17.1 Designing HVAC Systems (Psychrometrics, Enthalpy & Airflow)

17.1 Designing HVAC Systems (Psychrometrics, Enthalpy & Airflow)

Trade Area A (Pre-Installation) is 14 percent of the Class A trade exam and 12 percent of Class B. After you can read the drawings and the load, the outline asks you to design HVAC systems from efficiency, psychrometrics, loads, SEER, pressure-enthalpy diagrams, airflow and duct design, pipe sizing, wiring, controls, and refrigeration. That is one cluster: the air side and the refrigerant side have to close on the same Btu/h. The 2026 Air Conditioning CBT list that supports this chapter is Refrigeration & Air Conditioning Technology, 9th Edition (2021), Florida Building Code — Mechanical, 2023, Florida Building Code — Energy Conservation, 2023, Energy Efficient Building Construction in Florida (2024), SMACNA HVAC Duct Construction Standards Metal and Flexible, 4th Edition (2020), and the Trane Ductulator (1976 or later).

Quick Answer: Sensible supply airflow is (\text{CFM} = \text{Btu/h} \div (1.08 \times \Delta T)). Wet-bulb, dry-bulb, relative humidity, and enthalpy locate the air on the psychrometric chart. A pressure-enthalpy (P-h) diagram locates the refrigerant. Class B designs the same way, but only up to 25 tons of cooling and 500,000 Btu of heating in any one system.

Why this cluster is on the exam

Pearson VUE will hand you a load, a ΔT, a psychrometric state, or a sketch of a vapor-compression cycle and ask whether the air quantity, coil leaving condition, or compressor work is consistent. You design with the listed books open. You do not invent a square-foot-per-ton rule and hope the latent load in Miami agrees.

Class A candidates design without a statutory capacity cap (F.S. 489.105(3)(f)). Class B candidates (F.S. 489.105(3)(g)) use the same psychrometric and airflow math, then stop at 25 tons / 500,000 Btu per system. The Class B outline drops 25–100 ton, over-100 ton, and centrifugal install items; it does not drop the idea that a 40-ton machine still has a P-h diagram. Twenty-five tons is 300,000 Btu/h of cooling. A single 30-ton air loop is Class A work even if two compressors share the cabinet.

Psychrometric properties you must name

Moist air is a mixture of dry air and water vapor. The psychrometric chart plots that mixture so you can read one property from two others. Learn the four names the outline uses, then the two that make the coil math work.

To place a state point you need two independent properties — commonly dry-bulb and wet-bulb, or dry-bulb and RH. Enthalpy is not independent of wet-bulb on the standard sea-level chart; they run almost together. Sensible heat moves dry-bulb at constant humidity ratio (horizontal). Latent heat changes humidity ratio at constant dry-bulb (vertical). A cooling coil does both: the process line slopes toward the apparatus dew point on the saturation curve. Sensible heat ratio (SHR) is sensible divided by total. Florida gulf and Atlantic coasts run high latent fractions; an oversized unit that short-cycles stays on a high-SHR process and does not wring the air.

Approximate enthalpy of moist air is (h \approx 0.24,t_{DB} + W(1061 + 0.445,t_{DB})) in Btu/lb dry air, with (W) in lb moisture per lb dry air. On CBT you will usually read (h) at the intersection of DB and WB rather than derive it.

Airflow from sensible heat — the 1.08 formula

The sensible-heat equation for standard air is:

(Q_s = 1.08 \times \text{CFM} \times \Delta T)

so

(\text{CFM} = \dfrac{Q_s}{1.08 \times \Delta T})

The constant 1.08 is density × specific heat × minutes per hour:

(1.08 \approx 0.075,\text{lb/ft}^3 \times 0.24,\text{Btu/lb·°F} \times 60,\text{min/h})

(Q_s) is sensible Btu/h, not total and not latent. (\Delta T) is the temperature difference of the air stream you are moving — room dry-bulb minus supply dry-bulb for cooling, or supply minus room for heating.

Worked example. A Sarasota living-dining zone has a Manual J sensible cooling load of 21,600 Btu/h. The design room is 75°F dry-bulb. The coil is selected to supply 55°F dry-bulb, so (\Delta T = 20°F).

(\text{CFM} = \dfrac{21{,}600}{1.08 \times 20} = \dfrac{21{,}600}{21.6} = 1{,}000\ \text{cfm})

That zone needs 1,000 cubic feet per minute of supply air at that ΔT. If you drop supply temperature to 50°F ((\Delta T = 25°F)), CFM falls to (21{,}600 / 27.0 = 800) cfm — less air, colder air, more dehumidification if the coil can actually leave at 50°F. If you apply a 400 cfm/ton rule to a 1.8-ton total load, you invent 720 cfm and starve the sensible equation. The 350–450 cfm per ton band is a starting guess, not the exam formula. Florida latent loads often want the lower end of that band so the coil stays cold long enough to condense moisture.

