Belt Tension and Shaft Load Calculator

How tight, measured the way you can actually measure it — deflect the span by a sixty-fourth of its length and read the force — and then the part every other belt-tension page leaves out: the resultant shaft load, in magnitude and direction, and what it does to the bearing’s rating life. The capstan relation with the groove angle done properly, the centrifugal term on both sides where it belongs, and ISO 281’s exponent to turn an over-tension into a number.

Last updated 24 Sep 2026

Belt tension and shaft load

Power and tension → the bearing load

Power to transmit (kW)

Small pulley speed (rev/min)

Small pulley pitch diameter (mm)Pitch diameter, not outside diameter. On a V-pulley they differ by about twice the distance from the pitch line to the top of the groove.

Large pulley pitch diameter (mm)

Centre distance (mm)

Belt sectionThe section sets two things this page needs: the mass per metre, which drives the centrifugal tension, and the groove angle, which multiplies the friction. Pick the flat-belt option and the groove term collapses to plain friction.

Included groove angle (degrees)The FULL included angle, not the half angle — 34, 36 and 38 are the common V-pulley values and ISO 4183 varies it with the pitch diameter. 180 means a flat belt.

Belt mass per metre (kg/m)A product property — take it from the belt maker if you have it. It only matters through the centrifugal tension, which is negligible below about 5 m/s and dominant above about 25.

Coefficient of frictionBelt on pulley, before the groove multiplies it. Khurmi’s worked examples use 0.25 and 0.28; 0.3 is a common design value for a rubber belt on cast iron. On a V-belt it barely matters — look at the tension ratio it produces.

Number of belts

Service factorFrom the driven machine, the driver and the daily hours. Work it out on the service factor page linked below — the V-belt family’s table runs 1.0 to 1.8.

How is the tension being set?

Tension margin over the slip limit1.0 puts the belt exactly on the point of slipping, which is not a place to run. Practice is somewhere between about 1.3 and 2. Where your belt maker publishes a deflection force for your section, sheave diameter and speed, that figure supersedes this whole calculation.

Deflection force you measured, one belt (N)Measured at mid-span with the drive stopped, perpendicular to one belt, at the deflection distance below. Gates: on a multiple-belt drive, multiply the force by the number of belts you are actually deflecting — the DISTANCE is not multiplied.

Not a circuit, and drawn at two scales at the top for a reason. The upper span carries its deflection at TRUE scale — one sixty-fourth of the span, which on 596 mm of belt is 9.3 mm and on this drawing is under five pixels. That is what you are being asked to judge with a rule, and it is worth seeing how small it is. Beneath it the same deflection at eight times the scale, with the force arrow, which is the shape the procedure actually has. Then four bars, all to ONE scale in multiples of the useful tension: the tension that does the work, the slack side, the tight side, and what one bearing actually feels. That fourth bar is the whole argument of this page. It is not the biggest of the other three — it is close to their SUM, because the two spans pull in nearly the same direction, and everything above the first bar is there only to stop the belt slipping. Each bar is gated on its own value alone, so nothing here depends on two moving quantities at once.

1.524kNExample

7.5 kW at 1,450 rev/min, 140 mm to 280 mm pulleys on 600 mm centres, three B-section belts, a 1.2 service factor, tensioned 1.5 times the slip limit

From power to the bearing, in one line of vectors

F_e = P·K_s / v · T_c = m·v² · R = e^(μθ / sin(γ/2)) · T₁ = T_i + T_c + F_e/2 · T₂ = T_i + T_c − F_e/2 · shaft load = √(T₁² + T₂² + 2T₁T₂·cos 2α) · sin α = (D − d)/2C · deflection force = T_i/16 at 1/64 in per inch of span

