EnginesFor the following special drive conditions, add 0.2 to the foregoing factors: a. 10 to 16 Hour Service b. Speed-Up Drives c. Frequent Starting d. Excessive Moisture e. Abrasive Dust f. Infrequent Maintenance g. Excess Heat h. Centers less than 20 pitchesFor the following special drive conditions, add .4 to the foregoing factors:

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a. 16 to 24 Hour Service b. Across the Line Starting c. Heavy Shock Load d. Heavy Starting Load e. Reversing Drives

The above factors which apply to a specific drive should be added together to get the total Overload Service Factor. The horsepowerof the driver is then multiplied by this total Overload Service Factor to get the Normal Rating of the Drive.

"Design Horsepower" RequiredDesign horsepower is the normal horsepower to be transmitted, multiplied by the composite Service Factor selected from Table 14and corrected with the above correction factor.

3.2 Drive SelectionThe most economical chain drive selection usually consists of a combination of:

a. b.

The shortest pitch single strand chain which has the required "design horsepower" rating when used with the desired sizesprocket at the selected sprocket speed.The selection of a center-distance that insures a chain wrap of at least 120º (meshing with at least 1/2 of the teeth) onthe smaller sprocket.

Table 15 indicates the horsepower transmitting capacity of No. 25 single strand chain. These roller chain power transmission ratingsare established for chain operation under the following conditions: 1. Service factor of 1. 2. Chain length of 100 pitches. 3. Use of recommended method of lubrication. 4. Two-sprocket drive arrangement. 5. Sprockets aligned and mounted on parallel horizontal shafts.A service life of approximately 15,000 hours may be expected when operated at full rated capacity under conditions as outlinedabove.

TABLE 15 HORSEPOWER RATINGS−−SINGLE STRAND ROLLER CHAIN No. 25

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TYPE l MANUAL LUBRICATION. Oil applied periodically with brush or spout can. The limiting RPM for eachlubrication

TYPE ll DRIP LUBRICATION. Oil applied between link plate edges from a drip lubricator. type is read from the column tothe

TYPE lll OIL BATH or OIL SLINGER. Oil level maintained in casing at predetermined height. right of the boundary linesshown.

TYPE lV OIL STREAM. Oil supplied by circulating pump inside chain loop on lower span.

3.3 Sprocket Selection RecommendationsWhere smooth performance and long service life are desired, sprockets with 17 or more teeth, but less than 67, should be used. On slow speed and special purpose installations or where space limitations are a factor, sprockets with less than 17 teeth canbe used. Use of sprockets with more than 67 teeth tends to reduce the normal service-life of the chain. Hardening the smaller sprocket teeth tends to equalize the rate of wear with the large sprocket in the drive. Sprockets shouldhave hardened teeth when any of the following conditions prevail.

1. Sprocket has less than 24 teeth and the rotative speed exceeds 600 R.P.M., or 1/8 of the maximum speedrecommended for sprockets with that number of teeth, whichever is smaller. (See horsepower rating, Table 15).

2. The chain speed is approximately 100 feet per minute or less and the chain loading is about 1/7 of the average tensilestrength or greater.

General experience shows that a sprocket speed ratio between driver and driven sprockets should not exceed about 7:1.

3.4 Selection of Center DistancesQuite often the center distance for a roller chain drive is determined within limitations by other machine components orconsiderations. Although it is not essential that the smaller sprocket have 1200 chain wrap for satisfactory operation andperformance, it is recommended as good practice when conditions permit. Chain wrap on the smaller sprocket of a two-sprocket drive with ratio 3.5 to 1 or less, will always be 1200 or more. Chain wrapincreases as center distance is increased. For an average application, a center distance of 30 to 50 pitches of chain represents good practice. For pulsating loads, shaftcenters as short as 20 pitches of chain may be preferable when sprocket sizes permit. Close attention should be given to propercenter distance for the chain length (preferably an "even" number of pitches).

3.5 Chain Drive Selections for Very Slow SpeedsWhere the linear speed of the chain is to be under 100 feet per minute and the installed chain length exceeds about 50 pitches,roller chains may be selected on the basis of their "endurance limit" and without regard for the customary horsepower ratings.Roller chains selected on this basis are considered primarily as "tension members." The "endurance limit" of a chain is the tensile load which it can accept for an infinite number of cycles without introduction ofmetal fatigue (failure). Although it is not directly related to the tensile strength, a reasonable approximation of linkplate"endurance limit" (allowable working load is 1/7 of the chain's average tensile strength.) The recommended "endurance limit" ofchains with offset links and/or connecting links with slip-fit coverplates is 1/2 of the chain's average tensile strength. Because oftheir lower capacity, offset links and slip-fit coverplates should be avoided in high load applications, and connecting links withpress-fit coverplates should be used. The tensile strength for the chains is given in the product pages as: #25 Hardened Steel−−925 lbs #25 Hardened Steel−−700 lbs

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Chains selected on the basis of "endurance limit" should be well lubricated to provide satisfactory service life. Although thefrequency of chain joint action in such drives is low, and joint wear due to articulation should be low, the lubrication is neededdue to the higher loads and to prevent corrosion.

Chain Length CalculationsChain Length may be calculated from the following formulae:

Where C1=

Center distance in number of pitches (links)−− low ratios and longcenters.

C2

=

Center distance in numberof pitches (links)−− high ratios and shortcenters.

L = Chain length in number ofpitches (links).

N = Number of teeth inlarge sprocket.

n = Number of teeth insmall sprocket.

P = Chain pitch.

If the center distance is already fixed because of other factors, an idler sprocket may be used with not less than a three toothwrap. When the idler is placed outside the closed span it should be placed nearer the small sprocket, and when inside the closedspan it should be placed nearer the large sprocket The 6Z9 series product group in this catalog provides adjustable drivetighteners and idlers.

3.6 Lubrication Chain life will vary appreciably depending on the way the drive is lubricated. A properly lubricated chain can last more than 100times as long as the same chain with poor lubrication. There are four basic types of lubrication for chain drives.

Type l −− Manual Lubrication −− Oil should be applied periodically with a brush or spout can, preferably once every 8 hours ofoperation.

Type ll −− Drip Lubrication −− Oil drops are to be directed between the linkplate edges from a drip lubricator at a rate of 5 to20 drops per minute per strand of chain. The higher rate of flow should be applied to chains operating at rotativespeeds approaching the limit for drip lubrication as shown on the horsepower rating tables. Precaution must betaken against misdirection of the drops by windage.

Type lll −− Bath or Slinger Disc Lubrication −− With bath lubrication, the lower strand of chain runs through a sump of oil Inthe drive housing. The oil level should reach the pitch line of the chain at its lowest point while operating. Withslinger disc lubrication, the chain operated above the oil level. The disc picks up oil from the sump and deposits itonto the chain, usually by means of a trough. The diameter of the disc should produce rim

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speeds between 600 f.p.m. minimum and 8000 f.p.m. maximum.Type lV −− Oil Stream Lubrication −− the oil is sprayed across the lower strand of chain in a continuous stream by a

circulating pump or a central lubricating system.

2. Miniature roller chains −− These chains are designed especially for use where space is limited, or where light weight is animportant factor. The high strength-to-weight ratio, durability, and extreme dimensional accuracy make this chain ideal forpositive power transmission or shaft synchronization where precision and compactness are important. Examples are:communications equipment, business machines, cameras, and other electronic or electro-mechanical devices. For chain speedsnot exceeding 100 feet per minute, maximum load on chain should be no more than 20 pounds. For higher chain speeds,maximum chain loading should be reduced depending on the operating conditions:length of service required, sprocket size, number of sprockets in the drive, chain speed, and the length of chain. For sprockets with 20 teeth, or less, speeds should not exceed 15,000 R.P.M. Larger sprockets should be limited to reducespeeds. For example, the speed of 60 tooth sprockets should not exceed 6,000 R.P.M. Lubrication is desirable where high loads, speeds and service life are required. Miniature chain is available from stock and listedin the Product Section of the SDP catalog.

