Industrial Exhaust Fan Sizing: From System Curve to Fan Selection
Industrial exhaust fan sizing is the process of selecting a fan that delivers the design CFM at the required static pressure by matching the fan performance curve to the system resistance curve at the design operating point. The system resistance curve — a parabola defined by ΔP = k × Q² — represents the pressure the fan must overcome at each flow rate, and the fan performance curve represents the pressure the fan can generate at each flow rate for a given speed. The intersection of these two curves is the operating point — the CFM and static pressure at which the fan and system are in equilibrium. Selecting a fan without knowing the system resistance curve results in a fan that either delivers less CFM than required (the system pressure is higher than the fan curve at the design CFM) or operates in an unstable region (the operating point is near or past the fan’s peak pressure point). This guide covers system resistance curve calculation, fan curve interpretation, operating point determination, fan laws for speed and density changes, VAV system fan selection, series and parallel fan operation, motor sizing including drive losses, and a complete worked example for a 15,000 CFM chemical scrubber system. For overall system duct design, see our industrial exhaust system design guide. For fan type selection, see our FRP blower selection guide.
Key Takeaways
- The fan operating point is the intersection of the fan performance curve and the system resistance curve — selecting a fan by CFM alone without considering the system curve is the most common fan sizing error. The system curve must be calculated from the duct layout and equipment pressure drop before a fan can be selected.
- The system resistance curve follows ΔP = k × Q² — k is calculated from one known pressure-flow point and used to plot the full curve. Doubling the CFM increases the system pressure by four times, and the fan power by eight times.
- Fan speed changes follow the fan laws: CFM ∝ RPM, pressure ∝ RPM², power ∝ RPM³ — increasing fan speed by 10 percent increases CFM by 10 percent, pressure by 21 percent, and power by 33 percent. Verify the motor has capacity before increasing fan speed.
- VAV fan selection requires the fan curve to remain stable from 100 percent down to 40 percent speed — below 40 percent speed, most centrifugal fans enter stall where flow becomes erratic and vibration increases. Size the fan so the minimum expected CFM is at least 40 percent of the fan’s design CFM at full speed.
- Motor sizing for belt-driven fans must include the 4 to 8 percent belt transmission loss — a fan requiring 25 HP at the shaft needs a motor rated for 27 to 33 HP. Select the next standard motor size (30 HP for this example).
System Resistance Curve Calculation
The system resistance curve is the foundation of industrial exhaust fan sizing. It represents the total static pressure that the fan must overcome to move the design CFM through the duct system. The curve follows a parabolic relationship: ΔP = k × Q², where ΔP is the system static pressure loss in in. W.G., Q is the airflow in CFM, and k is a constant derived from the system design. The constant k is calculated by first determining the total system pressure loss at the design CFM using the duct friction loss calculation method (see the industrial exhaust system design guide for the complete pressure loss calculation procedure). Once the total system pressure loss at the design CFM is known, k = ΔP_design / (Q_design)². For a system with 9.0 in. W.G. total pressure loss at 10,000 CFM, k = 9.0 / (10,000)² = 9.0 × 10⁻⁸. The system curve at any flow Q is then ΔP = 9.0 × 10⁻⁸ × Q². At 8,000 CFM, the system pressure loss is 9.0 × 10⁻⁸ × 8,000² = 5.76 in. W.G. At 12,000 CFM, the pressure loss is 9.0 × 10⁻⁸ × 12,000² = 12.96 in. W.G. — the system pressure increases with the square of flow, so a 20 percent increase in CFM from the design point increases the system pressure by 44 percent.