F_e: effective tension — the difference between the two sides, and the only part of the tension that does any work. Everything else is there to stop the belt slipping
T_c = m·v²: centrifugal tension. It adds to both sides equally, so it does no work and reduces the ratio the friction can use. Negligible under about 5 m/s; at 25 m/s it is often the largest term in the belt
γ: the FULL included groove angle — 34, 36 or 38° on a V-pulley, and 180° for a flat belt. It is halved inside the relation, which is the step most retellings of the formula get wrong in one direction or the other
1/sin(γ/2): the wedge amplification: 3.42 at 34°, 3.24 at 36°, 3.07 at 38°. This is what a V-belt is FOR
T_i: installed static tension — what you set with the adjuster and measure with a deflection gauge. The only quantity on this page you actually control
α: the angle each span makes with the line of centres. Small on most drives, which is why the shaft load is usually quoted as simply T₁ + T₂ — and that IS the answer when the pulleys are equal
T_i/16: the deflection force at 1/64 inch of deflection per inch of span. The 16 is 4 × 64 / 16… more usefully: P = 4T_i·h/t, and h/t = 1/64 gives T_i/16

Worked example

7.5 kW at 1,450 rev/min, 140 mm to 280 mm pulleys on 600 mm centres, three B-section belts, a 1.2 service factor, tensioned 1.5 times the slip limit
Belt speed = π × 140 × 1450 / 60000 = 10.63 m/s, and the span is √(600² − 70²) = 595.9 mm, giving 166.6° of wrap on the small pulley
Design power = 7.5 × 1.2 = 9.0 kW, so the effective tension is 9000 / 10.63 = 847 N across three belts, or 282 N each
A 34° groove amplifies the friction by 1/sin 17° = 3.420, so μ of 0.30 behaves like 1.026 and the slip-limit ratio is e^(1.026 × 2.9077) = 19.8. A V-belt does not slip
Centrifugal tension = 0.185 × 10.63² = 20.9 N per belt. The minimum installed tension that will not slip is 282/2 × (19.8+1)/(19.8−1) = 156 N; at 1.5 times that, 234 N
Tight side = 234 + 20.9 + 141 = 396 N; slack side = 114 N. The spans lean 6.70° either side of the line of centres, so the vector sum per belt is √(396² + 114² + 2×396×114×cos 13.40°) = 508 N, and 1,524 N for three belts
So the bearing sees 1.8 times the tension that is doing the work. Tighten to twice the slip limit instead of 1.5 and the shaft load goes to 1,988 N and the ball bearing's rating life to 15% of what it was at the slip limit
To set it: deflect the span by 596/64 = 9.3 mm at mid-span with the drive stopped, and the force should be 234/16 = 14.6 N (1.49 kgf, 3.29 lbf) on one belt, or 44 N if you deflect all three together

Belt sections, and what the groove buys

The masses are typical figures for the section and belong in the input field rather than in a table: take yours from the belt maker, because construction varies. The groove angles are the common values — ISO 4183 varies the angle with the pulley’s pitch diameter, so 34° on a small sheave becomes 38° on a large one in the same section. The three right-hand columns are the point. A 34° groove multiplies the friction by 3.42, which turns a modest μ of 0.30 into an effective 1.03 and a slip-limit tension ratio of about 20 — against 2.4 for the same belt running flat. That is why a V-belt essentially never slips if it is tensioned at all, and why the tension is set by a deflection table rather than by the slip calculation. It is also why over-tensioning is so easy: nothing goes wrong that you can hear. This is a first-pass calculation on an idealised part — uniform section, static load, no stress raisers beyond those stated, and room-temperature material properties. Real parts have fillets, keyways, surface finish and duty cycles that a closed-form answer cannot see.
Section	Top width (mm)	Mass (kg/m)	Included groove angle (°)	Wedge amplification 1/sin(γ/2)	Effective μ at μ = 0.30	Slip-limit ratio at 166.6° of wrap	Centrifugal tension at 10.63 m/s (N)
Z / 10 mm classical	10	0.062	34	3.420	1.026	19.8	7.0
A / 13 mm classical	13	0.105	34	3.420	1.026	19.8	11.9
B / 17 mm classical	17	0.185	34	3.420	1.026	19.8	20.9
C / 22 mm classical	22	0.315	36	3.236	0.971	16.8	35.6
D / 32 mm classical	32	0.630	36	3.236	0.971	16.8	71.2
SPZ narrow wedge	9.7	0.074	34	3.420	1.026	19.8	8.4
SPA narrow wedge	12.7	0.120	34	3.420	1.026	19.8	13.6
SPB narrow wedge	16.3	0.195	34	3.420	1.026	19.8	22.0
SPC narrow wedge	22	0.375	36	3.236	0.971	16.8	42.4
Flat belt (no groove)	—	0.150	180	1.000	0.300	2.4	16.9