(b) Ladder ChainLadder chain links are fabricated from wire which has been formed so that each link interlocks with its adjacent links to form acontinuous chain whose drive members appear as rungs on a ladder. Hence the name; Ladder Chain. Ladder Chain is inexpensiveand easy to install. It finds wide use in household appliances, recorders, vending machines and in various timing and electricaldevices. Ladder Chains are flexible enough so that in a 2 sprocket drive, one of the sprockets can be rotated well over 90 degreeswith respect to the other sprocket plane provided of course that sufficient center distance is allowed. Thus offering a decidedadvantage over roller chain and timing belts in this regard. Ladder Chains are assembled manually by spreading the connectingloops on the end link slightly apart with a pair of needle nosed pliers, and retightening the loops after engaging the link from theother end of the chain. Table 16 indicates the approximate H.P. ratings of the various chain sizes and material combinations listed in the SDP productcatalog.

TABLE 16

Chain No.Approx. H.P. Capacity @ 500 RPM

Hi-Tensile Steel

Basic Steel Brass

14171819

13/41/21/3

1/21/31/41/6

1/31/41/61/8

(C) Bead ChainBead Chain consists of a series of hollow metal beads, linked together by dumbbell shaped wire links. Since the beads are free-toswivel, the chain can be easily-twisted, thus making it extremely versatile for unusual drives. Bead chain because of itsconstruction should only be considered on a low speed or manual drives With low torque requirements.

483

4.0 TENSIONING MECHANISMS

Correct operating tension is an important factor in the satisfactory performance and life of any V-bell or chain drive. As V-beltswear they seat themselves deeper in the sheave grooves. This seating, along with belt stretch lessens the initial tension. Theresult is slippage and loss in horsepower capacity unless some form of take up is used to restore and maintain original tension. As chains wear they elongate, which result in sway or slap and increased vibration. This puts undue shock into the drive andincreases bearing wear. Both V-belt and chain drives wear at an increased rate if allowed to run with insufficient tension, Adjustment of center distanceis the best method of maintaining proper tension. In cases of fixed centers, drive tighteners and idlers provide the necessarymeans of take up. Drive tighteners also provide a means of obtaining extra belt or chain wrap frequently needed for extremelyhigh ratios, multiple shaft or serpentine drives. Drive tighteners and idlers are listed in the SDP catalog.

1. Too much tension in a drive causes excessive belt, chain and bearing wear.2. Too little tension in a drive allows belt slippage or chain vibration, resulting in loss of power and additional wear.3. All idlers should be used on the slack side of the drive.4. V-belt drive idlers should be used on the inside of the belt. Allowance should be made for horsepower loss due to the

reduced arc of contact.5. Where necessary to use V-belt drive idlers on the outside of the belt, the reverse bending will reduce belt life.6. Flat face pulley idlers can be used on either inside or outside of the belt.7. Sprocket idlers should be used on the outside of chain drives, and with at least three teeth engaged in the chain.8. Idlers used on the inside of a drive should be located approximately 1/3 of the center distance from the large sheave,

pulley or sprocket.9. Idlers used on the outside of a drive should be located approximately 1/3 of the center distance from the small sheave,

pulley or sprocket.

484

Table of Contents

B. SYNCHRONOUS BELT DRIVES

1.0 INTRODUCTION 480

2.0 SYNCHRONOUS BELTS 486

2.12.22.32.42.5

Belt ConstructionCharacteristics of Reinforcing FibersCord Twist and its Effect on the DriveFactors Contributing to Side TravelCharacteristics of Belt Body Materials

488489491493493

3.0 BELT TOOTH CONFIGURATIONS 493

4.0 PULLEY PITCH AND OUTSIDE DIAMETERS 498

5.0 DESIGN AND INSTALLATION SUGGESTIONS 498

6.0 STANDARDS APPLICABLE TO BELTS 500

7.0 PULLEY DIMENSIONS AND TOLERANCES 503

8.0 COMPARATIVE SIZES OF DIFFERENT PITCH PULLEYS 516

9.0 DRIVE RATIO TABLES 522

10.0 CENTER DISTANCE FACTOR TABLES 529

11.0 DESIGN OF BELT DRIVES

11.111.211.3 11.411.511.6

Design Based on Horsepower Design Based on Allowable Torque for Optimum Life Drive Selection Procedure−−Based on Horsepower Drive Selection Procedure−−Based on Allowable Torque Example of Drive Design−−Based on Horsepower Example of Drive Design−−Based on Allowable Torque in Expectation of Optimum Life

599600600621622 623

485

B. SYNCHRONOUS BELT DRIVES1.0 INTRODUCTIONSynchronous drives represent an important category of drives. Characteristically these drives employ the positive engagement oftwo sets of meshing teeth. Hence, they do not slip and there is no relative motion between the two elements in mesh. Due to this feature different parts of the drive will maintain a constant speed ratio or even a permanent relative position. This isextremely important in applications such as automatic machinery in which a definite motion sequence and or indexing is involved. The positive nature of these drives makes them capable of transmitting large torques and withstanding large accelerations. Synchronous drives can be divided into the following categories: 1. Belt Drives* 2. Chain Drives 3. Gear Drives This section of the Handbook will encompass the first category only. The other categories have been dealt with in precedingchapters. Since it is directed primarily towards designers of smaller mechanisms and machinery we shall limit the tabulation ofdata to miniature and light belts and pulleys. Inasmuch as there is relatively limited information readily available for these belts,we have found it necessary to computerize the center-distance tables in order to cover the range of belts and pulleys carried inour catalog. Belt drives are particularly useful in applications where layout flexibility is important. They enable the designer to placecomponents in more advantageous locations at larger distances without paying a price penalty. Motors, which are usually thelargest heat source, can be placed away from the rest of the mechanism. Achieving this with a gear train would represent anexpensive solution.

2.0 SYNCHRONOUS BELTS

Synchronous belts, commonly known as timing belts, are basically flat belts with a series of evenly spaced teeth on the insidecircumference, thereby combining the advantages of the flat belt with the positive grip features of chains and gears. There is no slippage or creep as with plain flat belts. Required belt tension is low, therefore producing very small bearing loads.Synchronous belts wilt not stretch and do not require lubrication. Speed Is transmitted uniformly because there is no chordal riseand fall of the pitch line as in the case of roller chains. The tooth profile of commonly known synchronous belts is of trapezoidal shape with sides being straight lines which generatean involute, similar to that of a spur gear tooth. As a result the profile of the pulley teeth are involute. Unlike the spur gearhowever, the outside diameter of a timing pulley is smaller than its pitch diameter, thus creating an imaginary pitch diameterwhich is larger than the pulley itself. This is illustrated in Figure 1. Backlash between pulley and belt teeth is negligible. Recently a new curvilinear tooth profile was developed which has many desirable and superior properties. Belts and pulleys ofthis construction have features protected by US. Patent Number 3,756,091. Advantages of this type of drive are as follows: • Proportionally deeper tooth; hence tooth jumping or loss of relative position is less probable.__________* Credit for portions of this technical section are given to: Uniroyal Sales Engineering Dept., Rubber Manufacturers Association(RMA), International Organization for Standardization (ISO).

486

• Lighter construction, with correspondingly smaller centrifugal loss. • Smaller unit pressure on the tooth since area of contact is larger.

• Greater shear strength due to larger tooth cross section. • Lower cost since a narrower belt will handle larger load. • Energy efficient, particularly if replacing a "V" belt drive which incurs energy losses due to slippage. • Installation tension is small, therefore, light bearing loads.

On Figure 2 the photoelastic pattern shows the stress distribution within teeth of different geometry. There is a definite stressconcentration near the root of the trapezoidal belt tooth, with very low strains elsewhere. For the Curvilinear tooth there is auniform, nearly constant strain distribution across the belt. The load is largest in the direction of the tension member to which itis transferred.

