EngineeringMechanical EngineeringPart 15 of 15
Power Screw Design
With power screws, friction is the enemy and the safety feature at the same time. Remove enough of it and the efficiency rises; remove a little too much and the load runs back down under its own weight.
- Thread forms
- Torque and efficiency
- Self-locking condition
- Thread stresses
Executive summary
Power screws, also called translation screws, convert rotary motion into linear motion — rotational torque and power into linear force and power. The essential difference from a fastening screw is one of intent: a power screw is used constantly and efficiency matters, while a fastening screw is used rarely and high friction is desirable so that it stays done up.
Taken to the limit, that distinction is absolute. At zero friction a power screw would be 100 per cent efficient, and a fastening screw would not fasten at all.
Arrangements and terminology
- Arrangement 1Screw rotates, located by a thrust collar or bearing. The nut is prevented from turning but free to slide, and advances axially with the load.
- Arrangement 2Nut rotates, located by a thrust collar or bearing. The screw is prevented from turning but free to slide, and advances axially with the load.
- Arrangement 3Screw rotates and translates, transmitting load through a thrust collar. The nut is fixed and cannot turn. Least common for power screws; usual for fastening.
Thread forms and proportions
| Form | Face angle | Character |
|---|---|---|
| Square | 0° | Most efficient. Can only be cut on a lathe — not with a die, milling cutter or thread rolling — so rarely used in practice. |
| Modified square | 5° | The practical substitute for the square thread. Producible by any method. |
| Trapezoidal metric | 15° (30° included) | The metric equivalent of the ACME thread, which has a 29° included angle. The general-purpose choice. |
| Buttress | Varies | Asymmetric, for unidirectional loads only. Several variations exist with differing face angles and thread depths. |
Unlike fastening threads, power screws have no standardised metric pitch and depth series. Pitch is chosen for good proportions relative to nominal diameter — broadly, pitch rises from a few millimetres on 10 mm screws to around 17 mm on 100 mm screws. Manufacturers publish their own preferred combinations.
Friction, torque and self-locking
Coefficient of friction
Between lubricated metal sliding faces at the thread or thrust collar, the coefficient depends on lubrication type and frequency, workmanship — accuracy and surface finish — and the number of cycles, since mating surfaces bed in with use.
- α
- thread face angle — zero for a square thread, in which case φ′ = φ
- F
- axial thrust, N
- d
- pitch diameter, m or mm consistently with T
Self-locking
If the lowering torque is positive, torque must be applied to lower the load — and with zero applied torque the load stays where it is. That is self-locking, an inbuilt safety feature valuable in jacks, lifting screws and clamping devices.
Large helix angles arise from multi-start threads. If overhauling is required — for a mechanism that must back-drive — a multi-start thread is the usual route. Conversely, a design that relies on self-locking must be checked against the lowest credible friction coefficient, because self-locking is lost as friction falls.
Self-locking under static conditions does not guarantee the load will hold under vibration, shock or thermal cycling, and running-in reduces the friction coefficient over time. Where the consequence of a dropped load is severe, a positive mechanical brake or locking device is required regardless of the calculated self-locking condition.
Efficiency and collar friction
- F
- axial thrust, N
- L
- lead, m — the linear distance moved in one revolution
- T
- torque applied over that revolution, Nm
Worked example
A single-start trapezoidal metric thread of 30 mm nominal diameter carries a 2 kN axial load at average friction, with a pitch of 8 mm.
| Quantity | Thread only | With thrust collar |
|---|---|---|
| Pitch diameter | d = 30 − 0.5 × 8 = 26 mm | |
| Friction angle | tan φ′ = 0.125 / cos 15° = 0.1294, so φ′ = 7.37° | |
| Helix angle | tan θ = 8 / (π × 26) = 0.0979, so θ = 5.59° | |
| Torque raising | 5.99 Nm | 10.7 Nm |
| Torque lowering | 0.81 Nm | 5.5 Nm |
| Efficiency raising | 42.5 per cent | 23.9 per cent |
- rm
- mean radius of the thrust face. For a collar of 50 mm outside and 25 mm inside diameter, rm = 18.75 mm.
Collar friction nearly halved the efficiency in this example and almost doubled the raising torque. Replacing a plain rubbing thrust face with a rolling element thrust bearing makes that term negligible — usually the cheapest and largest single improvement available to a power screw design. Note the corollary: with collar friction removed, the lowering torque falls to 0.81 Nm and the margin to overhauling shrinks considerably.
Thread stress analysis
Accurate stress analysis of a screw and nut is complicated. Deflection means the first one or two threads carry most of the load, and clearance, fillet radii, surface finish and machining accuracy all matter. The simplified analysis below assumes even load distribution across all engaged threads — useful and instructive, provided its limits are understood.
- a
- nut thickness, usually in the range 0.75 to 1.5 times nominal diameter
- t
- thread depth
- h
- thread height or base thickness — 0.5p square, 0.5437p modified square, 0.634p trapezoidal metric, 0.875p buttress
- b
- total thread length in contact
| Rubbing speed (m/s) | Maximum pressure (MPa) |
|---|---|
| Below 0.05 | 20 |
| 0.05 to 0.1 | 10 |
| 0.1 to 0.2 | 5 |
| Above 0.2 | 2.5 |
A well designed power screw does not fail by crushing the thread. Bearing pressure must be held low enough to maintain lubrication and limit wear, and that constraint sits well below the crushing strength of the material. Note also that bending stress peaks at the outer fibre while shear peaks at the centre where bending is zero — so the two need not be combined.
Design checklist
- Transmission arrangement chosen and the rotating and translating members identified.
- Thread form selected, with modified square or trapezoidal preferred over true square for manufacturability.
- Pitch chosen in good proportion to nominal diameter.
- Pitch diameter, minor diameter and thread depth derived from the chosen form.
- Friction coefficient chosen for the actual lubrication and finish, with start-up increase applied.
- Raising and lowering torque both calculated.
- Self-locking condition checked against the lowest credible friction coefficient.
- Positive brake or lock provided where a dropped load would be serious.
- Collar friction included or eliminated by a thrust bearing, and the decision reflected in both torque and efficiency.
- Efficiency calculated and carried into the prime mover sizing.
- Nut thickness set within the usual proportion range and thread count derived from it.
- Bearing pressure checked against the allowable value for the actual rubbing speed.
- Thread bending and shear stresses checked, with a reduced effective thread count where prudent.