The force-deflection method, and the makers’ own rules

Two things about the arithmetic behind this. First, the deflection force is not an empirical number: a span at tension T deflected by h at its midpoint pushes back with 2T·sin(arctan(2h/t)), which for small h is 4Th/t — and at h/t = 1/64 that is exactly T/16. This batch checked the linearisation against the exact expression and it is 0.049% low, so T/16 is the deflection force to four figures. Second, and this is what makes the 1/64 rule work at all: it is a FRACTION of the span, so the same rule gives the same fractional deflection angle whatever the span, and therefore the same relation between force and tension. The rule is dimensionless dressed up as a length. This is a first-pass calculation on an idealised part — uniform section, static load, no stress raisers beyond those stated, and room-temperature material properties. Real parts have fillets, keyways, surface finish and duty cycles that a closed-form answer cannot see.
Step	What to do	Who says so
Span	t = √(C² − ((D − d)/2)²). Not the centre distance — the tangent length between the pulleys.	TB Wood’s Tensioning Drives prints the formula in exactly this form.
Deflection	1/64 inch of deflection per inch of span. In metric that is 15.625 mm per metre; European makers generally round it to 16 mm per metre, which is 2.4% more deflection.	TB Wood’s: “the deflection height (h) is always 1/64″ per inch of span length (t)”. Gates give the same figure.
Where	Perpendicular to ONE belt, at the mid point of the span, with the drive stopped.	TB Wood’s. That the drive is stopped matters: the centrifugal tension is absent, so the force measures the installed static tension and nothing else.
Multiple belts	Multiply the FORCE by the number of belts you are deflecting. Do not multiply the deflection distance.	Gates: “you must multiply the deflection force by the number of strands that you will be deflecting”, and for a PowerBand the distance is “not multiplied by the number of ribs”.
A new belt	Goes on tighter, and drops fast. Recheck within the first 24 hours.	TB Wood’s: initial tension at twice the tabulated minimum, falling “rapidly during the first few hours”, with retensioning between the minimum and 1.5 times it. Gates: “any belt that has not been run more than about 24 hours is considered a new belt”.
Which force	A maker’s tabulated force for your section, sheave diameter and speed beats the one on this page.	TB Wood’s table is indexed by section, small sheave diameter, speed range AND drive ratio. It is also sized for the belt’s RATED power, not for the power you happen to be transmitting, which is why a tabulated figure is usually higher than the one computed here for a lightly loaded drive.

What each extra turn on the adjuster costs — the default drive, tension by tension