487

Because of their superior load carrying capabilities the curvilinear belts are marketed under the name of HTD® * drives. This isan abbreviation of High Torque Drives, and they are usually offered in 8mm and 14mm pitches. The 3mm and 5mm curvilinearbelts and pulleys will be referred to as Mini-HTD® * or TRUE METRIC® **, These drives are not necessarily used for high-torqueapplications, but their other superior properties make them an excellent choice for small drives. Timing belts are also available in double-sided designs, which offer an infinite number of new design possibilities on computerequipment, business machines, office equipment, textile machines and similar light-duty applications. Belts with driving teeth onboth sides, make it possible to change the direction of rotation of one or more synchronized pulleys with only one belt. The insideand outside teeth are identical as to size and pitch and operate on standard pitch-diameter pulleys.Double-sided timing belts have teeth on both sides as shown in Figure 3. The outside teeth do not have nylon facing; hence, thehorsepower rating of the outside teeth is only 45% of the total load

For example: assuming the drive pulley and bet are capableof transmitting 1 horsepower, 0.55 hp can be transmittedfrom the inside teeth to pulley (A) and 0.45 hp can betransmitted by the outside teeth to pulley (B) for a total of 1hp the rated capacity of the driver pulley.

2.1 Belt Construction

The load-carrying element of the belts are the tension members built into the belts (see Figure 4 These tension members can bemade of: a) Spirally wound steel wire b) Wound glass fibers c) Polyester cords d) Kevlar The tension members are embedded in neoprene or polyurethane. The neoprene teeth are protected by a nylon fabric facingwhich makes them wear resistant.

*Trademark of UNIROYAL, lnc.**Registered Trademark of SDP

488

The materials used for tension members as well as for the body of the belt itself have a profound effect on the life and usefulnessof the drive for a particular application. We will limit ourselves to a description of characteristics of the materials used and leavethe choice of the materials to the designer, who is best equipped to make the proper selection or compromise.

2.2 Characteristics of Reinforcing Fibers

POLYESTER

Tensile Strength 160,000 lbs/ln2Elongation at break 14.0%Modulus (approx.) 2,000,000 lbs/ln2

One of the main advantages of polyester cord over higher tensile cords is the lower modulus of polyester, enabling the belt torotate smoothly over small diameter pulleys. Also, the elastic properties of the material enable it to absorb shock and dampenvibration. In more and more equipment, stepping motors are being used. Polyester belts have proven far superior to fiber glass or Kevlarreinforced belts in these applications. High-speed applications with small pulleys are best served by polyester belts under low load.

KEVLAR

Tensile Strength 400000 lbs/ln2Elongation at break 2.5%Modulus 18,000,000 lbs/ln2

High tensile strength and low elongation make this material very suitable for timing belt applications. Kevlar has excellent shockresistance and high load capability.

FIBER GLASS

Tensile strength 350,000 lbs/ln2Elongation at break 2.5 - 3.5%Modulus 10,000,000 lbs/ln2

The most important advantages are: 1.) High strength 2.) Low elongation or stretch 3.) Excellent dimensional stability 4.) Excellent chemical resistance 4.) Absence of creep, 100% elongation recovery

Disadvantages: 1.) High modulus (difficult to bend) 2.) Brittleness of glass. Improper handling or installation can cause permament damage. 3.) Poor shock resistance. No shock absorbing quality when used in timing belts.

Additional characteristics of tension members and their effect on the drive design are shown in tabulated form in Table 1.

489

490

2.3 Cord Twist and Its Effect on the Drive

There is a specific reason for not applying the yarn directly in the form of untwisted filaments around the mold. If the filamentwould be applied continuously the top and bottom of the belt body would be prevented from being properly joined, andseparation could result. See Figure 5.1.

Two strands each composed of several filaments are twisted around each other, thus forming a cord which is subsequentlywound in a helical spiral around the mold creating a space between subsequent layers, which corresponds to the step of the helix.The two strands however can be twisted two ways, to create an "S" or a "Z" twist construction. See Figure 5.2.

The "S" twist is obtained if we visualize the two strands being held stationary with our left hand on one end, while a clockwiserotation is imparted by our right hand to the two strands, thus creating a twisted cord. The "Z" twist is obtained similarly if a counterclockwise rotation is imparted to the two strands.

491

Different type of cord twist will cause side thrust in opposite direction. The "S" twist will cause a lateral force direction whichwill obey the "Right Hand" rule as shown in Figure 5.3.

A "Z" type cord twist will produce a direction of lateral force opposite to that of "S" cord. Therefore in order to produce a beltwith minimum lateral force standard belts are usually made with "S" and "Z" twist construction, in which alternate cordscomposed of strands twisted in opposite directions are wound in the belt. This is illustrated in Figure 6.

The lay of the cord is standard as shown on Figure 6, and It is wound from left to right with the cord being fed under the mold.The smaller the mold diameter and the fewer the strands of cord per inch, the greater will be the helix angle, and the greater thetendency of the lay of the cord to make the belt move to one side. In general a standard belt of S" and "Z" construction, as shown in Figure 6r will have a slight tendency to behave as apredominantly "S" twist belt, and will obey the "Right Hand" rule accordingly.

492

2.4 Factors Contributing To Side Travel

The pulleys in a flat belt drive are crowned to keep the belt running true. Since crowned pulleys are not suitable for a timing belt,the belt will always track to one side. Factors contributing to this condition include:

a) lN THE DRIVE

1. Misalignment: A belt (any belt any construction) will normally climb to the high (or tight) side.2. Tensioning: In general, lateral travel can be altered or modified by changing tension.3. Location of Plane: Vertical drives have a greater tendency to move laterally due to gravity.4. Belt width greater than O.D. of Pulley: This condition creates an abnormal degree of lateral travel.

5. Belt length: The greater the ratio of length/width of the belt, the less the tendency to movelaterally.

b) lN THE BELT 1. Direction of the lay of the cords in the belt (Figure 6.) 2. Twist of the strands In the cord (Figure 5.2.)

2.5 Characteristics of Belt Body Materials

Basic characteristics of the three most often used materials are shown on Table 1a. The tabulated characteristics give rise tofollowing assessment of these materials:

NATURAL RUBBER • High resilience, excellent compression set, good molding properties • High coefficient of friction, does not yield good ground finish • High tear strength, low crack growth • Can withstand low temperatures • Poor oil and solvent resistance; usable for ketones and alcohol • Ozone attacks rubber, but retardants can be addedNEOPRENER®

• High resilience, good compression set • Flame resistant • Aging good with some natural ozone resistance • Oil and solvent resistance fairPOLYURETHANE • Excellent wear resistance, poor compression set • Low coefficient of friction • Oil and ozone resistance good • Low-temperature flexibility good • Not suitable for high temperatures

3.0 BELT TOOTH CONFIGURATIONS

There are several belt tooth configurations (Figure 7, Table 2) which are the result of different patented features, marketing andproduction considerations.

493

Inspection of Table 2 brings forth the following favorable characteristics of the 3mm an 5mm TRUE METRIC® belts:

• The jacket size (E) of the 3mm and 5mm belts is the same as the one used on the 3/8 pitch L belts.

• The reinforcing cord of the 3mm pitch belts is of the same diameter (.023) and winding density (26/inch) as the cord inthe 1/5" pitch XL and 3/8" pitch L belts.

• The reinforcing cord of the 5mm pitch belts is nearly of the same diameter (.042 vs .050) as the one used in the 1/2"pitch H belts.

• The allowable working tension of the 3mm pitch belt is (60 lbs/inch) higher than that of the 3/8" pitch belts (55 lbs/inch).

• The allowable working tension of the 5mm pitch belt is (100 lbs/inch) nearly double that of the 3/8" pitch belts (55lbs/inch).

• The relatively low centrifugal loss constants, Kc, indicate that the 3mm and 5mm belts are lighter; hence, centrifugaltension losses are smaller.

If the belt width is less than 1" the belt torque carrying capacity will decrease in a disproportionate manner as shown in Table3. This is due to inefficiency of the tension members at the edge of the belt. The narrower the belt the smaller the ratio of fullload carrying tension cords relative to the total number of cords in the belt. When desing data for 1" wide belts is applied to narrower belts a Width Factor multiplier is used which takes into account thedata shown in Table 3.