The system curve calculation assumes that the system resistance elements — duct friction, fitting losses, equipment pressure drops (scrubber, filters, mist eliminators), and stack losses — all follow the square law relationship. For equipment with pressure drop that varies with flow, the square law is accurate within ±5 percent for most industrial exhaust components. For equipment with a fixed pressure drop regardless of flow — such as a constant-pressure regulator or a backdraft damper that opens fully at a set pressure — the system curve deviates from the square law at low flow, and the fan sizing must account for the fixed pressure component separately. In practice, for chemical exhaust systems, the square law approximation is accurate enough for fan selection because the scrubber pressure drop (typically 3.0 to 8.0 in. W.G. at design flow) dominates the system curve, and scrubber pressure drop follows the square law closely. The system curve must be calculated for every fan selection to verify that the operating point falls within the fan’s stable performance range. A fan selected without the system curve can operate at a CFM that is 15 to 30 percent below the design CFM if the actual system pressure is higher than the assumed value. For the complete duct pressure loss calculation procedure, see our industrial exhaust system design guide.
Fan Performance Curves and Operating Point
Fan performance curves are published by manufacturers for each fan model and show the relationship between CFM and static pressure at various speeds. A typical performance curve for a backward-curved FRP centrifugal fan shows static pressure decreasing as CFM increases — the curve starts at the shutoff point (maximum pressure at zero flow) and drops to the free-air point (zero pressure at maximum flow). The usable portion of the curve is between 40 percent and 100 percent of the free-air CFM — below 40 percent, the fan operates in the stall region where the airflow separates from the impeller blades, causing fluctuating pressure, vibration, and potential motor overload. The fan’s peak efficiency point is typically at 60 to 80 percent of free-air CFM, and the fan should be selected to operate within ±15 percent of the peak efficiency CFM for optimal energy consumption. The fan curve is certified by AMCA (Air Movement and Control Association) for fans that carry the AMCA seal — the certified curve guarantees that the fan delivers within ±5 percent of the published CFM at the stated static pressure.
The operating point is found by plotting the system resistance curve and the fan performance curve on the same axes and finding their intersection. For a system requiring 10,000 CFM at 9.0 in. W.G., the system curve ΔP = 9.0 × 10⁻⁸ × Q² is plotted from 0 to 12,000 CFM. The fan performance curve for a 630 mm backward-curved FRP centrifugal fan at 1,200 RPM is plotted on the same graph. The intersection of the two curves is the operating point — if the intersection is at 10,000 CFM and 9.0 in. W.G., the fan selection is correct. If the intersection falls below 10,000 CFM — for example, 8,500 CFM at 6.5 in. W.G. — the fan is undersized, and either a larger fan or a higher speed is required. The operating point must also be checked against the fan’s safe operating range — the AMCA-defined Class I, II, or III limits based on the fan wheel design and material stress limits. For FRP fans, the maximum safe tip speed is 45 m/s for backward-curved impellers and 84 m/s for compression-molded radial impellers. An FRP fan with a 630 mm impeller at 1,200 RPM has a tip speed of π × 0.630 × 1,200 / 60 = 39.6 m/s — within the safe range. At 1,600 RPM, the tip speed is 52.8 m/s, exceeding the 45 m/s limit for backward-curved FRP impellers. This tip speed limit must be verified as part of the fan selection process — exceeding it causes the impeller to delaminate at the blade root within 6 to 12 months.
Fan Laws: Speed, Diameter, and Density Relationships
The fan affinity laws predict how a fan’s performance changes with speed, impeller diameter, and air density. These three laws apply to all centrifugal fans and must be used in industrial exhaust fan sizing whenever the fan speed or operating conditions differ from the manufacturer’s published rating. Law 1 — flow varies directly with speed: Q₂ = Q₁ × (RPM₂ / RPM₁). A fan delivering 10,000 CFM at 1,200 RPM delivers 11,000 CFM at 1,320 RPM — a 10 percent speed increase provides a 10 percent flow increase. Law 2 — pressure varies with the square of speed: P₂ = P₁ × (RPM₂ / RPM₁)². The same fan at 1,320 RPM generates 9.0 × (1,320/1,200)² = 9.0 × 1.21 = 10.9 in. W.G. — the 10 percent speed increase provides a 21 percent pressure increase. Law 3 — power varies with the cube of speed: HP₂ = HP₁ × (RPM₂ / RPM₁)³. The same fan requiring 25 HP at 1,200 RPM requires 25 × (1,320/1,200)³ = 25 × 1.33 = 33.3 HP at 1,320 RPM — the 10 percent speed increase requires a 33 percent power increase. The cube relationship between speed and power is the most critical practical implication of the fan laws: a 10 percent speed increase requires a motor with 33 percent more capacity. A fan that appears to have margin at 25 HP (selected with a 30 HP motor operating at 83 percent load) will be overloaded if the speed is increased by 10 percent (the new requirement of 33.3 HP exceeds the 30 HP motor rating).
