12.4 System Commissioning, Testing, Adjusting & Balancing (TAB)

System Commissioning, Testing, Adjusting & Balancing (TAB)

Core TAB Principle: Design engineering on paper is only as good as physical field execution. Testing, Adjusting, and Balancing (TAB) and structured System Commissioning (Cx) per ASHRAE Standard 111, ACCA Standard 5 (ANSI/ACCA QI), and the North Carolina Energy Conservation Code ensure that air and hydronic distribution systems deliver exact design airflow, design thermal capacity, proper indoor air quality (IAQ), and maximum energy efficiency.

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Total External Static Pressure (TESP) & Airflow Determination

Total External Static Pressure (TESP) is the total resistance that the air distribution duct system (supply ducts, return ducts, filters, coils, registers) imposes on the indoor blower fan.

                      [ Return Duct ]
                             |
                             v
                 (P_return) [ Static Probe 1 ] (Entering Air Handler / Before Blower)
                             |
                       [ Air Filter ]
                             |
                      [ Blower Motor ]
                             |
                      [ Heating Elements / Heat Exchanger ]
                             |
                 (P_supply) [ Static Probe 2 ] (Leaving Air Handler / Before A-Coil)
                             |
                     [ Evaporator Coil ]
                             |
                             v
                      [ Supply Plenum ]

1. The TESP Calculation Formula

$\text{TESP} = |P_{\text{supply}}| + |P_{\text{return}}|$

• Measurement Protocol: Using a calibrated dual-port digital manometer and static pressure tips pointing directly into the airstream:
  • $P_{\text{return}}$ is measured in the return plenum between the air filter and the blower inlet (typically a negative pressure, e.g., $-0.25\text{ in. w.c.}$). Drop the negative sign and take the absolute value.
  • $P_{\text{supply}}$ is measured in the supply plenum downstream of the furnace/air handler (typically a positive pressure, e.g., $+0.25\text{ in. w.c.}$). In electric air handlers with internal coils, measure after the coil; in gas furnaces with cased uncased add-on coils, measure between furnace and coil to isolate furnace TESP from coil pressure drop. $\text{TESP} = |-0.25\text{ in. w.c.}| + |+0.25\text{ in. w.c.}| = 0.50\text{ in. w.c.}$

2. Interpreting Manufacturer Blower Performance Tables

Once TESP is measured, the technician references the manufacturer's engineering blower performance chart for the specific motor speed tap or ECM profile to determine delivered CFM:

Sample Blower Performance Table (CFM vs. TESP in inches w.c.)

Standard Rule: Residential cooling systems require $350\text{ to }450\text{ CFM per ton}$ ($400\text{ CFM/ton}$ nominal in North Carolina's humid climate to balance sensible cooling and latent dehumidification).

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Quantitative Airflow Measurement Methods

When blower tables are unavailable or duct geometry requires physical verification, three primary engineering methods are used:

1. Electric Resistance Heat Temperature Rise Method

Because electric resistance strip heaters convert electrical energy into heat at $100%$ efficiency ($1\text{ kW} = 3,412.14\text{ BTU/hr}$), measured temperature rise and electric power provide an exact calculation of system CFM:

$Q_{\text{sensible}} = \text{Volts} \times \text{Amps} \times 3.41214 = \text{kW} \times 3,412.14$

$Q_{\text{sensible}} = 1.08 \times \text{CFM} \times \Delta T$

$\text{CFM} = \frac{\text{Volts} \times \text{Amps} \times 3.41214}{1.08 \times \Delta T} = \frac{\text{kW} \times 3,412.14}{1.08 \times (T_{\text{supply}} - T_{\text{return}})}$

Where $1.08$ is the sensible heat constant for standard air density ($0.075\text{ lb/ft}^3 \times 0.24\text{ BTU/(lb}\cdot^\circ\text{F)} \times 60\text{ min/hr}$). (For high-altitude adjustments, use $c_p \cdot \rho \cdot 60$).

2. Pitot Tube Duct Traverse & Velocity Pressure ($P_v$)

A Pitot tube connected to a digital manometer senses Total Pressure ($P_t$) through its open impact tip and Static Pressure ($P_s$) through its side radial ports. The manometer calculates Velocity Pressure ($P_v$):

$P_v = P_t - P_s$

$\text{Air Velocity } (V\text{ in FPM}) = 4,005 \times \sqrt{P_v}$

$\text{Airflow Volume } (\text{CFM}) = \text{Duct Cross-Sectional Area } (\text{sq ft}) \times V_{\text{average}}$

ASHRAE 111 Duct Traverse Grid Methods:
├── Round Ducts: Log-Tchebycheff Rule or Equal Area Method (Min 12 to 20 traverse points across 2 perpendicular diameters)
└── Rectangular Ducts: Minimum 16 to 64 traverse grid points centered in equal rectangular areas.