TABLE 1ACOMPARISON OF DIFFERENT BELT BODY MATERIALS*

Common Name Natural Rubber Neoprene® Urethene, PolyurethaneChemical Definition Polyisoprene Polychloroprene Polyester/Polyether UrethaneDurometer Range (Shore A) 20 - 100 20 - 95 35 - 100Tensile Range (P.S.I.) 500 - 3500 500 - 3000 500 - 6000Elongation (Max. % 700 600 750Compression Set Excellent Good PoorResilience - Rebound Excellent Excellent ExcellentAbrasion Resistance Excellent Excellent ExcellentTear Resistance Excellent Good ExcellentSolvent Resistance Poor Fair PoorOil Resistance Poor Fair GoodLow Temperature Usage (Fº) -20º to -60º +10º to -50º -10º to -30ºHigh Temperature Usage (Fº) to 175º to 250º to 175ºAging Weather - Sunlight Poor Good ExcellentAdhesion to Metals Excellent Good to Excellent Fair to Good

*Courtesy of Robinson Rubber Products

494

495

TABLE 3ALLOWABLE WORKING TENSION FOR DIFFERENT BELT WIDTHS

(Ta in LBS. not corrected for centrifugal force loss)

TABLE 4MINIMUM PULLEY DIAMETERS

*Smaller pulleys then shown under "SuggestedMinimum" may be used if a corresponding reduction inbelt life is satisfactory. Use of pulleys smaller thanthose shown will be at customers own responsibilityfor performance and belt life.

496

For the sake of completeness two additional belt profiles are shown on Figure 8. These belts are used in Europe and aresometimes found on machinery imported from Europe and Japan. They are not produced in the USA and are not covered by RMAstandards. These belts are made of polyurethane and steel is usually used as the tension member.

Pulley and belt geometry as indicated in Figure 1 shows reference to a Pitch Circle which is larger than the pulley itself. Its sizeis determined by the relationship

where P is the belt tooth spacing (pitch) and N is the number of teeth on the pulley. The reinforcing cord centerline will coincidewith the pulley pitch diameter while the belt is in contact with the pulley. At the same time the outside diameter of the pulley willbe in contact with the bottom of the belt tooth. Hence, the distance "U" between the reinforcing cord centerline and the bottom ofthe belt tooth will determine the outside diameters of pulleys for different pitches. See Table 5.

TABLE 5 BASIC BELT DIMENSIONS

Distance from Pitch Line to Belt Tooth Bottom

"U"

Common

Description

Pulley OD = pd −− 2U

.010 inches

.007 inches

.010 inches

.010 inches.015 inches

0.5 millimeters1.0 millimeter.015 inches.0225 inches.027 inches

Minipitch 0.080 MXL 40 DP

Minipitch .0816 1/5" XL3/8" L

"T5" (5mm)"T10" (10mm)3mm Mini HTD5mm Mini HTD

8mm HID

497

Due to the particular geometry of the 8mm HTD belts, some corrections are needed for small size pulleys only. Hence, consultpulley specification tables given later in this text pertaining to the 8mm HTD pulley.

4.0 PULLEY PITCH AND OUTSIDE DIAMETERS

As previously noted the pitch and the number of teeth will determine the pitch diameter of the pulley, whereas its outsidediameter will depend on the "U dimension (distance from tooth bottom to centerline of cord) as shown on Table 5. In order to provide fast reference the following tables show pitch and outside diameters of different pitch pulleys:

Table 15: 0.080 Minipitch and 0.0816-40 Diametral PitchTable 16: 1/5" - XL pitchTable 17: 3/8" - L pitchTable 18: 3mm pitchTable 19: 5mm pitchTable 20: 8mm pitch

5.0 DESIGN AND INSTALLATION SUGGESTIONS

There are some general guidelines, which are applicable to all timing belts including miniature and double sided belts:

1. Drives should always be designed with ample reserve horsepower capacity. Use of overload service factors is important. Beltsshould be rated at only 1/15th of their respective ultimate strength.

2. For miniature pitch belts the smallest recommended pulley will have 10 teeth. For other pitches Table 4 should be used.3. The pulley diameter should never be smaller than the width of the belt.4. Belts with Fibrex-glass fiber tension members should not be subjected to sharp bends or rough handling, since this could

cause breakage of the fibers.5. In order to deliver the rated horsepower, a belt must have six or more teeth in mesh with the grooves of the smaller pulley.

The number of teeth in mesh may be obtained by formula given in Design Theory Section. The shear strength of a singletooth is only a fraction of the belt break strength.

6. Because of a slight side thrust of synchronous belts in motion, at least one pulley in the drive must be flanged. When thecenter distance between the shafts is 8 or more times the diameter of the smaller pulley, or when the drive is operating onvertical shafts, both pulleys should be flanged.

7. Belt surface speed should not exceed 5500 feet per minute for larger pitch belts and 10,000 feet per minute for minipitchbelts. For the TRUE METRIC® belts a speed of 6500 feet per minute (33.02 m/s) is permitted.

8. Belts are in general rated to yield a minimum of 3000 hours of useful life if all instructions are properly followed.9. Belt drives are inherently efficient. It can be assumed that the efficiency of a synchronous belt drive is greater than 95%.

498

10. Belt drives are usually a source of noise. The frequency of the noise level increases proportionally with the belt speed. Thehigher the initial belt tension, the greater the noise level. The belt teeth entering the pulleys at high speed act as acompressor and this creates noise. Some noise is the result of a belt rubbing against the flange, which in turn may be theresult of the shafts not being parallel.

11. If the drive is part of a sensitive acoustical or electronics sensing or recording device it is recommended that the backsurfaces of the belt be ground to assure absolutely uniform belt thickness.

12. For some applications no backlash between the driving and driven shaft is permitted. For these cases special profile pulleyscan be produced without any clearance between the belt tooth and pulley. This may shorten the belt life but it eliminatesbacklash.

13. Synchronous belts are often driven by stepping motors. These drives are subjected to continuous and large accelerationsand decelerations. If the belt reinforcing fiber, i.e. tension member, as well as the belt material, have high tensile strengthand no elongation, the belt will not be instrumental in absorbing the shock loads. This will result in sheared belt teeth.Therefore, take this into account when the size of the smallest pulley and the materials for the belt and tension memberare selected.

14. The choice of the pulley material (metal vs. plastic) is a matter of price, desired precision, inertia, color, magneticproperties, and above all personal preference based on experiences. Plastic pulleys with metal inserts or metal hubsrepresent a good compromise.

The following precaution should be taken when installing all timing belt drives: 1. Timing belt installation should be a snug fit, neither too tight nor too loose. The positive grip of the belt eliminates the need

for high initial tension. Consequently, a belt, when installed with a snug fit (that is, not too taut), assures longer life, lessbearing wear and quieter operation. Preloading (often the cause of premature failure) is not necessary.

When torque is unusually high, a loose belt may "jump teeth" on starting. in such a case, the tension should be increasedgradually until satisfactory operation is attained. A good rule of thumb for installation tension is as shown in Figure 9. Thecorresponding tensioning force is shown in Table 6. For widths other than shown increase force proportionally to the beltwidth.

499

2. Be sure that shafts are parallel and pulleys are in alignment. On a long center drive, it is sometimes advisable to offset thedriven pulley to compensate, because of the tendency for the belt to run against one flange.

3. On a long center drive, it is imperative that belt sag is not large enough to permit teeth on the slack side to engage theteeth on the tight side.

4. It is important that the frame supporting the pulleys be rigid at all times. A non-rigid frame causes variation in centerdistance and resulting belt slackness. This in turn, can lead. to jumping of teeth −− especially under starting load with shaftmisalignment.

5. Although belt tension requires little attention after initial installation, provision should be made for some center distanceadjustment for ease in installing and removing belts. Do not force belt over flange of pulley.

6. Idlers, either of the inside or outside type, are not recommended and should not be used except for power take-off orfunctional use. When an idler is necessary, it should be on the slack side of the belt. Inside idlers must be grooved unlesstheir diameters are greater than an equivalent 40-goove pulley. Flat idlers must not be crowned (use edge flanges). Idlerdiameters must exceed the smallest diameter drive pulley. Idler arc of contact should be held to a minimum.