Read the two life columns against the shaft-load column and the whole argument about over-tensioning is there. Doubling the installed tension raises the shaft load by about 92% and cuts the ball bearing’s rating life to 15.1% — because ISO 281 makes the basic rating life go as the load raised to the power 3 for a ball bearing and 10/3 for a roller bearing. That exponent is the only thing taken from the standard here and it is doing all the work: a load you cannot feel with your thumb is a life you lose by a factor of three. Note also the slack-side column. At the slip limit the slack side is only 36 N, and a belt that slack will flap on a shock load — which is the honest reason not to run at 1.0 and the reason practice sits somewhere between about 1.3 and 2. This page sizes a part; it does not certify one. Where the answer carries a consequence — a load path, a lifting duty, a pressure boundary, a fastener holding something that can fall — confirm it against the design code that governs the application, and against the manufacturer’s own rating, before relying on it.
Tension margin over the slip limit	Installed tension (N)	Tight side (N)	Slack side (N)	Total shaft load (N)	Shaft load (%)	Ball-bearing life (%)	Roller-bearing life (%)	Deflection force per belt (N)
1.00	156.2	318.2	35.9	1,060	100.0	100.0	100.0	9.76
1.10	171.8	333.8	51.6	1,152	108.7	77.8	75.6	10.74
1.25	195.2	357.2	75.0	1,292	121.9	55.2	51.7	12.20
1.50	234.3	396.3	114.0	1,524	143.8	33.6	29.8	14.64
1.75	273.3	435.3	153.1	1,756	165.7	22.0	18.6	17.08
2.00	312.3	474.4	192.1	1,988	187.6	15.1	12.3	19.52
2.50	390.4	552.4	270.2	2,453	231.5	8.1	6.1	24.40
3.00	468.5	630.5	348.3	2,918	275.3	4.8	3.4	29.28

Every relation this page uses, and where it comes from

Two of these are derived here rather than quoted, and both are flagged as such. The resultant shaft load is the one that matters: it is elementary statics, nothing more than adding two vectors, and yet none of the belt-tensioning material consulted in this batch printed it. That is the gap this page exists to fill. Everything upstream of it — the arc of contact, the capstan relation with the wedge factor, the centrifugal term — is standard textbook material with a named source, and everything downstream of it is one exponent from ISO 281.
Quantity	Expression	Source
Belt speed	v = π·d·n / 60000 m/s (d in mm)	Definition.
Span	t = √(C² − ((D − d)/2)²)	TB Wood’s, printed in this form. Identical to the chain drive’s tangent length — see the chain length calculator.
Arc of contact	θ_small = π − 2·asin((D − d)/2C)	Plain geometry, and the same expression as the chain wrap angle.
Effective tension	F_e = P_design / v	Definition. P_design is the power times the service factor.
Centrifugal tension	T_c = m·v²	Khurmi. The Cambridge DANotes call ρv² “often conveniently if erroneously referred to as the centrifugal tension” — it is a tension that does no work and must be taken off BOTH sides before the friction relation is applied.
Tension ratio	(T₁ − T_c)/(T₂ − T_c) = e^(μθ/sin(γ/2))	Khurmi writes it as 2.3 log(T₁/T₂) = μθ·cosec β with β the groove HALF angle; Cambridge as f ≡ μ∗cosec β. γ here is the FULL included groove angle, so γ/2 is Khurmi’s β. A flat belt is γ = 180°, where sin(90°) = 1 and the expression becomes the plain capstan relation.
Minimum installed tension	T_i,min = (F_e/2)·(R + 1)/(R − 1), R = e^(μ_eff·θ)	Follows from T₁ − T₂ = F_e and the ratio above, with T₁ + T₂ = 2(T_i + T_c). This is the tension at which the belt is exactly on the point of slipping.
Tight and slack sides	T₁ = T_i + T_c + F_e/2, T₂ = T_i + T_c − F_e/2	The belt’s length is fixed, so T₁ + T₂ is set by the installed tension and the speed, and the difference by the power.
Resultant shaft load	R = √(T₁² + T₂² + 2·T₁·T₂·cos 2α), sin α = (D − d)/2C	DERIVED HERE. No source consulted in this batch printed it. The two spans are tangent to both pulleys and symmetric about the line of centres, each inclined to it by α, so the pulley feels the vector sum. It collapses to T₁ + T₂ when the pulleys are equal. Confirmed three ways over 500 random geometries: the closed form, a component sum, and an explicit tangent-point construction.
Direction of the shaft load	atan2((T₁ − T₂)·sin α, (T₁ + T₂)·cos α), off the line of centres	Same derivation. It leans toward the tight side, and the lean is small — a few degrees on a normal drive — which is why the load is usually taken as acting along the line of centres.
Deflection distance	h = t / 64	TB Wood’s and Gates.
Deflection force	P = 4·T_i·h/t = T_i/16 at h/t = 1/64	DERIVED HERE by string statics, and shown to be 0.049% below the exact 2T_i·sin(arctan 2h/t).
Bearing life	L₁₀ ∝ (C/P)³ for a ball bearing, (C/P)^(10/3) for a roller bearing	ISO 281. Only the exponent is used.

