8.3 Centrifugal & Axial Fans, Affinity Laws & Airflow

Centrifugal Fan Blade Configurations & Operational Characteristics

Centrifugal fans (blowers) utilize a rotating impeller mounted inside a scroll-shaped volute housing. Air enters the fan axially through the center inlet eye, turns 90°, and is accelerated radially outward off the blade tips by centrifugal force into the expanding scroll housing, converting kinetic velocity pressure into static pressure.

Centrifugal fan performance is governed primarily by blade curvature relative to impeller rotation:

1. Forward-Curved Blades (Squirrel Cage)

• Geometry: Features numerous small, closely spaced blades curved forward in the direction of rotation.
• Performance: High volume air delivery at low operating speeds and low static pressures (< 4 in. w.g. / 1 kPa).
• Power Curve: Overloading characteristic. As static pressure resistance drops, airflow increases rapidly, causing brake horsepower (BHP) demand to surge continuously. If operated with open ducts, the drive motor will overload and trip.
• Limitation: Shallow blade pockets collect dirt easily; restricted to clean indoor HVAC applications.

2. Radial Blades (Paddle Wheel)

• Geometry: Flat, straight radial blades extending outward like spokes on a wheel.
• Performance: Medium pressure (up to 30+ in. w.g. / 7.5 kPa), moderate efficiency (55–65%).
• Key Advantage: Self-Cleaning Geometry. High structural strength and open blade design prevent material buildup. Ideal for heavy industrial pneumatic conveying (dust collection, wood sawdust, paper trim, foundry sand, and corrosive gas streams).

3. Backward-Inclined (BI) & Backward-Curved (BC) Blades

• Geometry: Flat or curved blades inclined backward, away from the direction of rotation.
• Performance: High efficiency (75–80%), suitable for high-volume, medium-static pressure industrial applications.
• Power Curve: Non-Overloading Characteristic. BHP rises to a peak near the maximum efficiency point and then levels off or decreases as airflow increases. Drive motors cannot be overloaded by opening duct dampers.

4. Airfoil Blades

• Geometry: Backward-curved blades engineered with a double-surfaced aerodynamic airfoil cross-section.
• Performance: Highest energy efficiency (> 85%) and quietest acoustic profile of all fan designs.
• Limitation: Hollow airfoil blades are susceptible to internal moisture corrosion and abrasive wear. If particle-laden air enters a damaged airfoil, dust accumulates inside the hollow core, causing severe rotor unbalance.

Axial Fan Types & Applications

Axial fans move air parallel to the rotation axis of the fan shaft. Impeller blades accelerate air by aerodynamic lift, similar to an airplane wing. Axial fans are classified into three distinct structural categories based on housing and vane geometry:

   PROPELLER FAN               TUBE-AXIAL FAN               VANE-AXIAL FAN
 (No Duct Housing)         (Cylindrical Housing)        (Guide Vanes Included)
      \  |  /                   +---------+                  +----+---+----+
       \ | /                    |  \ | /  |                  | > | \|/ | < | (Guide
   =====(O)=====            ====|===(O)===|====          ====|===|=(O)=|===|==== Vanes)
       / | \                    |  / | \  |                  | > | /|\ | < |
      /  |  \                   +---------+                  +----+---+----+
  Low SP (<0.5")               Med SP (<2.0")              High SP (Up to 10")

Stationary Guide Vane Aerodynamics

In a standard axial fan, air leaving the impeller blades possesses a high rotational velocity swirl component (turbulent helical motion), which wastes energy. In a vane-axial fan, stationary air-foil guide vanes positioned immediately behind the rotor straighten the swirling air stream. This converts rotational velocity energy into useful static pressure, increasing fan mechanical efficiency up to 80–85%.

Static vs. Velocity vs. Total Pressure Relationships

Air flowing through industrial ventilation ducts possesses energy in two distinct mechanical states: potential energy (static pressure) and kinetic energy (velocity pressure). Fluid mechanics dictates that Total Pressure (P_T) is the algebraic sum of Static Pressure (P_S) and Velocity Pressure (P_V):

$P_T = P_S + P_V$

1. Static Pressure (P_S or SP)

• Definition: The potential pressure exerted equally in all directions against the interior walls of a duct, independent of air movement.
• Characteristics: Can be positive (downstream of blower discharge, pushing outward on duct walls) or negative (upstream of fan inlet suction, collapsing duct walls inward).
• Measurement: Measured perpendicular to duct airflow using a static pressure tap connected to a water manometer or Magnehelic gauge, expressed in inches of water gauge (in. w.g.) or Pascals (Pa).