TABLE 6 BELT TENSIONING FORCE

PITCH WIDTH FORCE.080 (MXL) .080 (MXL)

1/8"1/4"

1 oz2 oz

1/5 (XL)1/5 (XL)

1/4"3/8"

2 1/2 oz3 1/2 oz

3/8 (L)3/8 (L)3/8 (L)

1/2"3/4"1"

7 oz11 oz1 lb

PITCH WIDTH FORCE3mm3mm3mm

6mm9mm15mm

4 1/2-9 oz7 1/2-15 oz

7/8-1 3/4 lbs5mm5mm5mm

9mm15mm25mm

1-2 lbs2-4 lbs

3 1/2-7 lbs8mm (HTD) 20mm 4 lbs

6.0 STANDARDS APPLICABLE TO BELTS

Different belt tooth configurations are shown on Figure 7 and their characteristics are described in Table 2. Since synchronousbelts are manufactured by several manufacturers, each has established individual standards. Subsequently the following generalstandards have been published:

1) Specifications by the Rubber Manufacturers Association for Drives using Synchronous Belts (IP-24-1978).2) Synchronous Belt Drives −− Belts (ISO 5296-1978), specification by the International Organization for Standardization.

Based on these, as well as standards developed by belt manufacturers, the following information is presented in this Handbook:

Total measuring tensionStandard belt widths and tolerancesPitch length tolerances

Table 7Table 8aTable 8b

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Length Measurement

The pitch length of a synchronous belt is determined by placing the belt on a measuring fixture comprising two pulleys of equaldiameter, of applying tension and of measuring the center distance between the two pulleys. One of the pulleys is fixed inposition while the other is movable along a graduated scale. The fixture is shown schematically in Figure 9a. Any pair of equal-diameter pulleys of the proper pitch and manufactured tospecifications may be used for measuring. The measuring tension is given in Table 7. In measuring the length of a synchronous belt, the belt should be rotated at least two revolutions to seat it properly and todivide the tension equally between the two strands. The pitch length is calculated by adding the pitch circumference of one pulley to twice the center distance:

501

TABLE 8A STANDARD BELT WIDTHS AND TOLERANCES

TABLE 8B PITCH LENGTH TOLERANCES

502

503

TABLE 10ISO PITCH TO PITCH TOLERANCES

TABLE 11UNIROYAL PITCH TO PITCH TOLERANCES

504

TABLE 12ISO AXIAL PULLEY RUNOUT

TABLE 13ISO RADIAL PULLEY RUNOUT

TABLE 14UNIROYAL AND ISO PULLEY O.D. TOLERANCES

505

TABLE 15 (Sheet 1 of 3)

506

TABLE 15 (Sheet 2 to 3)

507

TABLE 15(Sheet 3 of 3)

508

TABLE 161/5" (.200) Pitch −− For "XL" Belts

509

TABLE 173/8" (.375) Pitch −− For "L" Belts

510

511

512

513

514

515

8.0 COMPARATIVE SIZES OF DIFFERENT PITCH PULLEYS TABLE 21

PURPOSE: Comparative sizes of different pitch pulleys can be of great help if the envelope dimensions of adrive are being evaluated. Furthermore, if a drive of a certain pitch is replaced by another pitchdrive, these tables show the nearest available pulley sizes which will be of similar diameter.

EXAMPLES: a) A 5:1 ratio. .200" pitch drive using 60 and 12 tooth pulleys is to be replaced by a 3mm pitchdrive.

From the table we see that the 60 tooth .200" pitch pulley has an outside diameter of96.5mm. The closest 3mm pulley has 100 teeth and 94.7mm outside diameter. Accordingly a100/20 teeth 3mm drive will fit within the same envelope as a 60/12 tooth .200" pitch drive.

b) Miniaturize an existing 4:1 .200" pitch drive which uses a 20 and 80 tooth .200" pitch pulleysby using 3mm pitch components.

From the Drive Ratio Tables we choose the smallest available pulleys which will be 10 and 40teeth. Outside diameters will be 8.8 and 37.4mm.

Minimum center distance will be:

Cmin 7.958 x 3mm = 23.87mm.Since the smallest available belt has 48 teeth NB − N1 = 48 − 40 = 8 and N1 − N2 = 40 − 10= 30, from the center distance factor tables we obtain a center distance of 10.381 pitches or31.14mm.

516

TABLE 21517

TABLE 21518

TABLE 21519

In the design of belt drives, we usually know the speed ratio (transmission ratio) and we need to determine pulley sizes, centerdistance and belt length. These quantities are shown in Figure 10, for an open (uncrossed) belt. The transmission tables are designed to facilitate the determination of these quantities. They list the following information:

N1/N2 = the transmission ratio obtained when the larger pulley (N1 teeth) is the input and the smaller pulley (N2teeth) is the output. Given to 3 decimal places.

N2/N1 = the transmission ratio obtained when the larger pulley (N1 teeth) is the output and the smaller pulley (N2teeth) is the input. Given to 3 decimal places. (Note that N1/N2 is the reciprocal of N2/N1).

N1 = number of teeth on larger pulley.N2 = number of teeth on smaller pulley.

N1-N2 = difference between number of teeth on larger and smaller pulleys. This number is useful in center-distancedetermination.

C MIN = The minimum center distance between pulleys for a belt of unit pitch. If the pitch is denoted by p, the actualminimum center distance is the product of C MIN and p. The minimum center distance is determined fromthe condition that at the minimum center distance, the pitch circles of the pulleys can be assumed to touch.This will generally give a satisfactory approximation to the practical minimum center distance. The table isbased on the equation:

At the beginning of the table a list of standard pulley sizes is shown. The smallest pulley has 10 teeth and the largest 156teeth. A standard size will be the most economical. If a nonstandard size is needed, however, please contact Stock Drive Productsfor assistance.

520

The use of the tables is best illustrated by means of examples. Example 1: For a transmission ratio of 1.067 find the number of teeth of the pulleys and the minimum center distance for a beltof 5mm pitch. When the transmission ratio is greater than unity, the larger pulley is the input and the smaller pulley is the output. That is tosay, the transmission ratio is equal to N1/N2. The table is organized in order of increasing values of N1/N2 and decreasing valuesof N2/N1. Referring to the table we find at this value of N1/N2, we have the following entries:

N1/N2 N2/N1 N1 N2 N1-N2 C MIN

1.067 0.938 1632

1530

12

4.9349.868

Hence, there are 2 different pulley combinations for the given transmission ratio of 1.067. For each of these the minimum centerdistance is 5 x (C MIN) in mm. If the smaller pulley were driving, the transmission ratio would have been 0.938. The quantity(N1-N2) is needed in center-distance calculations, as described in the next Section.

Example 2: Given a transmission ratio of 0.680, determine the pulley sizes. Since the transmission ratio is less than one, the smaller pulley is the input and the transmission ratio is given by N2/N1 =0.680. Looking up this ratio in the table, we find N1 = 25, N2 = 17, N1-N2 = 8. In this case only one pulley combination isavailable.

Example 3: Given a driving pulley of 48 teeth and a driven pulley of 19 teeth, find the minimum center distance for a belt pitchof 3mm. The transmission ratio is N1/N2 = 48119 = 2.526. Looking up this ratio in the table, we find C MIN = 10.663. The minimumcenter distance, therefore, is given by 3 x 10.663 or 31.989 mm.

Example 4: Given a transmission ratio of 2.258, find the pulley sizes. Looking through the table, there is no entry at this value of the transmission ratio. The nearest entries are:

N1/N2 N2/N1 N1 N2 N1-N2

2.250 0.444 3672

1632

2040

2.273 0.440 25 11 14

Since the difference between the desired ratio and the nearest available ratios is only about 0.008, it is likely that the 2.250 or2.273 ratios will be acceptable. If this is not the case, however, the design may require review, or a nonstandard pulleycombination may be considered.