The tension is the easy half. The bearing is the other half

Every belt-tensioning page stops at the tension. The bearing does not. Tensioning a belt correctly is the easy half: measure the span, deflect it by a sixty-fourth of its length at mid-point with the drive stopped, and read the force. The hard half is that the tension you set does not go into the belt and stay there. It goes into the shaft, and from the shaft into two bearings, as a resultant that is the vector sum of the tight and slack side tensions — on the default drive here, 1.8 times the tension that is actually doing the work. And bearing life does not go with load; it goes with load cubed. That is why over-tensioning is the commonest cause of premature bearing failure on a belt drive, and why this page does not end at the tension.

The number nobody prints. The two belt spans are tangent to both pulleys, so they are symmetric about the line of centres and each leans away from it by α, where sin α = (D − d)/2C. The pulley therefore feels T₁ along one span and T₂ along the other, and the resultant is √(T₁² + T₂² + 2T₁T₂·cos 2α) — which collapses neatly to T₁ + T₂ when the pulleys are the same size, which is the form everybody quotes. It also has a direction, a few degrees off the line of centres toward the tight side, and that direction is what tells you which way to orient a self-aligning housing. None of the belt-drive material consulted in this batch printed either the magnitude or the direction. It is elementary statics; it is also the whole reason for this page, so it was derived here and then checked three independent ways over 500 random geometries — the closed form, a component sum, and an explicit construction that finds the tangent points on the two circles numerically and adds the vectors.

Why a V-belt essentially never slips, and why that is a problem. The capstan relation for a flat belt is (T₁ − T_c)/(T₂ − T_c) = e^(μθ). Put the belt in a groove and the wedging action multiplies the effective friction by 1/sin(γ/2), γ being the full included groove angle: 3.42 at 34°, 3.24 at 36°, 3.07 at 38°. So a modest μ of 0.30 behaves like 1.03, and over 166° of wrap the belt can hold a tension ratio of about 20 — against about 2.4 for the same belt running flat. The consequence is that slip is almost never the binding constraint on a V-belt drive, which is exactly why makers publish a deflection-force table rather than asking you to compute a slip limit, and exactly why over-tensioning is so easy to do: there is no audible feedback. A belt that is 60% too tight sounds better than one that is 10% too loose.

The centrifugal tension belongs on both sides, and it does no work. T_c = m·v², and it adds equally to the tight and the slack side, so it must be subtracted from both before the friction relation is applied — the Cambridge engineering notes go so far as to call ρv² “often conveniently if erroneously referred to as the centrifugal tension”, because it is a tension in the belt that transmits nothing. Below about 5 m/s you can ignore it. At 25 m/s it is often the largest single term, and Khurmi’s condition for maximum power transmission says the useful limit is reached when it hits one third of the total: past that, running faster transmits LESS power. The page prints its share of the tight-side tension so you can see where you are.

What a new belt needs, and what this page will not do. A new belt is set tighter and loses it fast: TB Wood’s put the initial tension at twice their tabulated minimum, say it falls “rapidly during the first few hours”, and ask for a recheck inside 24 hours, with retensioning between the minimum and 1.5 times it — Gates define a new belt as one that has run less than about 24 hours. Where your maker publishes a deflection force for your section, sheave diameter and speed, use theirs: their table is indexed by four things this page does not ask about, and it is sized for the belt’s RATED power rather than the power you are transmitting, which is why a tabulated figure is usually higher than the one computed here for a lightly loaded drive. And this page will not compute a belt length or a centre distance for a given belt — that is a different calculation and the site already answers it elsewhere. What it will do is take the span, which is what the deflection is measured over, and carry the tension all the way to the bearing. The service factor that starts it off comes from the drive service factor calculator; the same shaft-load question on a chain drive is a different shape, because a chain has no slack-side tension to speak of — see the roller chain selection calculator and the chain length and centre distance calculator.

