Pipe Condensation Calculator

• Ambient intake temperature, RH and altitude. The air the blower draws in. RH × the saturation vapour pressure at the intake temperature (Buck equation) gives the vapour partial pressure, and with the local barometric pressure P_atm (standard atmosphere at the site altitude) the humidity ratio W — the moisture the air carries. The blower changes temperature and pressure, not W.
• Blower outlet temperature and discharge pressure. The pipe inlet state: the air temperature after compression (typically intake + 25–35 K for a lobe or turbo blower) and the gauge discharge pressure, added to P_atm. At line pressure the same W means a higher vapour partial pressure, so the dew point at line pressure is above the intake dew point — the pipe inlet RH shown is this partial pressure over the saturation pressure at the blower-outlet / pipe-inlet temperature.
• Flow rate. On the normal basis (0 °C, 1.013 bar) and converted to a mass flow with 1.293 kg/m³ — the dry-air normal density standing in for the moist intake air, an error of 1–2 % at these humidities. Mass flow is what the energy and water balances need; velocity and Reynolds number follow from it at the line density.
• Pipe material, schedule, size and length. The schedule database gives the inner and outer diameters; the wall's own resistance is neglected — a few mm of steel is about 0.0002 m²·K/W against roughly 0.08 m²·K/W for the inner air film (plastic walls are 0.01–0.03 m²·K/W, still an order below). The temperature difference to ambient decays exponentially with length (the NTU form), not linearly.
• Surface, insulation and wind. Bare: the pipe outer surface loses heat with a fixed outer coefficient by wind class (10 / 20 / 40 W/(m²·K)). Insulated: a cylindrical conduction resistance for the chosen material and thickness is added in series, and the outer film acts on the insulation surface.
• Heat transfer and outlet temperature. Inside: Dittus–Boelter for turbulent flow (Re > 4,000), Nu = 3.66 laminar, interpolated between. Air density, viscosity and conductivity are evaluated at the mean of inlet and outlet temperature, in two passes so the mean uses the computed outlet; the Re, h_i and U reported are the second-pass values. The overall U on the pipe outer diameter feeds the single-stream NTU relation T_out = T_amb + (T_in − T_amb)·exp(−U·π·D·L / (ṁ·c_p)), which is what the outlet temperature is.
• Dew point at line pressure and the verdict. The temperature at which the air in the pipe is saturated at its pressure — evaluated at constant line pressure and constant humidity ratio, so it is one figure for the whole run, not something that drifts along the pipe. If the outlet temperature stays above it, no condensation; if it falls below, the air leaves saturated at the outlet temperature and the difference in W is the water dropped out along the run.
• Condensate, heat loss and flow figures. Condensate in L/h and L/day (1 kg ≈ 1 L), the total heat lost to ambient, and the velocity, Reynolds number, regime, inner coefficient h_i and overall U so the heat-transfer side can be checked against your own numbers.

• Single-stream NTU heat exchange with constant c_p (1,005 J/(kg·K)) and the ambient as an infinite sink at one temperature along the whole run; air properties (Sutherland viscosity, linear conductivity, ideal-gas density at line pressure) are taken at the mean of inlet and outlet in two passes.
• Inside: Dittus–Boelter with the Pr^0.4 exponent for Re > 4,000 (the cooling form, Pr^0.3, would put h_i about 4 % lower — retained for simplicity because the outer film, not h_i, governs the uncertainty), Nu = 3.66 for laminar flow, and a linear blend across 2,300–4,000 that is an engineering interpolation, not a published correlation; entrance effects are ignored.
• Outside: a fixed film coefficient by wind class (10 / 20 / 40 W/(m²·K)) that stands for convection plus radiation together — the still-air value corresponds to natural convection plus radiation from a high-emissivity (ε ≈ 0.9) surface; a bright bare stainless pipe radiates less and would run a little warmer. The pipe wall's resistance is neglected. For surface temperature, emissivity and wind resolved properly use the insulation calculator.
• Dry cooling: the latent heat released when vapour condenses on the wall is not fed back into the air's energy balance, and the condensate film is assumed drained rather than insulating the wall. In the worked example the latent release is about 40 % of the sensible loss, so when condensation is heavy the calculator over-cools the air noticeably — treat its condensate as an upper bound; when the outlet sits just below the dew point the effect is small.
• Constant line pressure along the run (no friction pressure drop) and the dew point evaluated at that pressure; water is conserved from intake to pipe inlet — the blower adds none and an aftercooler, if any, must be represented by entering its outlet temperature.
• Vapour pressure by the Buck (1981) equation over liquid water; humidity ratio from the ideal-gas relation W = 0.622 p_w/(P − p_w) without the pressure enhancement factor (under 0.2 % at blower pressures).
• Steady state at one ambient. Condensate per day multiplies the hourly rate by 24; the design case for drains and insulation is the coldest ambient at the highest intake humidity, not the average day.

The outer-surface model here is the simplified fixed-h form; the insulation calculator resolves the surface temperature with Churchill–Chu convection and radiation. The SimuPipe solver's own accuracy for the air network is documented on the validation page.

Correlations and sources behind this page:

• Incropera, DeWitt, Bergman and Lavine. Fundamentals of Heat and Mass Transfer, Wiley — the NTU relation, the cylindrical resistance network and the internal-flow correlations.
• Dittus, F. W. and Boelter, L. M. K. (1930). "Heat transfer in automobile radiators of the tubular type." University of California Publications in Engineering, 2, 443–461; reprinted Int. Commun. Heat Mass Transfer 12(1), 3–22 (1985). doi:10.1016/0735-1933(85)90003-X — the turbulent internal-flow Nusselt correlation.
• Buck, A. L. (1981). "New Equations for Computing Vapor Pressure and Enhancement Factor." Journal of Applied Meteorology, 20(12), 1527–1532. doi:10.1175/1520-0450(1981)020<1527:NEFCVP>2.0.CO;2 — the saturation vapour pressure equation.
• ASHRAE (2021). Handbook — Fundamentals, Chapter 1: Psychrometrics — humidity ratio, dew point and the psychrometric relations.
• Sutherland, W. (1893). "The viscosity of gases and molecular force." Philosophical Magazine, 36, 507–531 — the temperature dependence of air viscosity used for the Reynolds number.
• Water Environment Federation. Design of Water Resource Recovery Facilities (MOP 8), 6th ed. — aeration blower and air-piping design practice, including condensate drainage.