ISO Metric Thread Sizes Explained: M6 to M20 Decoded

A metric thread designation like "M10x1.5" tells you the nominal outer diameter (10 mm) and the thread pitch — the distance between adjacent threads, measured along the axis (1.5 mm). If the pitch is omitted (just "M10"), it defaults to the coarse pitch defined in ISO 261/ISO 262 for that diameter. Coarse threads are the default for a reason — they're what you should reach for unless you have a specific reason not to.

Coarse vs. fine: what actually changes

This is the part most quick-reference charts skip. Coarse and fine pitch aren't just "two options of the same thread" — they trade off differently:

• Coarse pitch has deeper thread flanks and a smaller minor (core) diameter than a fine thread of the same nominal size. That makes coarse threads more resistant to cross-threading and more forgiving of dirt, paint buildup, and minor damage during assembly — but it also means less thread engagement per unit length in soft materials. This is why coarse is the default for general fastening, weld nuts, and anything that gets assembled and disassembled in the field.
• Fine pitch has a larger stress area for the same nominal diameter (more thread engagement per unit length), holds torque adjustments more precisely (useful for micrometers, adjusting screws), and resists loosening under vibration slightly better in some joint designs. The tradeoff: fine threads are more sensitive to cross-threading and gall more easily in soft or dissimilar metals.

Rule of thumb: default to coarse. Reach for fine pitch only when you have a specific reason — thin-walled tapped holes where you need more thread engagement, adjustment mechanisms, or a standard that specifically calls for it (some automotive and aerospace hardware).

Reference table: coarse and fine pitch, M6–M20

Nominal size	Coarse pitch	Common fine pitch options
M6	1.0 mm	0.75 mm
M8	1.25 mm	1.0 mm, 0.75 mm
M10	1.5 mm	1.25 mm, 1.0 mm
M12	1.75 mm	1.5 mm, 1.25 mm
M16	2.0 mm	1.5 mm
M20	2.5 mm	2.0 mm, 1.5 mm

Sources for this table: Newman Tools metric coarse thread data and Machining Doctor's metric thread chart, cross-checked and in agreement on all six sizes above. These figures follow ISO 261 (general purpose metric screw threads) and ISO 262 (selected sizes).

Property classes: what "8.8" or "10.9" actually means

Metric bolts are marked with a two-number property class (e.g. 8.8, 10.9, 12.9), and the numbers are not arbitrary — they encode real mechanical properties:

• The first number × 100 ≈ the minimum tensile strength in MPa (N/mm²). Class 8.8 → roughly 800 MPa tensile strength; class 10.9 → roughly 1000 MPa; class 12.9 → roughly 1200 MPa.
• The second number × the first, divided by 10 ≈ the yield-to-tensile ratio ×10. For class 8.8, 8×8=64 → yield strength is about 640 MPa (roughly 80% of tensile). This ratio matters because it tells you how much margin exists between yielding (permanent deformation) and fracture.

In practical terms: 8.8 is the standard grade for general structural and machine assembly work, 10.9 is used where higher clamp load or a lighter fastener for the same load is needed, and 12.9 is reserved for high-stress applications where the fastener's designer has specifically accounted for its lower ductility and higher sensitivity to hydrogen embrittlement. Don't default to 12.9 "because it's stronger" — higher-grade bolts are more brittle and less forgiving of over-torque or shock loading; picking the highest class available is not automatically the safer choice.

Why we won't give you a single "correct" torque number

Search for a metric bolt torque chart and you'll find numbers that disagree by 30–40% between reputable sources for the same size and property class. That's not an error on anyone's part — tightening torque depends on the coefficient of friction at the joint (which changes with plating, lubrication, and surface finish), and different charts assume different friction coefficients and different target percentages of proof load. As one example: two commonly cited charts give M10 class 8.8 dry torque as roughly 41 Nm in one source and roughly 57 Nm in another — both are "correct" for their respective friction assumptions, and neither is correct for your joint unless you've matched the assumption.

The practical approach:

1. Use the general formula T ≈ K × D × F, where T is torque, D is nominal diameter, F is the target clamp force (commonly ~75% of proof load), and K is a friction coefficient — typically 0.2 for dry steel-on-steel, lower for lubricated or coated fasteners.
2. Treat any published torque chart as a starting point, not a final answer, unless it explicitly states the friction coefficient and lubrication condition it assumes.
3. For anything safety-critical, use the torque value specified by the fastener manufacturer or the assembly drawing — not a generic web chart.