Frequently asked questions

How tight should a V-belt be?

Tight enough not to slip and no tighter, and the two ends of that range are further apart than people expect. The slip limit on the default drive here is 156 N of installed tension per belt. Practice sits somewhere between about 1.3 and 2 times that: below about 1.3 the slack side is nearly dead and a shock load or a start will let the belt slip and glaze, and above about 2 you are paying for it in bearing life. The correct answer, when you can get it, is the deflection force your belt maker publishes for your section, your small sheave diameter and your speed — their table is indexed by four variables and supersedes this calculation.

How do I measure belt tension without a gauge?

With a rule and a spring balance, by the force-deflection method. Measure the span — the tangent length between the pulleys, which is √(C² − ((D − d)/2)²) and not the centre distance. Deflect it at mid-span by 1/64 inch per inch of span, which is 15.625 mm per metre; European makers usually round that to 16 mm per metre, 2.4% more. Push perpendicular to ONE belt with the drive stopped and read the force. On the default drive that is 9.3 mm of deflection and 14.6 N of force. If you deflect several belts at once, multiply the force by the number of belts — Gates’ rule — and do not multiply the distance.

Why is the shaft load so much bigger than the useful tension?

Because only the DIFFERENCE between the two side tensions does any work, and the bearing feels something close to the SUM. On the default drive the effective tension is 847 N and the resultant shaft load is 1,524 N, a factor of 1.8. Everything above the effective tension is there to stop the belt slipping, and all of it is carried by the bearings. That ratio is worst on a lightly loaded, heavily tensioned drive, which is the commonest way a belt drive is set up: three belts where two would do, tightened by feel.

Why does over-tensioning kill bearings?

Because of one exponent. ISO 281 makes a bearing’s basic rating life go as the load raised to the power 3 for a ball bearing and 10/3 for a roller bearing. So a 30% over-tension, which raises the shaft load by roughly 30%, leaves about 45% of the rating life; doubling the tension leaves about an eighth. Nothing about the drive sounds or looks different — a V-belt has so much friction reserve that it will not squeal to tell you — and the bearing fails eighteen months later, where nobody connects it to the adjuster. The table on this page prints the life against the tension so the trade is visible before the spanner goes on.

What is the groove angle for, and what happens with a flat belt?

The groove wedges the belt, so a given belt tension presses harder on the pulley faces than it would on a flat rim, and the effective friction coefficient becomes μ ÷ sin(γ/2) with γ the full included groove angle. At 34° that is a multiplier of 3.42. Over 166° of wrap it turns a tension ratio of about 2.4 into about 20. Select the flat-belt option on this page and the groove term collapses — sin(90°) is 1 — and you can see the whole difference in the results. Note that γ is the FULL angle: Khurmi’s relation is written with the HALF angle β, so μθ·cosec β there is μθ/sin(γ/2) here. Getting that factor of two wrong changes the tension ratio by a large factor in either direction.

Does the centrifugal tension matter?

Below about 5 m/s, no — it is well under a per cent of the tight-side tension. Above about 20 m/s it is one of the largest terms in the belt, and it is worse than useless: T_c = m·v² adds equally to both sides, so it raises the tension the belt has to carry without raising the difference that does the work. Khurmi’s condition for maximum power says the useful limit is where the centrifugal tension reaches a third of the total; beyond that, a faster belt transmits less power. The page prints T_c and its share of the tight side. It needs the belt’s mass per metre, which is a product property — the defaults per section are typical figures and your belt maker’s number is better.

Does this page give me the belt length?

