Evaluate the friction loss, fitting loss, total pressure drop, and fan power of an HVAC air duct against the ASHRAE standard. Vary airflow, cross-section (circular, rectangular, or flat-oval), wall material, air temperature, and fitting count to see the Darcy-Weisbach + Swamee-Jain pressure drop and the recommended velocity range update in real time.
Parameters
Airflow Q
m³/h
Duct shape
Selecting a shape auto-updates the hydraulic diameter formula
Height H (diameter for circular)
mm
Width W (rectangular / flat-oval)
mm
Duct length L
m
Fittings N
Elbows, branches, dampers (K = 0.5 each assumed)
Wall material
Absolute roughness ε [mm] is set automatically
Air temperature T
°C
Density and viscosity scale with temperature
Results
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Hydraulic diameter D_h (mm)
—
Velocity V (m/s)
—
Reynolds number Re
—
Friction loss (Pa)
—
Total pressure drop (Pa)
—
Fan power (W)
—
Duct section & velocity profile
Air flows from left (high pressure) to right (low pressure). The dotted profile shows the turbulent velocity distribution (faster in the centre). Yellow ticks are fittings.
Pressure drop vs airflow
Pressure drop by duct shape
Theory & Key Formulas
$$\Delta P = f\frac{L}{D_h}\frac{\rho V^{2}}{2} + \sum K\frac{\rho V^{2}}{2},\qquad f = \frac{0.25}{\left[\log_{10}\!\left(\dfrac{\epsilon}{3.7\,D_h} + \dfrac{5.74}{Re^{0.9}}\right)\right]^{2}}$$
f: Swamee-Jain friction factor, K: local loss coefficient, D_h: hydraulic diameter (4A/P for rectangular), ε: absolute wall roughness, Re: Reynolds number.
$$D_h = \frac{4A}{P},\qquad Re = \frac{\rho V D_h}{\mu},\qquad P_{\text{fan}} = \frac{Q\,\Delta P}{\eta_{\text{fan}}}$$
A: cross-section area, P: wetted perimeter, ρ: temperature-corrected air density, μ: dynamic viscosity, η_fan: assumed fan efficiency (70%).
HVAC Duct Pressure Drop — ASHRAE Standard Calculation
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An air duct is just a sheet-metal box that carries air, right? Why do HVAC engineers go through all this "pressure drop" math?
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Great starting point. As air moves through the duct, friction on the walls plus losses at every elbow, branch and damper steal pressure from the stream. That total loss ΔP is exactly what the fan has to overcome, so it sets the fan power: P_fan = Q·ΔP/η. For a 5000 m³/h system at 90 Pa, that's already about 180 W just to push the air—multiply by the dozens of ducts in a building and you see why ΔP drives the annual electric bill. Get ΔP wrong and you mis-size the fan, the duct, and the energy budget.
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Got it. When I switch "Duct shape" from rectangular to flat-oval to circular at the same flow, the pressure drop changes a lot. Circular comes out lowest.
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Exactly. For the same cross-section area, a circle has the smallest wetted perimeter, so its hydraulic diameter Dh = 4A/P is the largest. Because ΔP scales as 1/Dh, circular ducts always win on pressure drop, then flat-oval, then rectangular. So why are commercial buildings full of rectangular ducts? Because they fit into a shallow plenum above a dropped ceiling. Real designers constantly trade off "low pressure loss (round)" against "fits the building (rectangular)".
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Switching to "Flexible duct" makes the friction loss explode. What's going on?
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Flex duct has a corrugated inner surface, giving an absolute roughness ε ≈ 3.0 mm—more than 30× galvanized steel at 0.09 mm. In the Swamee-Jain term ε/(3.7·Dh), that pushes the bracket way up, and the friction factor f doubles or triples. That's why you should never run long flex from the air handler. Industry practice is to use flex only at the last 1-2 m to the diffuser, with a generous bend radius and well stretched—not coiled in the ceiling.
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The verdict labels velocity as "Low / OK / High / Excessive". What's that based on?
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It's straight from the ASHRAE recommended-velocity ranges. Mains 4-8 m/s and branches 2-5 m/s read as "OK". Above 10 m/s air rush noise exceeds NC30—not acceptable in offices or homes. Below 2 m/s the duct is oversized and you waste money on sheet metal. With the defaults you're at 5.56 m/s, which is the textbook sweet spot for a commercial main. Try doubling the airflow—velocity jumps past 11 m/s and the verdict flips to "Excessive".
Frequently Asked Questions
ASHRAE Handbook (Fundamentals) uses the Darcy-Weisbach equation, ΔP = f·(L/Dh)·(ρV²/2), for friction loss. The friction factor f is implicitly defined by the Colebrook-White equation, but in practice the explicit Swamee-Jain approximation f = 0.25/[log10(ε/3.7Dh + 5.74/Re^0.9)]² is widely used. This tool evaluates f with Swamee-Jain and is accurate to roughly ±2% in the turbulent regime above Re ≈ 4000.
Use the hydraulic diameter Dh = 4A/P, where A is the cross-sectional area and P is the wetted perimeter. For a rectangle W×H, Dh = 4WH/(2(W+H)) = 2WH/(W+H). For a flat-oval section (width W, height H, two semicircular ends) A = (W-H)·H + π·H²/4 and P = 2(W-H) + π·H. ASHRAE also publishes Huebscher's equal-friction equivalent circular diameter De = 1.30·(WH)^0.625/(W+H)^0.25, which differs from the hydraulic diameter by only a few percent for aspect ratios up to 1:8.