The fan laws for impeller diameter change apply when comparing geometrically similar fans of different diameters: flow varies with the cube of diameter (Q ∝ D³), pressure varies with the square of diameter (P ∝ D²), and power varies with the fifth power of diameter (HP ∝ D⁵). A 700 mm fan operating at the same speed as a 630 mm fan delivers (700/630)³ = 1.37 times the flow and requires (700/630)⁵ = 1.82 times the power — a moderate 11 percent diameter increase provides 37 percent more flow but requires 82 percent more motor power. The density fan law applies when the fan operates at a different air temperature or altitude than the standard condition (20°C, sea level, 1.2 kg/m³): pressure varies directly with density (P ∝ ρ), and power varies directly with density (HP ∝ ρ). A fan operating at 80°C exhaust temperature with air density of 1.0 kg/m³ generates 9.0 × (1.0/1.2) = 7.5 in. W.G. at the design CFM — the fan must be selected for the actual operating density, not the standard density. The density correction for altitude is 3.5 percent pressure loss per 1,000 ft elevation — a fan at 5,000 ft elevation generates 9.0 × (1 – 5 × 0.035) = 9.0 × 0.825 = 7.4 in. W.G. These density effects are often overlooked in industrial exhaust fan sizing, resulting in a fan that delivers 15 to 25 percent less static pressure than the system requires at the operating condition. For the fan type that is most suitable for chemical exhaust applications, see our FRP blower selection guide.
Fan Selection for Variable Flow Systems
Variable flow systems — where the exhaust CFM varies with hood sash position, process conditions, or occupancy — require fan selection that accounts for the full operating range from maximum to minimum CFM. In industrial exhaust fan sizing for VAV systems, the fan must operate stably at the minimum expected CFM, which is typically 40 to 60 percent of the maximum design CFM. The fan is selected at the maximum CFM and static pressure (the design point), and the minimum CFM is set by the fan’s stall limit — generally 40 percent of the fan’s free-air CFM. A fan selected at 10,000 CFM from a fan with 15,000 CFM free-air capacity reaches its stall limit at 0.40 × 15,000 = 6,000 CFM. If the VAV system requires the fan to operate at 4,000 CFM during low-load conditions, the fan stalls at 6,000 CFM — the airflow becomes erratic, the motor amperage fluctuates, and the fan bearings experience accelerated wear. To avoid stall in VAV systems, select a fan with a wide stable operating range (turndown ratio of 4:1 or higher) or use a fan array where individual fans can be shut off at low load to keep the remaining fans above the stall limit.
VFD (variable frequency drive) selection for VAV fan systems must account for the fan’s speed range and the motor’s cooling characteristics. Standard TEFC (totally enclosed fan-cooled) motors are cooled by a shaft-mounted fan that delivers cooling airflow proportional to motor speed — at 40 percent speed, the cooling fan delivers only 40 percent of its rated airflow, and the motor may overheat if operating at full load torque. For VAV fan applications where the motor operates below 60 percent speed for extended periods, specify an inverter-duty motor with a separate constant-speed cooling fan or a TENV (totally enclosed non-ventilated) motor rated for the variable-speed operating range. The VFD itself must be located in a conditioned space or a NEMA 4X enclosure rated for the installation environment — a VFD installed in an unconditioned chemical plant area fails within 6 to 12 months from heat and corrosive gas exposure. The VFD adds 3 to 5 percent harmonic distortion to the electrical supply, which must be verified against the facility’s power quality standards — for facilities with sensitive electronics, a 3 percent line reactor or a 5 percent DC link choke on the VFD reduces harmonics to acceptable levels. The VFD energy savings in VAV systems are typically 20 to 35 percent compared to constant-speed fan operation with inlet guide vanes or discharge dampers. The payback period for the VFD investment in a continuously operating exhaust fan above 20 HP is 12 to 24 months.
