Gear Tooth Bending Stress (Lewis)
Calculate the bending stress at a gear tooth root by the Lewis equation, σ = F_t ÷ (b·m·Y), from the tangential force F_t (N), the tooth face width b (mm), the module m (mm) and the Lewis form factor Y (dimensionless, a function of tooth count). Lewis's 1892 equation was the first rational treatment of gear-tooth strength and is still the basis of BENDING design. It models the tooth as a cantilever beam fixed at the root: the tangential force transmitted between teeth (from the torque) creates a bending moment that tends to break the tooth at the root — the catastrophic failure where a tooth cracks and snaps. The form factor Y accounts for tooth geometry (gears with more teeth are 'fatter' at the root and stronger, higher Y). The computed stress is compared with the material's bending fatigue strength (with safety factors), since gears endure millions of cycles. The basic Lewis formula is then refined by the AGMA standard with stress-concentration, dynamic, load-distribution and surface-condition factors. It is one of the two fundamental gear design criteria (the other is contact stress). Enter the tangential force, face width, module and Lewis form factor.
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Gear tooth bending stress (Lewis)
The Lewis equation (1892), σ = F_t ÷ (b·m·Y), gives the bending stress at the root of a gear tooth, from the tangential force F_t, the face width b, the module m and the Lewis form factor Y (a function of the tooth count). It was the first rational treatment of tooth strength and it remains the basis of design against bending. It models the tooth as a cantilever beam fixed at its base: the tangential force transmitted between the teeth (coming from the torque) creates a bending moment that tends to break the tooth at the root — the catastrophic failure mode in which a tooth cracks and snaps off. The form factor Y accounts for geometry (gears with more teeth are fatter at the base and stronger, so Y grows). The calculated stress is compared with the bending fatigue strength of the material (with safety factors), since gears see millions of cycles. The basic Lewis formula is later refined by the AGMA standard with factors for stress concentration, dynamic effects, load distribution and surface condition. It is one of the two fundamental criteria in gear design (the other is contact/pitting stress). Enter the tangential force, the face width, the module and the Lewis form factor.
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Gear Tooth Bending Safety Factor
Calculate the bending safety factor of a gear tooth, FS = σ_perm ÷ σ, from the material's allowable (permissible) bending stress σ_perm (MPa, the bending fatigue strength with its factors) and the acting bending stress σ (MPa, from Lewis/AGMA from the load). It is the final bending-design check: the material strength must exceed the demand with adequate margin. Gears run for millions to billions of cycles, so the allowable stress is the material's bending FATIGUE strength (at the design-life cycle count), adjusted by reliability, temperature and life factors. Required safety factors depend on criticality and uncertainty (typically 1.5-3 for bending). If FS is insufficient, the module is increased (bigger, stronger teeth), the face width, or a better material/heat treatment used. A tooth breaking by bending fatigue is a CATASTROPHIC, sudden failure (unlike pitting, which gives progressive signs), since the broken tooth comes loose and can damage the whole drive — so bending is designed with generous margins. With the pitting (contact) safety factor, it defines the gear's robustness. Enter the allowable stress and the acting stress.
Gear Contact Stress (Hertz)
Calculate the Hertzian contact stress on gear tooth surfaces, σ_H = C_p·√(F_t ÷ (b·d·I)), from the elastic coefficient C_p (√MPa, a function of the pair's elastic moduli), the tangential force F_t (N), the face width b (mm), the pinion pitch diameter d (mm) and the geometry factor I (dimensionless). This is the basis of SURFACE FATIGUE (pitting) design — the second fundamental gear failure mode, distinct from bending breakage. When two teeth touch, the contact is practically a LINE, and even moderate loads create very high contact stresses (hundreds of MPa) in the tiny contact area, per Hertz theory. Under repeated cycles, these stresses cause sub-surface fatigue that flakes off small bits of material, forming craters (pitting) that progress, destroy the tooth profile, generate noise and vibration and lead to failure. The elastic coefficient C_p gathers the materials' elastic properties (steel-steel, steel-bronze, etc.), and the geometry factor I, the curvature and contact ratio. Contact stress is compared with the material's pitting resistance (which depends strongly on surface HARDNESS — so gears are often case-hardened). It is one of the two central AGMA criteria. Enter the elastic coefficient, tangential force, face width, diameter and geometry factor.
Gear Dynamic Load
Calculate the effective dynamic load on gear teeth, F_d = F_t·K_v, from the nominal tangential force F_t (N) and the dynamic factor K_v. The dynamic load is the REAL tangential force the teeth bear in operation, larger than the nominal force (simply torque over radius) because of the dynamic effects of meshing at speed. These effects — vibrations, contact impacts, tooth deflections under load and manufacturing errors — make the instantaneous tooth load fluctuate and peak above the average, especially at high pitch-line velocities. The dynamic factor K_v (from Barth or other formulas) quantifies this amplification. The dynamic load is then used in strength checks: in Lewis bending stress (risk of tooth breakage at the root) and Hertzian contact stress (risk of surface fatigue and pitting). Using the nominal load without amplifying by the dynamic factor would underestimate the demands and lead to undersized gears that fail prematurely by fatigue. It is an essential, classic step in gear design. Enter the nominal tangential force and the dynamic factor.
Beam Bending Stress
Enter the bending moment in kN·m and the section modulus in cm³ to get the normal stress at the extreme fibre, σ = M/W, converted to MPa.
D/d Ratio (Pulley-Rope)
Calculate the D/d ratio between the pulley diameter D and the rope diameter d, r = D ÷ d (both in the same unit). The D/d ratio is the most important parameter for the FATIGUE LIFE of a wire rope working over pulleys and drums. Each time the rope passes a pulley, it is FLEXED (bent and unbent), and this repeated bending fatigues the wires — the SMALLER the pulley diameter relative to the rope (lower D/d), the TIGHTER the curve, the greater the wire bending strain and the faster the rope fatigues and breaks. So codes require MINIMUM D/d ratios: typically 18-25 for cranes (each bend costs life), and even higher (40+) for people elevators and high-durability applications. Too small a D/d ratio drastically reduces rope life — doubling the D/d ratio can multiply rope life several times. There is a design trade-off: larger pulleys (high D/d) extend rope life but increase the equipment's size, weight and cost. The D/d ratio, with contact pressure and tension, sets the rope durability. Checking that the D/d ratio meets the code minimum is essential in designing any lifting machine. Enter the pulley and rope diameters.
Tooth Thickness at Pitch Circle
Calculate the tooth thickness measured at the pitch circle of a standard gear, s = (π·m) ÷ 2, from the module m (mm). In a standard (uncorrected) gear, the circular pitch (the distance from one tooth to the next, along the pitch circle) is p = π·m, and it splits equally between the TOOTH (the solid part) and the SPACE (the gap between teeth): half each, hence s = π·m/2. This equality between tooth thickness and space width is what lets two standard gears of the same module mesh perfectly, with one's tooth fitting the other's space with proper clearance. Tooth thickness is a fundamental parameter: it sets the tooth STRENGTH (thicker teeth resist bending more) and the mesh backlash. In CORRECTED gears (with profile shift, used to avoid interference in small pinions, adjust center distance or balance pinion-gear strength), the pitch-circle tooth thickness DIFFERS from π·m/2 — it increases in a positively corrected pinion (strengthening it) and decreases in the gear. Measuring tooth thickness (by the chordal method, with a gear-tooth caliper, or over pins) is a classic gear quality-control check. Enter the module.
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