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Gear tooth load sharing is the division of transmitted mesh load among simultaneously contacting tooth pairs. Gear contact ratio determines how engagement intervals overlap, but it does not by itself specify each pair's force. Tooth and support deformation, profile geometry and operating speed affect that distribution. A primary dynamic analysis explicitly coupled tooth stiffness with calculated dynamic loads rather than assigning a fixed full load to every active pair.^[ref]

An illuminating quasistatic model treats active pairs as parallel springs with a common normal approach $\delta$. With pair stiffnesses $k_j$ and no gaps, $F_j=k_j\delta$, soThe stiffness sum is gear mesh stiffness, and ${\eta }_j$ is a gear tooth load fraction. If two active pairs have stiffness100 and150N/µm under total normal load1000N, their common approach is4µm and loads are400 and600N. Equal sharing requires equal effective stiffness in this particular model.

Include a geometric gap $e_j$ by writing $F_j=k_j\max (\delta -e_j,0)$. If the second pair has a2µm gap, the same1000N total load gives $100\delta +150(\delta -2)=1000$ with $\delta$ measured in micrometres. The result is $\delta =5.2\mu \mathrm{m}$, $F_1=520\mathrm{N}$ and $F_2=480\mathrm{N}$. At only100N, the first pair alone compresses1µm, leaving the second pair out of contact.

For an ideal spur gear mesh with transverse contact ratio between1 and2, the fraction of a tooth-entry cycle having two pairs in contact is ${ϵ}_{\alpha }-1$. A ratio1.635 therefore gives about63.5% two-pair contact and36.5% single-pair contact. This is geometric timing, not a guarantee of equal forces in the overlapping part. A gear friction work per engagement calculation must use each pair's own load history; dynamic motion can require more than the quasistatic spring approximation.

Gear transverse module $m_t$ specifies tooth spacing in the plane perpendicular to a cylindrical gear's axis. If $N$ is tooth count and $d$ is reference diameter, $m_t=d/N$, so reference-circle gear transverse pitch is $p_t=\pi m_t$. For a helical gear, this transverse section differs from the section normal to the tooth direction.^[ref]

Unwrapping the reference cylinder makes the tooth direction a slanted line. Projecting its normal spacing into the circumferential direction giveswhere $m_n$ is gear normal module and $\beta$ is reference gear helix angle. At zero helix angle the normal and transverse modules coincide, recovering the spur gear case.

For an original example, take $m_n=2.5\mathrm{mm}$, $\beta =30{}^{\circ }$ and $N=24$. Then $m_t\approx 2.88675\mathrm{mm}$ and $d\approx 69.282\mathrm{mm}$. A compatible 48-tooth gear has reference diameter $138.564\mathrm{mm}$, giving standard separation $103.923\mathrm{mm}$. Treating2.5 as the transverse module instead would incorrectly give diameters60 and120mm and separation90mm.

The transverse specification is convenient for reference diameters and axis separations, while the normal specification is convenient for the tooth-normal geometry. A drawing must identify which module and pressure-angle planes it uses; a numerical module value alone is incomplete for a helical gear. These are reference dimensions: a changed operating separation must additionally be interpreted through the mounted geometry, rather than redefining the gear's original module.

The operating pitch circles of two meshing gears are the imaginary circles that reproduce their motion by rolling without slipping. They are properties of the working pair and its mounting, whereas a gear reference circle specifies one gear's nominal geometry. Working pitch circle and intermeshing pitch circle are alternative names.^[ref]

Consider an external pair of spur gears on fixed parallel axes, with axis separation $a$ and tooth counts $N_1,N_2$. Their operating radii $r_{w1},r_{w2}$ must satisfy both contact geometry and the tooth-count gear ratio:Solving givesThus the operating radii scale with the actual separation, while the gear ratio stays fixed by tooth counts. These are kinematic equations, conditional on the teeth actually maintaining compatible contact.

For an original example, choose gear module $m=2\mathrm{mm}$ and tooth counts 20 and 40. Reference radii are $20\mathrm{mm}$ and $40\mathrm{mm}$; their standard separation is $60\mathrm{mm}$. If a compatible mounting uses $a=61\mathrm{mm}$ instead, operating radii become $20.333\dots \mathrm{mm}$ and $40.666\dots \mathrm{mm}$. At input angular velocity $30rad/s$ the output has $-15rad/s$, and both operating tangential speed magnitudes are $0.61m/s$. Using the reference radii in that speed calculation would incorrectly give $0.60m/s$.