Mathematical Thread Equations and Torque Mechanics

To transition from general rule-of-thumb sizing to rigorous engineering validation, several core equations must be applied:

1. Tensile Stress Area ($A_s$)

A bolt does not fail at its nominal diameter $d$, nor at its minor diameter. It fails across an effective stress area. According to ISO 898-1, the tensile stress area ($A_s$) for metric threads is:

$$A_s = \frac{\pi}{4} \left( d - 0.938194 \cdot P \right)^2$$

Where:

• $d$: Nominal thread diameter (e.g., 10 mm for M10)
• $P$: Thread pitch (e.g., 1.5 mm for M10 coarse)

2. Thread Stripping Shear Area ($A_{shear}$)

To prevent the internal threads (in a tapped hole or nut) from stripping before the bolt breaks, the shear area of the engaged threads ($A_{shear}$) must be evaluated:

$$A_{shear} \approx \pi \cdot d \cdot L_{eng} \cdot 0.5$$

Where $L_{eng}$ is the engaged thread length. Generally, to ensure the bolt breaks in tension rather than stripping the threads, $L_{eng} \geq 1.0 \cdot d$ in steel, and $L_{eng} \geq 1.5 \cdot d$ in aluminum.

3. Tightening Torque and Bolt Preload ($T$)

The relationship between tightening torque ($T$) and the resulting clamp force (preload $F_i$) is governed by friction:

$$T = K \cdot d \cdot F_i$$

Where:

• $K$: Torque coefficient (friction factor). $K \approx 0.20$ for dry steel, $0.15$ for lightly oiled steel, and can drop to $0.10$ for specialized anti-seize coatings.
• $F_i$: Tightening preload (tension). Safe target preloads are usually set to $90%$ of the bolt material's yield strength ($f_y$): $$F_i = 0.9 \cdot f_y \cdot A_s$$

Case Study: M10 Coarse vs. M10 Fine under 25 kN Static Load

An assembly joint is subjected to a static axial tensile working load of $F_{load} = 25$ kN. We compare M10 coarse and M10 fine thread configurations.

• M10 Coarse ($P = 1.5$ mm):
  • $A_s = \frac{\pi}{4} \left( 10 - 0.938194 \cdot 1.5 \right)^2 \approx 58.0 \text{ mm}^2$
• M10 Fine ($P = 1.25$ mm):
  • $A_s = \frac{\pi}{4} \left( 10 - 0.938194 \cdot 1.25 \right)^2 \approx 61.2 \text{ mm}^2$

Selecting Property Classes: We evaluate Class 8.8 (yield strength $f_y = 640$ MPa) versus Class 10.9 (yield strength $f_y = 940$ MPa).

• Option A: M10 Coarse Class 8.8
  • Yield Limit Force: $F_{yield} = A_s \cdot f_y = 58.0 \text{ mm}^2 \cdot 640 \text{ N/mm}^2 \approx 37.1 \text{ kN}$
  • Safety Factor: $S_f = \frac{F_{yield}}{F_{load}} = \frac{37.1}{25} \approx 1.48$
• Option B: M10 Coarse Class 10.9
  • Yield Limit Force: $F_{yield} = 58.0 \text{ mm}^2 \cdot 940 \text{ N/mm}^2 \approx 54.5 \text{ kN}$
  • Safety Factor: $S_f = \frac{54.5}{25} \approx 2.18$
• Option C: M10 Fine Class 8.8
  • Yield Limit Force: $F_{yield} = 61.2 \text{ mm}^2 \cdot 640 \text{ N/mm}^2 \approx 39.2 \text{ kN}$
  • Safety Factor: $S_f = \frac{39.2}{25} \approx 1.57$

Analysis: Option A provides a safety factor of 1.48, which may be tight under dynamic loads. Bumping up the property class to 10.9 (Option B) increases the safety factor to 2.18 without changing the physical footprint. Alternatively, switching to a fine pitch Class 8.8 (Option C) gives a modest $6%$ safety increase due to the larger tensile area.

Choosing a size for a real load (a worked approach, not just "bigger is safer")

Oversizing a fastener isn't free — it adds weight, cost, and can require more material around the joint to support a larger hole. A rough sizing approach:

1. Estimate the working load on the joint (shear, tension, or both).
2. Apply a safety factor appropriate to the application (2–4× for general machinery, higher for safety-critical or fatigue-loaded joints).
3. Check that a candidate size's proof load (from the property class) comfortably exceeds that factored load — proof load, not tensile strength, is the number you should be comparing against for static joints, since proof load represents the load at which the bolt starts to take a permanent set.
4. For joints with dynamic or vibrational loading, thread engagement length and locking method (adhesive, nylon insert, safety wire) usually matter more than bumping up one thread size.

Quick reference: when each size shows up in practice

Size	Typical use
M6	Light equipment covers, small brackets, electronics enclosures
M8	General machine assembly, motor mounts, moderate structural brackets
M10	Structural machine frames, heavier motor mounts, medium-duty flanges
M12	Heavy machinery, structural steel connections, higher-load flanges
M16 / M20	Heavy structural steel, large flange connections, foundation bolts

The bottom line

A thread size chart alone doesn't tell you what to specify — it tells you what's available. The actual decision (coarse vs. fine, which property class, what torque) depends on the joint's load path, the material it's threading into, and how it will be assembled and maintained. When in doubt, match an existing proven design rather than guessing from a table in isolation.

Sources: Newman Tools — Metric Coarse Thread Data · Machining Doctor — Metric Thread Charts · Fastenal — Torque-Tension Relationship for Metric Fasteners.