Closed-loop hydronic system — boiler output, flow rate, pump TDH, expansion tank, and terminal units. Animated loop diagram shows temperature gradient from hot supply to cooler return as heat is released at each terminal.
This simulator models a closed-loop hydronic heating system — calculating flow rate, pump TDH, expansion tank size, and terminal unit count from boiler output, supply temperature, and system delta-T. Mechanical engineers use it to develop preliminary sizing for hot water heating systems before detailed pipe-by-pipe calculations.
The fundamental hydronic flow equation is GPM = BTUH / (500 × ΔT), derived from the sensible heat equation Q = m × Cp × ΔT for water where Cp × ρ ≈ 500 BTU/hr·GPM·°F. The simulator uses this to compute system flow from design heat load and supply-return temperature differential. Pump total dynamic head (TDH) is then calculated using the Darcy-Weisbach friction equation applied to the recommended pipe size at the computed flow rate, with an additional 5 psi allowance for fittings and control valves.
The animated loop diagram shows the temperature gradient from hot supply to cooler return as heat is extracted at each terminal unit. Terminal unit count is based on the rated BTUH per unit for the selected type — fin-tube baseboard, fan coil units, radiant panels, or cast iron radiators. The expansion tank is sized using the ASME method from estimated system volume, supply temperature, fill pressure, and relief valve setting.
Hydronic heating systems must comply with ASHRAE 90.1 energy efficiency requirements for boiler efficiency and controls, including outdoor air reset and optimal start. ASME Boiler and Pressure Vessel Code Section I covers power boilers above 15 psig; Section IV covers heating boilers at 15 psig or less. IMC (International Mechanical Code) Chapter 10 governs hydronic piping and boiler installation. For glycol systems, ASHRAE Guideline 12 provides guidance on freeze protection, fluid concentration, and corrosion inhibitor maintenance.
The system ΔT selection significantly affects all downstream sizing. Conventional cast iron boilers use 20°F ΔT (180°F supply, 160°F return) to prevent condensation on the heat exchanger. Condensing boilers achieve peak efficiency at return temperatures below 130°F, which drives designers toward 30–40°F ΔT and lower supply temperatures. Lower ΔT increases system flow rate and pump energy; higher ΔT reduces pump energy but may not be compatible with terminal unit ratings at lower supply temperatures. Always verify terminal unit output ratings at the actual supply water temperature — a fin-tube baseboard rated at 750 BTU/hr at 180°F supply may produce only 400 BTU/hr at 160°F supply.
Adjust the design heat load slider to match the building or zone load in BTU/hr. Set the supply water temperature based on the boiler type and terminal unit requirements. Adjust system ΔT based on the design approach — 20°F for conventional, 30–40°F for condensing. Select the terminal unit type and set the loop length to reflect the piping layout. The simulator updates all results in real time: flow rate, recommended pipe size with velocity check, pump TDH and BHP, number of terminal units required, estimated system volume, and expansion tank size. The velocity check flag alerts when the recommended pipe size produces flow outside the ASHRAE 1.5–4.5 fps range.
Condensing boilers achieve their highest efficiency (up to 95–98% AFUE) when the return water temperature is below 130°F, which causes water vapor in the flue gases to condense and release latent heat. Design return temperatures of 110–120°F (with 30–40°F ΔT and 150–160°F supply) are typical for condensing boiler systems. Low-temperature radiant floor systems at 100–120°F supply achieve even greater condensing savings.
In a variable flow system where pump speed varies with a variable frequency drive (VFD), the affinity laws govern: flow is proportional to speed (Q ∝ N), head is proportional to speed squared (H ∝ N²), and power is proportional to speed cubed (P ∝ N³). Reducing flow to 50% of design requires only 12.5% of design power. This makes VFDs highly effective for part-load energy savings in variable flow primary-secondary hydronic systems.
ASHRAE recommends a minimum velocity of 0.5 fps in hydronic piping to carry entrained air bubbles toward air separators and vents. Below 0.5 fps, air accumulates in high points of the system, causing airlocks that reduce or stop flow through terminal units. The practical design target is 1.5–4 fps to balance air transport against erosion risk in copper pipe, which limits velocity to 4 fps to prevent pitting corrosion from flow-accelerated erosion.
Primary-secondary systems use dedicated primary pumps for each boiler or chiller and a separate secondary pump for distribution, with a common pipe between circuits allowing decoupled operation. Variable primary flow (VPF) eliminates the secondary pump, running a single set of pumps with VFDs while maintaining a minimum flow through each chiller via a bypass. VPF reduces capital cost and pump energy but requires careful minimum flow protection for chillers and sophisticated controls.
Adding glycol reduces heat capacity (a 50% propylene glycol solution has about 80% of water's heat capacity), which increases the required flow rate for the same BTUH at a given ΔT. The viscosity increase also raises friction losses, potentially requiring larger pipes or higher pump head. System flow must be increased by roughly 1/0.8 = 25% for 50% PG mixtures. Also size the expansion tank for the higher volumetric expansion of glycol solutions.
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