521

9.0 DRIVE RATIO TABLES TABLE 22

DEFINITION: Drive Ratio (Transmission Ratio) is the ratio of number of teeth of the input and output pulley. Ifthe input pulley is larger than the output, the Drive Ratio will be larger than one and we have astep-up-drive. If the input pulley is smaller than the output pulley the Drive Ratio will be smallerthan one and we have a step-down-drive.

NOMENCLATURE USED: N1 Number of teeth of large pulleyN2 Number of teeth of small pulleyN1/N2 Step-up Drive ratioN2/N1 Step-down Drive ratioN1-N2 Pulley tooth differential needed for the Center Distance TableCmin Minimum center distance for particular pulley combination expressed in belt pitches

PULLEY SIZESINCLUDED:

10, 11, 12, 13, 14, 15, 16, 17, 18, 19, 20, 22, 24, 25, 28, 30, 32, 36,40, 48, 60, 72, 84, 96, 120, 156

NOTE: These pulley sizes reflect the preferred sizes per ISO Standard 5294 for synchronous beltdrives−−Pulleys (First edition-1979-07-15).

522

TABLE 22

523

TABLE 22

524

TABLE 22

525

l. Nomenclature and basic equations.Figure 11 illustrates the notation involved.The following nomenclature is used:CBLpNBN1N2φ

=======

center distance, inches.belt length, inches = PNB.circumferential pitch of belt, inches.number of teeth on belt = LIP.number of teeth (grooves) on larger pulley.number of teeth (grooves) on smaller pulley.one half angle of wrap on smaller pulley, radians.

θ = =angle between straight portion of belt and line of centers, radians.

R1R2π

===

pitch radius of larger pulley, inches =(N1)p/2π.pitch radius of smaller pulley, inches =(N2)p/2π.3.14159 (ratio of circumference to diameter of circle).

The basic equation for the determination of center distance is: 2Csin φ = L− π(R1 + R2) − (π − 2φ) (R1 − R2) [1]where C cos φ = R1 − R2 [2] These equations can be combined in different ways to yield various equations for the determination of center distance. Wehave found the formulations which follow useful.

526

ll. Exact center distance determination −− unequal pulleys. The exact equation is as follows: C = (1/2)p [(NB − N1) + k(N1 − N2)] [3]

where k = (1/π) [4a]

and φ is determined from: (1/π) (tan φ − φ) = (NB−N1)/(N1−N2) = Q (say) [4b] The value of k varies within the range (1/π, 1/2) depending on the number of teeth on the belt. All angles in equations [3,4]are in radians. The procedure for center distance determination is as follows: 1. Select values of N1, N2(in accordance with desired transmission ratio) and NB. 2. Compute Q = (NB−N1)/(N1−N2). 3. Compute φ by solving equation [4b] numerically. 4. Compute k from equation. [4a]. 5. Compute C from equation [3]. lll. Exact center distance determination −− equal pulleys. For equal pulleys, N1 = N2 and equation [3] becomes

[5]

lV. Approximate center distance determination. Approximate formulas are used when it is desirable to minimize computation time and when an approximate determination ofcenter distance suffices. An alternative to equation [1] for the exact center distance can be shown to be the following:

[6]

where S varies between 0 and 0.1416, depending on the angle of wrap of the smaller pulley. The value of S is given very nearlyby the expression: S = (cos²φ)/12 [7]

527

In the approximate formulas for center distance, it is customary to neglect S and thus to obtain following approximation for C:

[8]

The error in equation [8] depends on the speed ratio and the center distance. The accuracy is greatest for speed ratios close tounity and for large center distances. The accuracy is least at minimum center distance and high transmission ratios. in manycases the accuracy of the approximate formula is acceptable.

V. Number of teeth in mesh (TIM). It is generally recommended that the number of teeth in mesh be not less than 6.The number, TIM, teeth in mesh is given by: TIM = λ N2 [9] where λ = φ /(3.1416) when φ (see equation [4b]) is given in radians (See also the derivation given for TIM in this Handbook).

Vl. Determination of belt size for given pulleys and center distance. Occasionally the center distance of a given installation is prescribed and the belt length is to be determined. For given pitch,number of teeth on pulleys and center distance, the number of teeth on the belt can be found from the equation:

[10]

where the arc sin is given in radians and lies between 0 and π /2. Since NB in general will not be a whole number, the nearestwhole number less than NB can be used, assuming a slight increase in belt tension is not objectionable. An approximate formula can be used to obtain the belt length:

[11]

528

594

Table Number

Located On Page

11a23456789101112131415

490494495496496497500501502503504504505505505506

Table Number

Located On Page

161718192021222324252627282930

509510511513515516522529596601602603604605607

Table Number

Located On Page

31323334353637383940414243

608609610611612613614615616616616621626

Figure Number

Located On Page

1234567

487487488488491492495

Figure Number

Located On Page

89101112131415

497499520526606617618619

595

TABLE 24596

TABLE 24597

598

11.0 DESIGN OF BELT DRIVES

There are basically two approaches to the proper dimensioning of a belt drive, in both methods it is assumed that the loadcarrying ability of the pulley will exceed that of the belt. Therefore, the emphasis will be placed on the proper belt selection. The preoccupation with belt selection and complete neglect of the proper selection of pulley bores−−or shaft sizes−−as well asthe means of fastening the pulley to the shaft may also result in failure of the drive. For this reason this Handbook contains asection on design of shafts and fasteners.

11.1 Design Based on Horsepower

This method takes into account the tensile strength of the belt, after proper allowance is made for losses as a result ofcentrifugal force. The pulley diameter and rpm are treated as in dependent variables.The formula used for this calculation is derived as follows:

hp = 7.933 x pd x rpm (Tt − Tc) x 10-6 [13]

Tc = Kc x pd² x rpm² [14]

For Service Factor SF = 1, Ta = Tt, and for a one inch wide belt the values of Ta and Kc are given for different beltconstructions in Tables 2 and 3. Based on these values using (13) and [14] computerized Horsepower Capacity Tables are compiled and shown on Tables 25thru 29 for different belt constructions. Tension loss per one inch belt width is shown as a function of belt speed in form of a graph on Figure 12. The problem with this method of calculating is twofold:

a) It does not take into account that surface speed of the belt must not exceed 6500 fpm In the case of TRUE METRIC®(10,000 for MXL and 5500 for XL). This limitation has to be superimposed on the table subsequently, (See interruptedlines on Table 25) by means of additional calculations.

b) The life of the belt is not taken Into account at all. It was established empirically that the life of the belt (in hours) willdepend on the following factors:1.2.3.

Number of deformation cycles the belt can withstand.Length of belt used.inverse proportionality to the belt speed.

The number of deformation cycles the belt can withstand will vary depending on the tension load of the belt itself. The higherthe ratio of Ta/Tt is, the higher the number of cycles wilt be. Furthermore, for the same Ta/Tt larger diameter pulleys will yieldhigher number of cycles, hence longer life. To cope with the problem of belt life Table 4 gives minimum pulley diameters, below which a reduced belt life must beanticipated. On Table 25 the dotted line indicates the area beneath which reduced belt life will occur. In general, to increase beltlife horsepower ratings for higher rpm's should be reduced.

599

11.2 Design Based on Allowable Torque for Optimum Life (For HTD® −−TRUE METRIC® drives only)

If all the values and coefficients which govern the belt life are established experimentally. then by calculating simultaneouslybelt life and horsepower, a set of values can be derived for which a somewhat reduced horsepower rating will yield substantiallylonger life. Such evaluations have been made for the TRUE METRIC® 3mm and 5mm belts, and the findings have been put into the formof Tables 30 thru 33 which show variable allowable torque values as a function of rpm and pulley diameter as well. These valuesare given for specific belt widths. The torque values used with these tables must take into account the actual torque requirements derived by experiment orcalculation, and must be multiplied by a service factor given in Table 38. The service factor must be chosen based on actualoperating conditions. We will call this increased torque the Design Torque. By substituting torque values from Tables 30 thru 33 into

[15]

the horsepower Tables 34 thru 37 have been compiled. If these tables are used in lieu of Tables 28 and 29 the resulting beltwidths will have to be increased, but belt life will also be substantially improved. PLEASE NOTE THAT THIS METHOD OF BELT SELECTION BASED ON OPTIMUM LIFE, WAS DEVELOPED BY UNIROYAL INC. ONLYRECENTLY AND IS APPLICABLE ONLY TO THE HID® −−TRUE METRIC® BELTS. THERE ARE NO SIMILAR COMPUTATIONAL MEANSAVAILABLE FOR THE MXL, XL AND L TYPE BELTS THESE HAVE TO BE CHOSEN IN THE CONVENTIONAL WAY BASED ONHORSEPOWER. THE APPLICABLE PROCEDURE IS GIVEN IN THE FURTHER TEXT. If specific belt life values are required for the TRUE METRlC® belts, SDP Application Engineering will be glad to provide these ifthe basic design data is supplied. A minicomputer program exists, and telephone inquiries can be promptly expedited.