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The Governing Fan Laws & Sheave Sizing

The Fan Laws govern all aerodynamic performance changes in commercial and residential centrifugal and axial fans when motor speed, pulley diameter, or duct static pressures are altered.

+-----------------------------------------------------------------------------------------+
|                                 THE GOVERNING FAN LAWS                                  |
+-----------------------------------------------------------------------------------------+
| Fan Law 1 (Airflow vs Speed):      CFM_2 / CFM_1 = RPM_2 / RPM_1                        |
| Fan Law 2 (Static Press vs Speed): SP_2 / SP_1 = (RPM_2 / RPM_1)² = (CFM_2 / CFM_1)²   |
| Fan Law 3 (Power vs Speed):        BHP_2 / BHP_1 = (RPM_2 / RPM_1)³ = (CFM_2 / CFM_1)³ |
+-----------------------------------------------------------------------------------------+

The Critical Cubic Power Relationship (Fan Law 3)

[!WARNING] Fan Law 3 Mandate: Blower fan Brake Horsepower (BHP) increases with the CUBE of the speed ratio ($(\text{RPM}_2 / \text{RPM}_1)^3$). A modest $20%$ increase in airflow ($1.20\times\text{CFM}$) requires an $72.8%$ increase in motor horsepower ($1.20^3 = 1.728$). Attempting to solve airflow deficits by simply speeding up a belt-driven blower will rapidly overload the motor, causing nuisance overcurrent trips or winding burnout unless motor horsepower is upgraded.

Pulley (Sheave) Sizing Formula

For belt-driven commercial fans:

$D_{\text{fan}} \times \text{RPM}_{\text{fan}} = D_{\text{motor}} \times \text{RPM}_{\text{motor}}$

$D_{\text{motor}} = \frac{D_{\text{fan}} \times \text{RPM}_{\text{fan}}}{\text{RPM}_{\text{motor}}} \quad \iff \quad \text{RPM}_{\text{fan}} = \text{RPM}_{\text{motor}} \times \left(\frac{D_{\text{motor}}}{D_{\text{fan}}}\right)$

Where $D_{\text{fan}}$ is pitch diameter of fan pulley and $D_{\text{motor}}$ is pitch diameter of motor pulley.

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North Carolina Energy Code & Mechanical Code Commissioning Mandates

Under the North Carolina Energy Conservation Code (NCECC) and NC Mechanical Code (NCMC Chapter 4 & 6), contractors must execute verified commissioning:

NC Building Code Commissioning Checkpoints:
├── 1. Duct Leakage Pressurization Testing (Total Leakage ≤ 5 CFM25 per 100 sq ft CFA
│      or Leakage to Outdoors ≤ 4 CFM25 per 100 sq ft CFA when tested at 25 Pascals).
├── 2. Ventilation Verification (ASHRAE 62.2 / NCMC Chapter 4 mechanical outdoor air rates).
├── 3. Economizer Functional Testing (Air-side economizers must modulate dampers to 100% 
│      outdoor air when ambient conditions permit free cooling, with verified sensor calibration).
└── 4. Refrigerant Charge Verification (Subcooling verified within ± 3°F of nameplate on TXV 
       systems; Superheat within ± 5°F of manufacturer target chart on fixed orifice systems).

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Step-by-Step Worked Technical Examples

Example 1: Electric Heat Temperature Rise Airflow Calculation

Problem: A technician commissions a 3.5-ton split-system heat pump with an electric resistance strip heater. With only the electric heat operating, the technician measures:

• Operating Line Voltage: $240.0\text{ V}$
• Total Heater Element Current: $40.5\text{ A}$
• Return Air Temperature entering air handler: $68.0^\circ\text{F}$
• Supply Air Temperature leaving air handler: $92.5^\circ\text{F}$

Calculate: (1) Total electric heat output in $\text{Watts}$ and $\text{BTU/hr}$, (2) Air temperature rise ($\Delta T$), and (3) Actual delivered system airflow in $\text{CFM}$.