Multiple Fans: Series and Parallel Operation
Multiple fans in series or parallel are used when a single fan cannot meet the system pressure or flow requirement. Fans in series — where the discharge of one fan feeds the inlet of the next — add their pressure capabilities at the same CFM. Two fans in series, each rated at 10,000 CFM and 5.0 in. W.G., deliver 10,000 CFM at 10.0 in. W.G. total static pressure. Series fan arrangements are used for high-pressure applications such as deep bed scrubbers above 15.0 in. W.G. or long duct runs above 500 ft. In industrial exhaust fan sizing, series fans must be selected so that the second fan’s inlet pressure does not exceed its rated maximum — the first fan’s discharge pressure becomes the second fan’s inlet pressure, and if the second fan is not designed for the elevated inlet pressure, the shaft seal leaks and the bearings fail. Series fan installations in chemical exhaust systems are rare because a single FRP high-pressure blower (up to 15,000 Pa / 60 in. W.G.) covers 90 percent of chemical exhaust applications. Specify series fans only when the required system pressure exceeds the maximum single-stage FRP fan rating.
Fans in parallel — where multiple fans discharge into a common header — add their flow capabilities at the same static pressure. Two fans in parallel, each rated at 10,000 CFM and 5.0 in. W.G., deliver 20,000 CFM at 5.0 in. W.G. into a common duct. Parallel fan arrangements are used for systems where the flow varies widely and a single fan would operate in stall at low flow. A parallel fan array with two 10,000 CFM fans can operate both fans at full capacity (20,000 CFM), one fan at full capacity (10,000 CFM with the other isolated by a backdraft damper), or both fans at reduced speed via VFD (any flow from 4,000 to 20,000 CFM). The parallel fan arrangement provides turndown from 20,000 CFM to 4,000 CFM — a 5:1 turndown ratio — compared to a single fan with a 2.5:1 turndown ratio (20,000 CFM stall limit at 8,000 CFM). Parallel fan arrays are standard for large laboratory exhaust systems above 30,000 CFM where the CFM varies from 10,000 CFM at night (20 percent occupancy) to 40,000 CFM during peak daytime use. The cost of two fans in parallel is 50 to 70 percent more than a single fan of the same total capacity, but the turndown capability and N+1 redundancy justify the premium for systems where continuous exhaust is critical for life safety.
Motor Sizing: Power, Service Factor, and Drive Losses
Motor sizing for industrial exhaust fan sizing must account for three factors beyond the fan shaft power: belt drive transmission losses, motor service factor, and ambient temperature derating. The fan shaft power (the power required at the fan shaft to move the design CFM against the system pressure) is calculated from the fan selection process — a fan requiring 25 HP at the fan shaft at 10,000 CFM and 9.0 in. W.G. with 78 percent static efficiency uses the formula HP = (CFM × static pressure) / (6,356 × efficiency) = (10,000 × 9.0) / (6,356 × 0.78) = 25.0 HP. The belt drive transmission efficiency for a belt-driven FRP centrifugal fan is 92 to 96 percent — the motor must deliver 25.0 / 0.94 = 26.6 HP at the motor shaft to provide 25 HP at the fan shaft. The motor service factor — typically 1.15 for standard industrial motors — means the motor can operate at 15 percent above its nameplate rating without overheating. A 25 HP motor with 1.15 service factor can deliver 25 × 1.15 = 28.75 HP continuously. In this example, the motor must deliver 26.6 HP at the motor shaft — a 25 HP motor with 1.15 service factor has a capacity of 28.75 HP, which covers the 26.6 HP requirement with 8 percent margin. The motor selection is 25 HP (standard frame size) rather than 30 HP (next frame size, more expensive).