11.3 Drive Selection Procedure−−Based on Horsepower

Step 1: Determine design horsepower By definition design horsepower is a multiple of the rated horsepower, which takes into account overload and the specific drivecharacteristics.

Dhp = hp x SF [16]

Service Factor SF is a multiplier always greater than one which is established based or service requirements. Upon determiningthe class of driver per Table 39 the Basic Service Factor per Table 38 is selected. In case of a Speed-up Drive an additional factoris added to SF, whereas per Table 41 unusual conditions will further modify SF. Step 2. Select belt pitch From belt pitch selection graph Fig. 13 select the belt type which will be used. There is usually more than one choice which canbe considered. If a belt of stronger design is used it will result in a narrower belt. In many instances the selection will be madebased on: total cost of drive, compatibility with previous designs, required flexibility of drive belt, depth of engagement betweenbelt & pulley, availability of particular belt length, and finally personal preference of the designer.

600

FOR 1 INCH WIDE BELT■ BASED ON TENSILE STRENGTH■

Note: To obtain HP capacity forwidths other than oneinch use value in table,multiplied by widthfactor.

Exam ple:A one-inch-wide MXLbelt running on a 30 MXLdriver pulley at 7000 RPMhas a capacity of 1.35 HP.To find the HP capacity ofa 3/8-inch-wide MXL belt(6Z16-xxx037) multiply1.35 X 0.32 - 0432 HP per3/8-Inch-wide MXL belt.

•••••••Area beneath dotted line:

This pulley and RPMcan be used only if aCorrespondingreduction in beltservice life isallowable.

−−−Area beneath interruptedline: This pulley andRPM can be used onlyif a simultaneousreduction in beltservice life and torquereduction is takeninto account.

TABLE 25601

FOR 1 INCH WIDE BELT■ BASED ON TENSILE STRENGTH■

Note: To obtain HP capacity for widths other than one inch use value in table, multiplied by width factor.Example: A one-inch-wide XL belt runningona 30 XL driver pulley at 7000 RPM has a capacity of 3.84 HP. To find the HP

capacity of a 3/8-inch-wide XL belt (6Z3-XXXO37) multiply 3.84 X 0.29 = 1.11 HP per 3/8-inch-wide XL belt.

••••••••••Area beneath dotted line: This pulley and RPM can be used only if a corresponding reduction in belt service lifeis allowable.

TABLE 26602

FOR 1 INCH WIDE BELT■ BASED ON TENSILE STRENGTH■

Note: To obtain HP capacity for widths other than one inch use value in table, multiplied by width factor.Example: A one inch wide L belt running on a 22 L driver pulley at 4000 RPM has a capacity of 4.23 HP. To find the HP

capacity of a ½ inch wide L belt (6R4-XXX050) multiply 4.23 X 0.44 1.86 HP per ½ inch wide L belt.

••••••••Area beneath dotted line: This pulley and RPM can be used only if a corresponding reduction in belt service lifeis allowable.

TABLE 27

603

FOR 1 INCH WIDE BELT■ BASED ON TENSILE STRENGTH■

Important Note:This table is computed based ontensile strength similarly totables for 0.080, 0.200 and0.375 inch pitch belts.DO NOT USE FOR DESIGN. USEFOR COMPARISON ONLY.Tables later In this chapter arebased on allowable torque, andthe resulting computationyields optimum belt life.

Note: To obtain HP capacity for widths other than one inch use value in table, multiplied by width factor.

TABLE 28604

FOR 1 INCH WIDE BELT■ BASED ON TENSILE STRENGTH■

Important Note:This table is computedbased on tensile strengthsimilarly to tables for0.080, 0.200 and 0.375 inchpitch belts.DO NOT USE FOR DESIGN.USE FOR COMPARISONONLY.Tables later In this chapterare based on allowabletorque, and the resultingcomputation yieldsoptimum belt life.

Note: To obtain HP capacity for widths other than one inch use value in table, multiplied by width factor.

TABLE 29605

606

TABLE 30 6mm (0.24in.) Wide Belt

607

TABLE 31 9mm (0.35 in.) Wide Belt

608

TABLE 32 9mm(0.35 in.) Wide Belt

609

TABLE 33 15mm (0.59 in.) Wide Belt

610

BASED ON ALLOWABLE TORQUE FOR OPTIMUM LIFETABLE 34 6mm (0.24 in.) Wide Belt

611

BASED ON ALLOWABLE TORQUE FOR OPTIMUM LIFETABLE 35 9mm (0.35 in.) Wide Belt

612

BASED ON ALLOWABLE TORQUE FOR OPTIMUM LIFETABLE 36 9mm (0.35 in.) Wide Belt

613

BELT ON ALLOWABLE TORQUE FOR OPTIMUM LIFETABLE 37 15mm (0.59 in.) Wide Belt

614

615

TABLE 40 SPEED-UP DRIVESSpeed-Up Ratio

RangeAdditional

Factor1 Thru 1.241.25 Thru 1.741.75 Thru 2.492.50 Thru 3.493.50 Thru Over

None0.10.20.30.4

TABLE 41 UNUSUAL CONDITIONS

For 24-hour continuous operation and/or use of an idler, add0.2 to Service Factor. For intermittent or seasonal operation,deduct 0.2 from Service Factor.

616

617

618

Figure 15619

It is worth noting that the 3mm and 5mm TRUE METRIC® belts are capable of carrying loads normally associated with muchcoarser pitch belts. As a result, the use of TRUE METRIC® belts will yield the following advantages: smaller overall size, narrowerbelts, ultimately lower cost.

Step 3. Select pulley combination Use Drive Ratio Tables (Table 22) and choose pulleys depending on the space available; TRUE METRIC® pulley dimensions aregiven in Tables 18 and 19. Check surface speed of smaller pulley BS (fpm) = 0.262 x pd (in) x rpm [17] BS (m/s) = 0.0000524 x pd (mm) x rpm [18]

which should not exceed 6500 f pm for TRUE METRIC® (10,000 for MXL and 5500 for XL). Check compliance with requirement given on Table 4 for minimum pulley diameters.

Step 4. Determine belt length and nominal center distance Choice of longer belt length has the effect of increasing belt life. On the other hand it also increases the envelope dimensions ofthe drive, as well as its costs. It is desirable to choose belt lengths which are available as standard stock items. For calculation of accurate center distances see separate section with formulas and center distance factor-tables (Table 23).

Step 5. Select belt width From the appropriate Belt Horsepower Capacity Ratings (Tables 25 thru 27) and belt width factors calculate the width of thebelt needed to satisfy the calculated Dhp.

Step 6. Obtain actual Service Factor This step is introduced to recheck previous steps

Te is calculated from [20] and Tc from [14] or obtained from graph Figure 12. Ta for the appropriate belt width is taken fromTable 3. If SF is satisfactory, then design is finalized, if not, increase belt width and correspondingly the value of Ta will alsoincrease.

620

Step 7. Verify teeth in mesh (TIM) Number of teeth in mesh is calculated based on [12]. The minimum number of TIM which will satisfy the shear strength of thebelt teeth is 6. Namely, this is the specific relationship between the allowable working tension and shear strength of teeth perunit width of the belt. If TIM is smaller than 6, then the chosen belt width Wt must be increased using correction factor F perTable 34 and an increased belt width Ws must be specified.