Check the same air against total and latent equations when the outline hands you enthalpy or grains:

The 4.5 factor is (60 \times 0.075). If entering air is 80°F / 67°F WB with enthalpy about 31.5 Btu/lb and leaving air is about 23.2 Btu/lb, (\Delta h \approx 8.3) Btu/lb. At 1,000 cfm, (Q_t = 4.5 \times 1{,}000 \times 8.3 = 37{,}350) Btu/h (about 3.1 tons). Sensible was 21,600; latent is the remainder, which you can also get from grains. If that latent is less than the Manual J latent, the coil selection is wrong even if CFM and SEER2 look pretty.

Pressure-enthalpy diagrams (refrigerant side)

Air enthalpy and refrigerant enthalpy are different charts. A pressure-enthalpy (P-h or Mollier) diagram plots refrigerant pressure (log scale) versus enthalpy (Btu/lb). The saturated-liquid and saturated-vapor curves meet at the critical point. A simple vapor-compression cycle is four processes:

1. Evaporation (4 → 1): nearly horizontal (constant pressure) through the two-phase region, then a little superheat to the right of the dome. Refrigerating effect = (h_1 - h_4) Btu/lb.
2. Compression (1 → 2): pressure and enthalpy rise. Work of compression = (h_2 - h_1).
3. Condensation (2 → 3): nearly horizontal desuperheat, condensation, then subcooling to the left of the saturated-liquid line. Heat rejected = (h_2 - h_3).
4. Expansion (3 → 4): vertical (constant enthalpy) through a TXV, capillary, or electronic expansion valve. (h_4 = h_3).

Coefficient of performance for cooling is refrigerating effect divided by compressor work. Raising condensing pressure (dirty coil, 115°F Florida outdoor, recirculated condenser air in a screen well) moves line 2–3 up, increases (h_2), and cuts COP. Superheat is suction-line temperature minus saturated suction temperature; subcooling is saturated liquid temperature minus liquid-line temperature. Those two numbers keep the P-h state points honest; Chapter 29 drills the gauges. Here you only need to know that design picks a refrigerant, a condensing temperature, an evaporating temperature, and a coil SHR that matches the psychrometric load.

Efficiency, loads, pipe, wire, and controls in the same design

Loads come from Chapter 16: Manual J / ASHRAE 183, then Manual S with expanded data, a 1.15 cooling cap, and AHRI nominal not used for size (FBC Energy R403.7). SEER2 / EER2 is Chapter 15.3: Florida split air conditioners under 45,000 Btu/h sit on a 14.3 SEER2 regional floor. The indoor CFM you just calculated must be a point on the manufacturer's expanded data at the design outdoor dry-bulb and entering wet-bulb, and it must be a point the fan curve can reach at the duct's external static (Section 17.2).

Pipe sizing at design is a friction-and-velocity choice. Refrigerant suction lines need enough velocity for oil return without so much pressure drop that evaporating temperature collapses. Liquid lines are sized so pressure drop will not flash refrigerant before the metering device. Condensate drains are a code minimum 3/4 inch (FBC Mechanical 307) and a slope, not leftover 3/8-inch copper. Hydronic and steam pipe, boilers, and unfired pressure vessels are Class A statutory territory; Class B still reads a friction chart for condensate and for small water coils inside the 25-ton world.

Wiring on this license is limited: dedicated HVAC circuits, single-phase disconnect replace/reconnect, low-voltage control wiring, and a circuit-breaker lock (F.S. 489.105). Design still reads minimum circuit ampacity (MCA) and maximum overcurrent protection (MOCP) off the nameplate so the breaker matches. A three-phase 40-ton rooftop is a Class A equipment decision and usually an electrical-contractor power circuit.

Controls start at the thermostat: interior wall, representative air, not on a supply chase and not on an uninsulated exterior wall in a Florida afternoon sun. Humidity control, staging, and (Class A) pneumatic control piping belong in the design narrative. Class B statutory scope does not include pneumatic control piping.

Florida HVAC scenario

Gulfshore Comfort, a certified Class B shop in Lee County, designs a 2,200-square-foot slab-on-grade house with a Manual J sensible 21,600 Btu/h, latent 7,800 Btu/h, total 29,400 Btu/h. The salesperson wants “2½ tons at 400 cfm/ton” (1,000 cfm) because that is what the truck always carries. The 1.08 equation at a 20°F ΔT also wants 1,000 cfm for the sensible piece — that part accidentally matches — but the expanded-data sheet on the cheap 2.5-ton match shows only 5,500 Btu/h latent at 95°F outdoor / 67°F entering WB. Latent capacity is below the calculated latent load, which FBC Energy R403.7.1.1 forbids. The correct design moves to a match whose expanded data covers both 21,600 sensible and 7,800 latent at the design points, checks SEER2 ≥ 14.3, keeps total capacity inside 1.15 × 29,400 = 33,810 Btu/h, and then sizes duct (Section 17.2) so 1,000 cfm is actually deliverable. Class B is legal: 2.5 tons is under 25. The failure mode is psychrometrics, not the license card.

Traps: (1) Using 1.08 with total Btu/h instead of sensible. (2) Treating 400 cfm/ton as code. (3) Reading wet-bulb as “the humid temperature you set on the thermostat.” (4) Using AHRI nominal tons on the P-h / airflow problem. (5) Designing one 30-ton air loop on a Class B ticket.