The ambient temperature derating applies when the motor is installed in a high-temperature environment. Standard motors are rated for 40°C ambient — for each 10°C above 40°C, the motor capacity must be derated by 10 percent. A motor installed in a roof-mounted fan enclosure at 55°C ambient must be derated to 1.0 − (55 − 40) / 10 × 0.10 = 85 percent of its nameplate rating. A 25 HP motor in this environment can deliver 25 × 0.85 = 21.25 HP continuous without exceeding its insulation temperature limit. If the fan requires 26.6 HP at the motor shaft, the motor must be sized at 26.6 / 0.85 = 31.3 HP — select the next standard size of 30 HP with 1.15 service factor, which provides 30 × 1.15 × 0.85 = 29.3 HP at 55°C ambient, covering the 26.6 HP requirement. The motor enclosure must match the installation environment — TEFC for indoor clean areas, TENV or TEAO for outdoor or chemical areas where the motor cooling air may contain corrosive gases. For belt-driven FRP fans in chemical exhaust service, the motor is mounted outside the airstream and receives ambient cooling air — a TEFC motor with a NEMA 4X corrosion-resistant coating is the standard specification. The motor frame material must be cast iron or SS 304 (not aluminum) for chemical areas where atmospheric acid concentrations attack aluminum. Per AMCA Standard 99, fan motor selection must include all drive and environmental factors to ensure reliable operation at the design operating point.
Worked Example: Fan Selection for 15,000 CFM Chemical Scrubber
A chemical scrubber system requires 15,000 CFM at 12.0 in. W.G. total system pressure including the scrubber, ductwork, fittings, and stack. Step 1 — calculate the system resistance curve constant k: k = 12.0 / (15,000)² = 5.33 × 10⁻⁸. The system curve is ΔP = 5.33 × 10⁻⁸ × Q². At 12,000 CFM, ΔP = 7.68 in. W.G. At 18,000 CFM, ΔP = 17.3 in. W.G. Step 2 — select a fan model: a 700 mm backward-curved FRP centrifugal fan is selected as the starting point based on manufacturer data showing this diameter handles 12,000 to 20,000 CFM. At 1,100 RPM, the fan curve intersects the system curve at 13,200 CFM and 9.3 in. W.G. — 12 percent below the 15,000 CFM target. Step 3 — increase the fan speed using the fan laws to reach the design point. The required speed for 15,000 CFM is 1,100 × (15,000/13,200) = 1,250 RPM. At 1,250 RPM, the fan pressure at the system curve intersection must be checked: pressure at 1,250 RPM is 9.3 × (1,250/1,100)² = 9.3 × 1.29 = 12.0 in. W.G. — the pressure matches the system requirement at the design CFM. The operating point is 15,000 CFM at 12.0 in. W.G. at 1,250 RPM — the correct fan selection.