2. Velocity Pressure (P_V or VP)

• Definition: The kinetic energy of air moving in a specific direction along the duct.
• Characteristics: ALWAYS POSITIVE. Velocity pressure cannot be negative.
• Air Velocity Equation: Velocity pressure is directly related to air speed (V) in feet per minute (FPM):

$V = 4005 \times \sqrt{VP} \quad \text{(where } V \text{ is in FPM and } VP \text{ is in in. w.g. at standard air density)}$

Pitot Tube Measurements

A Pitot-static tube is a dual-concentric probe inserted into an airstream to capture duct pressure differentials:

• Impact Tip: Faces directly into the oncoming airstream to measure Total Pressure (P_T).
• Side Ports: Positioned parallel to flow to measure Static Pressure (P_S).
• Differential Connection: Connecting both ports across a manometer subtracts static pressure from total pressure, reading Velocity Pressure (P_V = P_T - P_S) directly.

Fan Affinity Laws & Formulas

The Fan Affinity Laws are fundamental mathematical relationships that govern how changes in fan rotational speed (RPM) or impeller diameter affect volumetric flow rate (CFM), static pressure (SP), and brake horsepower (BHP). When fan impeller diameter is held constant, performance varies strictly with shaft RPM:

The Three Fan Affinity Laws

$\text{Law 1 (Airflow CFM):} \quad \frac{CFM_2}{CFM_1} = \frac{RPM_2}{RPM_1}$

$\text{Law 2 (Static Pressure SP):} \quad \frac{SP_2}{SP_1} = \left(\frac{RPM_2}{RPM_1}\right)^2$

$\text{Law 3 (Brake Horsepower BHP):} \quad \frac{BHP_2}{BHP_1} = \left(\frac{RPM_2}{RPM_1}\right)^3$

Worked Calculation Example

An industrial exhaust fan operating at 1000 RPM delivers 5,000 CFM against a static pressure of 2.0 in. w.g. and draws 3.0 BHP. The plant engineer installs a larger motor pulley to increase fan speed to 1200 RPM. Calculate the new CFM, static pressure, and horsepower demand.

1. New Airflow (CFM_2): $CFM_2 = 5000 \times \left(\frac{1200}{1000}\right) = 5000 \times 1.20 = \mathbf{6,000 \text{ CFM}}$

2. New Static Pressure (SP_2): $SP_2 = 2.0 \times \left(\frac{1200}{1000}\right)^2 = 2.0 \times (1.20)^2 = 2.0 \times 1.44 = \mathbf{2.88 \text{ in. w.g.}}$

3. New Power Requirement (BHP_2): $BHP_2 = 3.0 \times \left(\frac{1200}{1000}\right)^3 = 3.0 \times (1.20)^3 = 3.0 \times 1.728 = \mathbf{5.184 \text{ BHP}}$

CRITICAL MILLWRIGHT WARNING: Notice that a 20% increase in RPM results in a 72.8% increase in required horsepower (3.0 → 5.18 BHP). Increasing fan speed without checking drive motor nameplate full-load amperage will burn out motor windings!

Fan Field Balancing & Vibration Diagnostics

Fans are susceptible to high vibration due to high rotational speeds and harsh operating environments. Systematic diagnostic troubleshooting is required:

Primary Causes of Fan Vibration

1. Mechanical Unbalance (1X RPM): The rotational center of mass does not coincide with the geometric center of the shaft axis. Caused by uneven dust/ash buildup on blades, localized blade erosion, loss of balance weights, or thermal warping.
2. Shaft Misalignment (1X & 2X RPM): Angular or offset misalignment between the motor shaft and fan shaft across a flexible coupling.
3. Aerodynamic Stall & Surge: Operating a fan outside its design curve (e.g., choked intake dampers or blocked filters) causes airflow separation off blades, generating violent low-frequency pressure pulsations.
4. Structural Mechanical Looseness: Loose hold-down bolts, cracked foundation pedestals, or worn pillow-block bearing seats.

Single-Plane Field Balancing Vector Procedure

When vibration spectrum analysis confirms a dominant 1X RPM sinusoidal vibration peak caused by mass unbalance, a millwright performs single-plane field balancing using a portable vibration analyzer and stroboscopic tachometer:

                          SINGLE-PLANE VECTOR BALANCE DIAGRAM

                                     Original Vector O (Amplitude @ Phase Angle)
                                    /
                                   / 
                                  /   + Trial Weight Vector T
                                 /   /
                                +---+---------------------> Resultant Vector O+T
                                    \
                                     \ Correction Weight Mass = W_trial * (O / T)
                                      \ Position Angle = Shift by (180 deg - Phase)

1. Initial Run (Original Vector O): Measure baseline 1X vibration amplitude (e.g., 0.60 in/sec peak) and phase angle (e.g., 45°).
2. Trial Run (Vector O+T): Stop the fan, attach a known trial weight (W_trial, e.g., 20 grams) at a precise angular radius (e.g., 0°). Re-run fan and measure new amplitude and phase angle.
3. Vector Subtraction: Subtract Vector O from Vector O+T to isolate the pure trial weight response vector (T).
4. Calculate Correction Weight: Compute required correction weight mass: $W_{\text{correction}} = W_{\text{trial}} \times \left(\frac{|O|}{|T|}\right)$
5. Mount Correction Weight: Place W_correction on the fan wheel at the calculated angular position opposite the unbalance vector, then remove the trial weight.