No, deliberately. It computes the SPAN — the tangent length between the pulleys — because that is what the deflection is measured over, and a span is not a belt length. Belt length and centre distance are a separate calculation and this site already answers it on its own page, so duplicating it here would put two pages in competition for the same question. What you get here is the other half: what the tension you set does to the shaft and the bearings.

Where does the resultant shaft load formula come from?

It was derived for this page, because none of the sources consulted printed it. The two spans are tangent to both pulleys, so they are symmetric about the line of centres and each is inclined to it by α with sin α = (D − d)/2C. The pulley feels T₁ along one and T₂ along the other, so the resultant is √(T₁² + T₂² + 2T₁T₂·cos 2α) and its direction leans toward the tight side by atan2((T₁ − T₂)sin α, (T₁ + T₂)cos α). Three checks were run over 500 random geometries: the closed form against a component sum, and both against an explicit construction that locates the tangent points on the two circles numerically and adds the vectors. All three agree to twelve figures, and with equal pulleys all three give T₁ + T₂, which is the textbook special case.

Related calculators

References

TB Wood’s (Altra Industrial Motion). Tensioning Drives, catalogue section BEV-1, P-1686-TBW. “The deflection height (h) is always 1/64″ per inch of span length (t)”, with the span itself “t = √(C² – ((D–d)/2)²)”; the force applied “perpendicular to any ONE of the belts at the mid point of the span”; initial tension at twice the tabulated minimum, falling “rapidly during the first few hours”, with retensioning between the minimum and 1.5 times it; and, on an older drive, excessive tension leading to “excessive bearing load or excessive shaft deflection”.
Gates Corporation. Design Flex force/deflection tensioning (Gates Belts, Hoses and Applications). Confirms the same “1/64 of an inch per inch of span length”, and adds two rules this page uses: on a multiple-belt drive “you must multiply the deflection force by the number of strands that you will be deflecting”, and “any belt that has not been run more than about 24 hours is considered a new belt”.
R.S. Khurmi and J.K. Gupta, A Textbook of Machine Design, chapter 21 Chain Drives and chapter 20 V-Belt and Rope Drives (the chapters as distributed by Al-Mustansiriyah University and Al-Mustaqbal University). Source for Table 21.2, factor of safety for bush roller chain against the smaller sprocket’s speed; for the service factor as K1 × K2 × K3 (load, lubrication, hours); for the recommended teeth on the smaller sprocket against velocity ratio; for the V-belt tension ratio 2.3 log(T1/T2) = μθ cosec β with β the groove HALF angle and a groove angle of 32° to 38°; for the centrifugal tension Tc = mv²; and for the maximum-power condition Tc = T/3.
University of Cambridge, Department of Engineering. DANotes: V-belt drives — Kinetics and fatigue. Gives the tension ratio with the centrifugal term separated, “(Fmax − ρv²)/(Fmin − ρv²) ≤ e^(fθ)min”, and the effective friction coefficient “f ≡ μ∗cosec β” with β the groove half angle. It also calls ρv² “often conveniently if erroneously referred to as the centrifugal tension”, which is why this page subtracts it from both sides rather than adding it to the load.
ISO 281:2007, Rolling bearings — Dynamic load ratings and rating life. Cited by number for the life exponent only: basic rating life goes as (C/P) raised to 3 for a ball bearing and 10/3 for a roller bearing. That exponent is what turns a 30% over-tension into a bearing-life figure on this page, and it is the only thing taken from the standard.
BestORQ. V-Belt Quick Selection Guide. The V-belt service factor table used here: four driven-machine groups with their named examples, driving Group 1 (normal-torque AC, shunt-wound DC, multiple-cylinder engines) against Group 2 (high-torque and single-phase AC, series and compound-wound DC, single-cylinder engines), and three service bands — intermittent 3–5 hours, normal 6–16 hours, continuous 16–24 hours or a start above 200% of rated load. Also the idler additions: inside slack side 0.0, outside slack side 0.1, inside tight side 0.1, outside tight side 0.2.