Solution:

1. Calculate Heat Output in Watts and BTU/hr: $P = V \times I = 240.0\text{ V} \times 40.5\text{ A} = 9,720.0\text{ Watts} = 9.72\text{ kW}$ $Q_{\text{sensible}} = 9,720\text{ W} \times 3.41214\text{ BTU/Wh} = 33,166.0\text{ BTU/hr}$

2. Calculate Air Temperature Rise ($\Delta T$): $\Delta T = T_{\text{supply}} - T_{\text{return}} = 92.5^\circ\text{F} - 68.0^\circ\text{F} = 24.5^\circ\text{F}$

3. Calculate System Delivered Airflow (CFM): $\text{CFM} = \frac{Q_{\text{sensible}}}{1.08 \times \Delta T} = \frac{33,166.0\text{ BTU/hr}}{1.08 \times 24.5^\circ\text{F}} = \frac{33,166.0}{26.46} = 1,253.4\text{ CFM}$

   • Normalized per ton: $1,253.4\text{ CFM} / 3.5\text{ tons} = 358.1\text{ CFM/ton}$ (acceptable for dehumidification in NC).

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Example 2: Fan Laws and Sheave Sizing for Commercial Air Handler

Problem: A commercial belt-drive air handler in Greensboro, NC currently delivers $3,200\text{ CFM}$ at a measured fan speed of $650\text{ RPM}$, with a measured static pressure of $0.80\text{ in. w.c.}$ and motor brake horsepower of $1.80\text{ BHP}$. The design specification requires increasing airflow to $3,800\text{ CFM}$.

Calculate: (1) New fan RPM required, (2) New system static pressure, (3) New brake horsepower required, and (4) The required motor pulley diameter if the motor spins at $1,725\text{ RPM}$ and the fan pulley diameter is $9.0\text{ inches}$.

Solution:

1. Calculate New Fan RPM (Fan Law 1): $\text{RPM}_2 = \text{RPM}_1 \times \left(\frac{\text{CFM}_2}{\text{CFM}_1}\right) = 650\text{ RPM} \times \left(\frac{3,800}{3,200}\right) = 650 \times 1.1875 = 771.9\text{ RPM}$

2. Calculate New Static Pressure (Fan Law 2): $\text{SP}_2 = \text{SP}_1 \times \left(\frac{\text{CFM}_2}{\text{CFM}_1}\right)^2 = 0.80 \times (1.1875)^2 = 0.80 \times 1.4102 = 1.128\text{ in. w.c.}$

3. Calculate New Brake Horsepower (Fan Law 3): $\text{BHP}_2 = \text{BHP}_1 \times \left(\frac{\text{CFM}_2}{\text{CFM}_1}\right)^3 = 1.80\text{ BHP} \times (1.1875)^3 = 1.80 \times 1.6746 = 3.01\text{ BHP}$ (Note: The existing $2.0\text{ HP}$ motor must be upgraded to a $3.0\text{ HP}$ or $5.0\text{ HP}$ motor to prevent overload burnout!)

4. Calculate Required Motor Pulley Diameter ($D_{\text{motor}}$): $D_{\text{motor}} = \frac{D_{\text{fan}} \times \text{RPM}_{\text{fan}}}{\text{RPM}_{\text{motor}}} = \frac{9.0\text{ in.} \times 771.9\text{ RPM}}{1,725\text{ RPM}} = \frac{6,947.1}{1,725} = 4.03\text{ inches}$

   • Install an adjustable $3.8" - 4.4"$ pitch diameter motor sheave adjusted to $4.03"$.

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Example 3: Pitot Tube Duct Traverse Air Velocity & Volume

Problem: A technician traverses a $24" \times 16"$ rectangular supply trunk duct with a Pitot tube. The average of 16 velocity pressure readings is $P_{v,\text{avg}} = 0.09\text{ in. w.c.}$

Calculate: (1) Average duct air velocity in Feet Per Minute (FPM), (2) Duct cross-sectional area in square feet, and (3) Total volumetric airflow in CFM.

Solution:

1. Calculate Average Air Velocity: $V = 4,005 \times \sqrt{P_v} = 4,005 \times \sqrt{0.09} = 4,005 \times 0.30 = 1,201.5\text{ FPM}$

2. Calculate Duct Cross-Sectional Area: $\text{Area} = \frac{24\text{ in.} \times 16\text{ in.}}{144\text{ sq in./sq ft}} = \frac{384}{144} = 2.667\text{ sq ft}$

3. Calculate Total Volumetric Airflow (CFM): $\text{CFM} = \text{Area} \times V = 2.667\text{ sq ft} \times 1,201.5\text{ FPM} = 3,204.0\text{ CFM}$