TABLE 42

TIM F FACTOR6 15 0.84 0.63 0.42 0.2

11.4 Drive Selection Procedure−−BASED ON ALLOWABLE TORQUE

NOTE: THIS PROCEDURE IS APPLICABLE TO TRUE METRIC® PITCHES ONLY ANDWILL YIELD OPTIMUM LIFE.

Step 1. Determine design torque (peak torque) Torque which is a result of the nominal shaft horsepower can be calculated from [19].

Upon choosing Service Factor SF per Tables 38-41 peak torque can be calculated Design Torque = T x SF [23] Step 2. Select belt pitch Based on Design Torque obtained from [23] and the rpm of smaller pulley choose from Figure 14 the belt type which will beused. Step 3. Select pulley combination From the Design Torque and Tables 30 through 33 which show the allowable torque values of TRUE METRIC® belts choose thesmaller pulley size. Subsequent to determination of the small pulley size use Drive Ratio Tables (Table 22) and choose whenever possible pulleysizes which are available from stock. TRUE METRIC® pulley diameters are shown in Tables 18 and 19. Check surface speed of smaller pulley BS (fpm) = 0.262 x pd (in) x rpm [17] BS (m/s) = 0.0000524 x pd (mm) x rpm [18]which should not exceed 6500 f pm for TRUE METRIC® (10,000 for MXL and 5500 for XL).

621

Check compliance with requirement given on Table 4 for minimum pulley diameters. Use procedure identical with the one described for calculation based on horsepower. See [8] and Table 23.

Step 5. Select belt width Using the Design Torque determined in Step 1 and the smaller pulley's speed and number of grooves, select the proper beltwidth from the appropriate 3mm or 5mm Allowable Pulley Torque Tables 30 through 33.

Step 6. Verify teeth in mesh (TIM) Use procedure identical with the one described for calculation based on horsepower. See [22] and Table 42.

11.5 Example of Drive Design−−BASED ON HORSEPOWER

GIVEN: Rated hp = 0.3, rpm = 1700, class of drive lll, Drive ratio 3:1, used for reciprocating compressor, usage is intermittent.

Step 1: Determine design horsepower. SF from Tables 38 through 41 is chosen asSF = 2.4 - 0.2 = 2.2

From [16] Dhp = hp x SF Dhp = 0.3 x 2.2 = 0.66

Step 2: Select belt pitch From Figure 13, for Dhp = 0.66 rpm = 1700 L type belt of 318" pitch is chosen.

Step 3: Select pulley combination From Drive Ratio Tables (Table 22) and desired overall size of drive the following is chosen:

N1 = 48 D1 = 5.730" N2 = 16 D2 = 1.910" Values of D1 and D2 are obtained from SDP Catalog. A check of belt speed per [17] reveals BS = 0.262 x pd x rpm = 0.262 x1.910 x 1700 = 850.71 (fpm) which is lower than the specified maximum of 5500 fpm. Therefore, no correction is needed. Per Table 4 our choice exceeds the minimum pulley diameter requirement of 1.671".

Step 4: Determine belt length and nominal center distance From the desired overall size of drive and availability of standard belt lengths (see SDP Catalog) the following is chosen:

NB = 86 BL = 32.250" Center distance is obtained from Table 23. NB-N1 = 38 N1-N2 = 32 CDF = 26.509 C = 26.509 x 0.375 = 9.941"

Step 5: Select belt width From Table 27 for ½" wide belt 1700 rpm and 16 teeth the horsepower capacity rating becomes: hp = 1.45 x 0.44 = 0.64

where 1.45 is rating for 1" wide belt and 0.44 the width factor for ½" wide belt. This capacity rating closely approximate the required Dhp 0.66, therefore the ½" width (which is standard) is chosen.

622

Step 6: Obtain actual Service Factor. From [20]

Tc is obtained from Figure 12 for the previously computed velocity of BS = 850.71 (fpm). Tc = 0.4 lbs. Tc can also be computed from [14] Tc = Kc x pd2 x rpm2 where Kc = 38 x 10-9 from Table 2. Tc = 30 x 10-9 x 1.9102 x (1.7 x 103)2 = 400.62 x 10-3 = 0.4 lbs. Ta is obtained from Table 3. Ta = 24 lbs. Therefore from [21]

This ServiceFactor is smaller than the desired 2.2. It is therefore proper to Increase belt width. Next standard belt width will be¾" (consult SDP catalog) and corresponding Ta from Table 3 will be 39 lbs.

It can be noted that the originally obtained true ServiceFactor of 1.99 can also be deemed satisfactory, provided that asomewhat shorter belt life is acceptable.

Step 7: Verify teeth in mesh (TIM) From [12]

which value is satisfactory and no correction is needed.

11.6 Example of Drive Design BASED ON ALLOWABLE TORQUE IN EXPECTATION OF OPTIMUM LIFE

GIVEN: Rated hp = 0.5, rpm = 1750, class of driver lll, Drive ratio approx. 2.5:1, used for reciprocating compressor, usage isintermittent.

Step 1: Determine Design Torque From [19]

623

Result of this calculation can be checked by Power Nomogram Figure 15. Service Factor SF from Tables 38 thru 41 is chosen as SF = 2.4 - 0.2 = 2.2 From [23] Design Torque = T x SF = 18 x 2.2 = 39.6 in lbs Step 2: Select belt pitch From Figure 14 the 5mm belt pitch is chosen based on 39.6 in lbs torque and 1750 rpm Step 3: Select pulley combination Small pulley with 28 teeth is chosen, consequently, based on Tables 19 and 22 N1 = 72 D1 = 4.511 N2 = 28 D2 = 1.754 The drive ratio will be 2.57:1. A check of belt speed 1 reveals BS = 0.262 x pd x rpm = 0.262 x 1.754 x 1750 = 804.21 (fpm)which is lower than the specified maximum of 6500 f pm. Per Table 4 our choice exceeds the minimum pulley diameterrequirement of 1.629".

Step 4: Determine belt length and nominal center distance From the desired overall size of drive and availability of standard belt lengths (see SDP catalog) the following is chosen: NB = 100 BL = 19.685Center distance is obtained from Table 23 NB-N1 = 28 N1-N2 = 44 CDF = 23.970 C = 23.970 x 0.19685 = 4.7185" Step 5: Select belt width From Design Torque value of 39.6 in-lbs and small pulley with 28 teeth and 1750 rpm, using Allowable Pulley Torque Table 33yields a belt width of 15mm. Step 6: Verify TIM From [12]

resulting value is satsifactory, and no correction is needed.

Step 7: Entering the parameters of the example problem into the SDP minicomputer results in a useful belt life calculation of2383 hrs.

624

Table 43 was constructed to provide comparison of different belt constructions using identical RPM and pulleys with the samenumber of teeth. It gives the Horsepower capacity of belts of different widths. It also illustrates the difference of ratings forHTD® −−TRUE METRIC® belts using the design criterion for optimum life, (allowable torque) as opposed to the tensilestrength/Horsepower method. A study of Table 43 reveals the following observations:

• For certain applications the 3mm HTD belt could be used instead of the XL belts

• The 5mm HTD belt compares favorably with the L type belt and therefore could be considered instead of the L belt.

• The influence of belt life factors is apparent in comparing the two methods of determining HP capacity as illustrated by

the dual ratings listed for the 3mm and 5mm HTD belts. Under more severe operating conditions the optimum life(allowable torque) rating is reduced below the value obtained by the tensile strength (H.P. method) for determiningbelt capacity.

• Under certain conditions such as a relatively large pulley and low RPM, calculations based on tensile strength (H.P.

method) will yield lower H.P. capacity than calculations based on allowable torque. This reflects relatively easyoperating conditions for the particular belt chosen. i.e.: 1750 rpm, 30 groove pulley and 3mm pitch belt.

• If a known H.P. is to be transmitted practically any belt pitch could be selected provided the width satisfies thehorsepower requirement. The final determining factor is the desire to select a standard belt and pulley to insure theiravailability. Please consult our catalog for a complete listing of standard belts and pulleys.

625

626