For general comfort applications ASHRAE recommends 4-8 m/s in main ducts, 2-5 m/s in branches, and 2-3 m/s at supply diffusers, balancing noise and pressure loss. Above 10 m/s air-rush noise exceeds NC30 and is a common source of occupant complaints. Below 2 m/s ducts become oversized, raising first cost and condensation risk. This tool reports a "Low / OK / High / Excessive" velocity verdict automatically.
ASHRAE typically recommends 0.5-1.0 Pa/m (about 0.08-0.10 in WG per 100 ft) as the design target for low-pressure systems. Residential designs often use 0.5 Pa/m, commercial offices 0.8 Pa/m, and high-velocity industrial systems 2-4 Pa/m. A lower friction rate yields larger ducts and higher first cost but reduces fan power and operating cost. The tool displays the current friction rate so you can compare against your target (commonly 1 Pa/m) and resize accordingly.
Real-world applications
Commercial and office central HVAC: Pressure drop is summed along the entire path from the air handling unit, through floor-level VAV boxes, into ceiling branches and out at the diffusers using the Darcy-Weisbach formulation that powers this tool. Commercial software such as Trane TRACE 3D Plus, Carrier HAP, and IES VE solves the same equation system internally. Engineers split the network into sections, total the ΔP of the longest run, and verify that the fan's external static pressure rating is comfortably above that figure.
Cleanrooms and data centres: Semiconductor fabs and server halls move enormous volumes of air (50,000 to several million m³/h) and fan power often represents 10-30% of total electricity use. To satisfy LEED or BREEAM targets such as SFP (Specific Fan Power) ≤ 2.5 W/(L/s), designers size ducts generously to hold friction rates below 0.5 Pa/m. Sweeping airflow and size in this tool exposes the sensitivity of SFP to those choices.
Residential and small-building energy-recovery ducting: Attic flex runs in balanced ventilation systems are typically 10-20 m long with many fittings and a rough interior surface—an easy place to underestimate pressure loss. Plugging "Flex / 20 m / 8 fittings" into the tool quickly reveals the >100 Pa total that leaves a small ERV fan gasping for air, which is a textbook failure mode in residential retrofits.
CFD pre-checks and sanity verification: Before running OpenFOAM or ANSYS Fluent on a duct network, engineers use 1D estimates like this tool to bound the expected ΔP. If the CFD answer is an order of magnitude off, that is a strong signal that the boundary condition, turbulence model, or mesh has an issue. Conversely, CFD output cannot be trusted without understanding the underlying Darcy-Weisbach physics.
Common pitfalls and cautions
The first trap is the assumption that the friction factor f is a constant. As the Moody chart shows, f is a function of both Reynolds number and relative roughness ε/Dh, so it shifts whenever air temperature, velocity, or duct size changes. Swamee-Jain becomes inaccurate in the transition region below Re ≈ 10⁵ and in the fully rough region above ε/Dh ≈ 0.05, so production software typically iterates the implicit Colebrook-White equation instead. The tool is accurate enough for routine commercial work (4×10³ < Re < 10⁸, ε/Dh < 0.05) but expect a few percent error on small-diameter flex.
The second pitfall is using a single fitting loss coefficient K of 0.5 for every component. This tool uses K = 0.5 per fitting (a generic elbow value) for quick estimates, but real K-values vary widely: 0.2 for a 45° long-radius elbow, 0.6-1.2 for a 90° short-radius elbow, 1-3 for the branch side of a tee, and 0.4 for a sudden contraction. Detailed sizing should cherry-pick coefficients from the ASHRAE Duct Fitting Database (DFDB) or the SMACNA HVAC Duct Design Manual; otherwise fan external static can be off by 50%.
The third pitfall is stopping at P_fan = Q·ΔP. The tool assumes a 70% fan efficiency, but the full electrical chain also includes motor efficiency (85-95%), drive losses (92% for a V-belt), and inverter losses (97%). Worse, fan efficiency itself drops to 30-40% at low load, so a VAV system rarely operates at its rated point. Comprehensive energy assessment requires hourly annual simulation in EnergyPlus or equivalent.
How to Use
Enter airflow rate in m³/hr (typical range 500–5000 for residential, 10000–50000 for commercial systems)
Input rectangular duct dimensions: height and width in mm (e.g., 200×400 mm for a residential branch duct)
Specify duct length in meters; the simulator calculates hydraulic diameter, velocity, Reynolds number, and friction losses per ASHRAE Duct Fitting Database methods
Review total pressure drop in Pa and required fan power in watts to assess system feasibility
Worked Example
Commercial office supply ductwork: 3000 m³/hr airflow through a 250 mm × 400 mm rectangular duct, 15 m total length. Hydraulic diameter D_h = 308 mm, velocity V = 5.0 m/s, Reynolds number Re ≈ 102,000 (turbulent). Friction loss ≈ 18 Pa/100m yields 2.7 Pa over 15 m. Total pressure drop (friction plus fittings) ≈ 8.5 Pa. Fan power required ≈ 7.1 W per 1000 Pa·m³/hr, here approximately 25.5 W mechanical input.
Practical Notes
Velocity exceeding 5 m/s in main ducts or 3.5 m/s in branches increases friction loss exponentially and noise (regenerated sound above 65 dB); upsize duct cross-section if pressure drop exceeds 10 Pa/100 m
Flexible ductwork adds 30–50% extra friction loss versus smooth rigid galvanized steel; always compare rigid metal to flexible in final design
Fitting losses (elbows, dampers, registers) often equal or exceed straight duct friction; use ASHRAE Duct Fitting Database coefficients and add equivalent length method for tees and transitions
Fan motor oversizing by 10–15% accommodates filter loading and future system modifications without performance degradation