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A gear contact pattern is the observed region of tooth-flank contact during a stated mesh test. Its position and extent reveal how the installed tooth surfaces meet. A gear contact marking test uses a thin marking layer and controlled rotation to make that region visible; the mounting position, transmission direction and applied torque belong in the test description. Manufacturer guidance treats surface-contact checks as evidence about the assembled mesh, including shaft alignment.^[ref]

A useful bench comparison holds the gear pair and test conditions fixed, records the initial mark, then changes one alignment setting and repeats the test. A mark moving toward one face edge indicates a change in where the flanks meet. This is an observation of contact location, rather than an automatic measurement of operating pressure or a universal diagnosis of the alignment error.

For example, a mark occupying the full20mm gear effective face width is compatible with a hypothetical uniform line load $w=50N/mm$, but also with the tilted positive distributionBoth integrate to1000N, yet their peak line loads are50 and80N/mm. Merely observing full-width contact cannot choose between those gear face load distributions. The gear face load factor or local stress needs additional load/deformation information.

Loaded and lightly loaded patterns can also differ because shaft, housing and tooth deformation change the gap. A quantitative loaded assessment may use strain measurements or a validated contact model alongside the pattern; the2006 gear-rating standard explicitly distinguishes comprehensive load assessment from the limited information in a bearing pattern.^[ref]

Gear effective face width is the usable axial width participating in a stated cylindrical gear mesh. It must be distinguished from each component's nominal gear face width, and from a claim that every part of that width carries equal load. The normal/transverse tooth geometry and gear overlap ratio use this engaged width, while gear face load distribution describes how the load actually occupies it.^[ref]

An elementary geometric model treats the two toothed widths as intervals $I_1=[a_1,b_1]$ and $I_2=[a_2,b_2]$ on their common axial coordinate. Their available overlap isFor one20mm-wide gear occupying $[-10,10]\mathrm{mm}$ and a16mm-wide mate centered at $6\mathrm{mm}$, the second interval is $[-2,14]\mathrm{mm}$. Their overlap is only12mm. The mean line load for a hypothetical uniformly shared normal force of1200N is then100N/mm, rather than the60N/mm obtained by using the first gear's entire nominal width.

With gear normal module $m_n=2\mathrm{mm}$ and gear helix angle $|\beta |=30{}^{\circ }$, gear axial pitch is $p_x=4\pi \mathrm{mm}$. The12mm overlap gives geometric gear overlap ratio $b_{\mathrm{eff}}/p_x=3/\pi \approx 0.9549$, compared with $5/\pi \approx 1.5915$ for20mm. Axial positioning can therefore alter overlap without changing the ideal gear ratio.

The interval calculation assumes the flanks remain usable throughout their common extent. Relief, damaged regions or partial contact can require a more specific working-width convention. Uneven loading is not generally corrected by shrinking width until a uniform-load formula looks plausible; a stated load-distribution model is needed to predict local loading.

Gear face load distribution describes how tooth-contact load is spread across the engaged gear face width. Equal total transmitted force can produce different local loads when shafts tilt, supports deform or tooth traces are imperfect. A geometrically available gear effective face width therefore does not establish uniform loading. The manufacturer's tooth-flank gear face load factor accounts for support arrangement and the expected quality of contact under load.^[ref]

For a simple axial line-load model, let $w(z)$ have units $N/mm$ over $0\le z\le b$. Its total normal force isUniform loading gives $w=F_n/b$. An illustrative tilted distribution isIts integral remains $F_n$, while its largest line load is $(1+\lambda )F_n/b$. With $F_n=1000\mathrm{N}$, $b=20\mathrm{mm}$ and $\lambda =0.6$, the two edges carry20 and80N/mm, compared with the50N/mm mean. The peak-to-mean ratio1.6 characterizes this constructed model; it is not a substitute for the manufacturer's standard factor calculation.

The gear contact pattern can reveal where mating tooth flanks touch, but an unloaded marking test alone does not measure the operating force density. A loaded comparison must specify applied torque, shaft positioning and support deformation. In helical gears, changing tooth-contact lines and multiple simultaneous contacts also require a consistent choice of time and projected width. A wider tooth can reduce mean load per width while still suffering a concentrated edge load; width and alignment must be considered together.

Gear face width $b$ describes the width of the toothed portion. For a cylindrical spur gear or helical gear, it is measured parallel to the rotation axis, rather than along the slanting tooth trace. A projecting hub need not be part of the toothed width. Manufacturer drawings distinguish face width, hub width and total length as separate dimensions.^[ref]

The available width of one gear is not necessarily the gear effective face width of a pair. For two aligned cylindrical gears with equal width $b$ and relative axial offset $e$, a simple interval-overlap model givesAn original $b=20\mathrm{mm}$ pair offset by $4\mathrm{mm}$ has only $16\mathrm{mm}$ of common axial extent. If a hypothetical normal load $1000\mathrm{N}$ were uniformly spread across that extent, mean load per axial width would rise from $50$ to $62.5N/mm$. This is bookkeeping for an ideal uniform load, not a tooth-strength prediction. Actual gear face load distribution also depends on alignment and deformation.

For helical gears, width also enters gear overlap ratio:With gear normal module $m_n=2\mathrm{mm}$ and gear helix angle $|\beta |=30{}^{\circ }$, reducing engaged width from $20$ to $16\mathrm{mm}$ reduces ideal overlap from $1.5915$ to $1.2732$. The overall gear ratio is unchanged. Width therefore affects engagement and loading even when tooth counts and shaft speeds remain the same.^[ref]