Stand Mixer Design Evaluation

Overview

For ME41 (Engineering Design II) our team tore down a "Kitchen in the Box" 3.2 quart stand mixer and evaluated it like a real product instead of a textbook problem. Once it was open, three subsystems were doing the actual work: a spur-and-planetary gear train, ball bearings sitting in a plastic housing, and a tilt-head hinge with a safety interlock. None of it came with manufacturer specs, so most of the numbers below are things we measured, counted, or back-calculated ourselves, then ran through free-body diagrams, life calculations, and FEA from Shigley's.

Compact spur plus planetary train that trades motor speed for mixing torque.

• Measured the paddle at 165 RPM about the bowl from slow-motion video, then counted every tooth in the train (13-tooth motor pinion through a 95, 9, and 75-tooth spur set into a 41-tooth ring and 13-tooth planet).
• Train value e = (13 x 9) / (95 x 75) = 0.0164, which back-calculates the motor to roughly 10,050 RPM. The 300 W motor plate then gives a no-load torque near 0.29 N-m.
• The planetary stage adds 2.15x, for a total 28.3:1 reduction, about 8.1 N-m at the paddle and roughly 153 N at the beater edge (r = 5.3 cm).
• Modelled the gears in Onshape (20 degree pressure angle, 1.25 mm module, one-third root fillet, 0.02 cm backlash), then ran a SolidWorks motion study and a static FEA with AISI 4130 steel and PA6 nylon, control-meshed at the tooth contacts under 2.09 N-m.
• The steel gears come out at a reasonable stress and the nylon runs well below that, but factor of safety is only just above 1 in the most loaded spots, so there isn't much room for overload or material variation.

A single deep-groove ball bearing carrying the belt and gear loads on the intermediate shaft.

• Identified the part as a 608 deep-groove ball bearing (8 x 22 x 7 mm, dynamic rating 3.3 kN) sitting in an injection-moulded glass-filled nylon housing.
• Traced the load path: belt tension to pulley to intermediate shaft at 17.4 N-m, giving a 2320 N belt radial load plus a 253 N gear separating force, or 2570 N radial and 257 N axial at the bearing.
• The ABMA equivalent load works out to 2714 N, so L10 = (C / P) cubed x 1e6 = 1.8 million revolutions, about 182 hours at 165 RPM (136 to 234 hours across a plus or minus 300 N band on the load rating).
• That short life is driven almost entirely by the constant belt tension. Housing compliance, press-fit creep, grease breakdown, and start-stop cycling all pull the real figure lower, so the bearing is likely to go loose or noisy before any gear fails.

The tilt-head lock, its torsion spring, and the safety switch that cuts power when the head is up.

• A button push slides a travel rod back so its thin section clears the curved latch slots, the head swings through 60 degrees, and a compression spring re-engages the thick section to lock it. The same motion trips a switch that kills motor power whenever the head is raised.
• Modelling the head as a rigid body (about 2.0 kg, centre of mass 0.20 m from the hinge), head weight alone gives a 3.9 N-m moment and a hard 50 N user push raises the total to 11.4 N-m.
• The torsion spring (2.2 mm wire, 8 active coils) supplies about 0.79 N-m of restoring torque at around 910 MPa wire stress against a 2050 MPa music-wire yield, a factor of safety near 2.3.
• Latch contact pressure runs 3 MPa in normal use and 10 MPa under the hard push, both well below yield, so wear in the slots rather than strength sets the life. Travel-rod axial stress is about 2 MPa, negligible.

Gear train CAD & FEA

SolidWorks static study of the driven spur pair under 2.09 N-m, control-meshed at the tooth contacts.

Full report

Tools