Step 4 — verify the tip speed: 700 mm impeller at 1,250 RPM, tip speed = π × 0.700 × 1,250 / 60 = 45.8 m/s. This slightly exceeds the 45 m/s limit for standard backward-curved FRP impellers. The options are: reduce the impeller diameter to 630 mm and increase speed (630 mm at 1,390 RPM, tip speed = π × 0.630 × 1,390 / 60 = 45.8 m/s — same issue), or select a compression-molded radial impeller with 84 m/s limit (no tip speed issue), or specify the same fan with a reinforced impeller (manufacturer-confirmed to 50 m/s). The recommended selection: 700 mm backward-curved FRP fan with manufacturer-confirmed reinforced impeller for 46 m/s max tip speed, at 1,250 RPM. Step 5 — motor sizing: shaft power at 15,000 CFM and 12.0 in. W.G. with 80 percent static efficiency: HP = (15,000 × 12.0) / (6,356 × 0.80) = 35.4 HP at the fan shaft. Belt drive at 94 percent efficiency: motor power required = 35.4 / 0.94 = 37.7 HP. Ambient temperature correction: motor installed outdoors at 45°C ambient — derating factor = 1.0 − (45 − 40)/10 × 0.10 = 0.95. Required motor nameplate rating = 37.7 / 0.95 = 39.7 HP — select the next standard size: 40 HP motor with 1.15 service factor. At 1.15 service factor, the motor can deliver 40 × 1.15 × 0.95 = 43.7 HP at 45°C ambient — covering the 37.7 HP requirement with 16 percent margin. Step 6 — VFD selection: for VAV operation with future turndown to 8,000 CFM, verify the fan’s stable operating range. The fan free-air CFM at 1,250 RPM is approximately 24,000 CFM. The stall limit is 40 percent of free-air = 9,600 CFM. The VAV minimum of 8,000 CFM is below the stall limit — for operation at 8,000 CFM, either accept stall operation (not recommended for continuous duty) or install a parallel fan that can be shut off at low load. Recommended: install two fans in parallel, each sized for 10,000 CFM at 8.0 in. W.G. (the system curve at 10,000 CFM is 5.33 × 10⁻⁸ × 10,000² = 5.33 in. W.G., but with two fans operating the combined system curve shifts — each fan sees 7,500 CFM at 9.0 in. W.G.). The parallel fan configuration provides turndown from 15,000 CFM to 5,000 CFM without stall.
Industrial Exhaust Fan Sizing FAQ
What is the system resistance curve and why is it needed for fan sizing?
The system resistance curve is ΔP = k × Q², representing the static pressure the fan must overcome at any CFM. It is needed because a fan selected without the system curve may deliver 15 to 30 percent less CFM than required if the actual system pressure differs from the assumed value.
How do the fan laws affect speed changes?
CFM varies with RPM (Q₂ = Q₁ × RPM₂/RPM₁), pressure varies with RPM², and power varies with RPM³. A 10 percent speed increase provides 10 percent more CFM, 21 percent more pressure, and requires 33 percent more motor power.
What is the minimum stable operating speed for a VAV fan?
40 percent of the fan’s free-air CFM. Below this point, the fan enters stall with fluctuating flow and vibration. For VAV systems requiring turndown below the stall limit, use a parallel fan arrangement.
How is motor power calculated for a belt-driven fan?
Fan shaft power HP = (CFM × SP) / (6,356 × efficiency). Add 4 to 8 percent for belt drive transmission loss. Apply ambient temperature derating (10 percent per 10°C above 40°C). Select the next standard motor size.
When are series or parallel fans used?
Series fans (discharge to inlet) add pressure at the same CFM — used for systems above 15.0 in. W.G. such as deep bed scrubbers. Parallel fans (multiple fans into common header) add CFM at the same pressure — used for systems requiring turndown below the single-fan stall limit.
What is the maximum tip speed for FRP fan impellers?
45 m/s for backward-curved FRP impellers, 84 m/s for compression-molded radial impellers. Exceeding the tip speed limit causes the impeller to delaminate at the blade root within 6 to 12 months. Verify tip speed for every fan selection.
Industrial exhaust fan sizing requires matching the fan performance curve to the system resistance curve at the design operating point. The system curve ΔP = k × Q² must be calculated from the duct layout and equipment pressure drop before the fan is selected. Proper industrial exhaust fan sizing accounts for fan laws, VAV turndown limits, belt drive losses, ambient temperature derating, and tip speed limits. Fan speed changes follow the law that power varies with the cube of speed — a 10 percent speed increase requires 33 percent more motor power. Per OSHA 29 CFR 1910.94, exhaust fans serving hazardous processes must maintain design airflow at all hoods — correct fan sizing is the foundation of this compliance requirement. For the complete fan selection methodology covering all fan types, see our FRP blower selection guide. Contact XICHENG EP LTD for industrial exhaust fan sizing assistance for your chemical exhaust system.
For overall system duct design, see our industrial exhaust system design guide. For fan type selection, see our FRP blower selection guide.