For tooth flanks formed as involutes, the fixed gear base circles permit constant-ratio motion over a range of separations. If their radii are $r_{b1},r_{b2}$, the common-normal geometry giveswhere ${\alpha }_w$ is the operating pressure angle. Increasing separation therefore increases that angle, while leaving each gear reference circle and gear base circle unchanged. This geometric possibility does not guarantee adequate tooth contact, clearance or strength at an arbitrary separation: the physical tooth extents and mounting must still be checked.^[ref]

The operating pressure angle ${\alpha }_w$ is the gear pressure angle of an actual meshing pair at its mounted axis separation. For external spur gears with involute profiles, it is the angle between the gear line of action and the common tangent to the operating pitch circles. Changing the separation changes this angle even though the individual gear base circles remain fixed.^[ref]

Let the two base radii be $r_{b1},r_{b2}$ and the mounted axis separation be $a$. The right-triangle geometry of the gear line of action givesIf $a_0=r_1+r_2$ is the reference separation and both reference gear pressure angles equal $\alpha$, then $r_{bi}=r_i\cos \alpha$, henceThe gear reference circles and gear ratio remain unchanged by this mounting adjustment; the operating pitch circles expand or contract to fit $a$.

For example, reference radii $20\mathrm{mm}$ and $40\mathrm{mm}$ with $\alpha =20{}^{\circ }$ give $a_0=60\mathrm{mm}$. Mounting a compatible pair at $a=61\mathrm{mm}$ givesAt an imposed gear tangential force $100\mathrm{N}$, the associated gear separating force is about $41.30\mathrm{N}$, compared with $36.40\mathrm{N}$ at the reference angle. Holding torque instead of tangential load would additionally require using the changed operating radius.

The angle formula is only a geometric condition. Usable tooth flanks must still reach the contact region and maintain an adequate gear contact ratio; excessive separation can destroy useful contact. Likewise, the existence of a real inverse cosine does not establish that a proposed mounting has suitable backlash or mechanical strength.^[ref]

A reverted gear train returns the output rotation axis to the input axis, usually through a two-stage compound gear train. The input and output gears can rotate independently on coaxial shafts, while the intermediate gears share another rotating body. Coaxial axes are a layout condition in addition to the desired gear ratio.

For standard unshifted spur gears, let gears1 and2 mesh in the first stage and gears3 and4 in the second, with2 and3 rigidly connected. The common intermediate-axis separation requiresIf both stages use the same gear module, this reduces to $N_1+N_2=N_3+N_4$. The reduction ratio independently satisfies $i=(N_2/N_1)(N_4/N_3)$. Both conditions must be satisfied; multiplying arbitrary stage ratios does not ensure a reverted layout.^[ref]

For an original design example, choose counts20,60,16 and64, with module $2\mathrm{mm}$ in both stages. Both sums are80, so both separations are $80\mathrm{mm}$. The reduction is $3\times 4=12$, taking an input of1440 revolutions per minute to an output of120. Two external meshes give the same input/output rotation direction.

The gears occupy different axial planes so that coaxial input and output do not require them to be one rigidly joined gear. A practical design still needs independent shaft support, compatible tooth geometry in each stage and adequate clearances. Operating with shifted tooth geometry or another mounting changes the applicable separation equations; the simple tooth-sum condition belongs to the stated standard-gear model.

A spur gear is a cylindrical gear with straight teeth parallel to its rotation axis. A conventional external pair transmits motion between parallel axes. Unlike the inclined teeth of a helical gear, the ideal straight-tooth geometry produces no axial thrust from the tooth-normal contact force; there are still tangential and separating loads that the mounting must carry.^[ref]

For compatible spur gears, reference diameter is $d=mN$, where $m$ is gear module and $N$ is tooth count. Standard reference axis separation is $(d_1+d_2)/2$. Two external spur gears have signed angular velocity relation ${\omega }_2/{\omega }_1=-N_1/N_2$: they rotate in opposite directions. Matching gear module alone is insufficient; tooth profiles and gear pressure angles must also be compatible.

For an original example, take gear module $3\mathrm{mm}$ and tooth counts 20 and 60. Reference diameters are $60\mathrm{mm}$ and $180\mathrm{mm}$, so standard axis separation is $120\mathrm{mm}$. Driving the smaller spur gear at $600$ revolutions per minute gives the larger one $200$ revolutions per minute in the opposite direction. In a steady ideal transmission, an input torque of $2\mathrm{N}\mathrm{m}$ becomes $6\mathrm{N}\mathrm{m}$ at the output. The corresponding input gear tangential force is $2/0.030\approx 66.7\mathrm{N}$; multiplying this same force by the output reference radius $0.090\mathrm{m}$ recovers the output torque.

A simple bench demonstration marks one tooth on each spur gear and counts complete turns, testing the three-to-one gear ratio. Measuring transmitted torques and rotational speeds additionally tests gear efficiency. Observing the rotation ratio alone cannot establish a tooth's allowable load or the losses caused by friction and backlash.