ASHRAE 62.1: uncommon calculations, approaches

ASHRAE Standard 62.1 is best known for its regulation of the amount of ventilation air delivered to each space by HVAC systems through its ventilation-rate procedure approach to system design. ASHRAE 62.1 offers many calculations and performance approaches.

By Peter Alspach, PE, Arup, Seattle August 14, 2017

Learning objectives: 

  • Know the basics of ASHRAE Standard 62.1: Ventilation for Acceptable Indoor Air Quality.
  • Summarize the less common calculations within this standard.
  • Review examples to enhance the understanding of these calculations.

Most HVAC engineers are familiar with ASHRAE Standard 62.1: Ventilation for Acceptable Indoor Air Quality. Likewise, the fundamental calculations required by the standard are also generally familiar and have been adopted into many building codes within the United States, either explicitly or as an alternative means of compliance.

However, there are some calculations that are less commonly applied and/or codified that are worth exploring. Some of these calculations are required when pursuing green building certifications, such as U.S. Green Building Council LEED compliance, as a part of the overall ASHRAE 62.1 compliance documentation (though some are often ignored by many practitioners and reviewers). Other calculations can help the designer improve the ventilation and indoor-air quality (IAQ) performance of their designs or offer performance-based alternatives to prescriptive requirements for consideration.

While this article offers an overview of some of these calculations, users of the standard are encouraged to purchase the ASHRAE 62.1-2016 User’s Manual and accompanying calculation spreadsheets from ASHRAE for a thorough and complete understanding of all aspects of the standard.

This article will summarize and provide examples for the following less common calculations included within ASHRAE Standard 62.1-2016:

  • Section 5.9.1: Relative Humidity (RH).
  • Section Short-Term Conditions.
  • Appendix A, A1.2.2: Secondary Recirculation Air Systems.
  • Appendix B, B2: Determining L Distance (for exhaust-air outlet separations).

Section 5.9.1: Relative Humidity

Section 5 is an important portion of the standard that many engineers and architects often fail to fully evaluate for compliance. This section of the standard governs many aspects of the HVAC and building (read: architectural) systems that impact IAQ. Aspects governed by this section include duct and air distribution system design, air handling system design, outside-air intake proximity to exhausts and other potential contaminant sources, controls, filtration, rain entrainment and façade moisture management, and operations and maintenance, to name just some of the many important aspects covered by the standard and required for compliance.

One important, and often overlooked, aspect of good ventilation and HVAC system design is the control of indoor RH. While the standard does not attempt to mandate design checks for all system operational situations, Section 5.9.1 requires the designer to address RH management of the HVAC system under a typical dehumidification design condition.

The dehumidification design condition within the standard is to maintain a maximum indoor RH of 65% under the following:

  • Outdoor conditions at the dehumidification design conditions (user-stipulated at the 2%, 1%, or 0.4% conditions); dewpoint and mean coincident dry-bulb temperatures.
  • Indoor sensible and latent loads at design conditions.
  • No solar loads (but shall include envelope conduction and infiltration loads).

It is important to consider this type of scenario, as well as other part-load scenarios with potentially high outdoor or indoor latent loads, when designing HVAC systems to properly manage indoor RH, especially in warm and humid climate zones that predominate in the eastern and southern United States.

Relative humidity: Example 1

  • Location: Baltimore.
  • 1% design conditions: 74.1°F dew point; 81.1°F mean coincident dry-bulb temperatures.
  • Indoor design target: 74°F dry-bulb; 65% RH.
  • Space: A 1,000-sq-ft conference room with an exterior window.
  • System type: Variable air volume (VAV).
  • Internal loads: 17,000 Btu/h sensible; 10,000 Btu/h latent.
  • Envelope load (conduction): 500 Btu/h sensible.
  • Infiltration load: 300 Btu/h sensible; 1,200 Btu/h latent.
  • Total space load: 17,800 Btu/h sensible; 11,200 Btu/hour latent; sensible heat ratio (SHR) = 0.61.

For the analysis, we will assume that the system is a VAV system with a typical supply air condition of 55°F dry-bulb, 54°F wet-bulb under the aforementioned design condition, with the supply airflow rate determined by the space-sensible cooling load. This results in a delivered airflow rate of 825 cfm.

The resulting psychrometric analysis shows that the indoor space conditions will be 74°F dry-bulb, 63% RH (see Figure 1). In this condition, the HVAC system design meets the ASHRAE design requirements as outlined in Section 5.9.1 of the standard. However, the designer should note that the conditions are very close to the maximum allowed—any adjustments in supply air temperature above the design condition should be avoided to maintain acceptable indoor RH conditions.

Relative humidity: Example 2

  • Location: Los Angeles.
  • 1% design conditions: 66.2°F dew point; 73°F mean coincident dry-bulb temperatures.
  • Indoor design target: 74°F dry-bulb; 65% RH.
  • Space: A 500-sqft cocktail lounge with two exterior windows.
  • System type: Dedicated outside-air system (DOAS) with chilled-water fan coil unit (FCU).
  • Internal loads: 17,250 Btu/h sensible; 13,750 Btu/h latent.
  • Envelope load (conduction): -200 Btu/h sensible.
  • Infiltration load: -50 Btu/h sensible; 525 Btu/h latent.
  • Total space load: 17,000 Btu/h sensible; 14,275 Btu/h latent; SHR = 0.54.
  • Peak-cooling space-sensible load: 25,000 Btu/h, including solar load (for FCU system sizing).

For the analysis, we will assume that the DOAS provides ventilation air at a constant rate of 450 cfm at a supply air condition of 55°F dry-bulb, 54°F wet-bulb. We will assume that the fan coil runs at a constant volume of 750 cfm and modulates its cooling coil to meet the space’s dry-bulb setpoint. The ventilation air provides 9,720 Btu/h of sensible cooling, with the fan coil picking up the remaining 7,280 Btu/h of sensible cooling at a supply air condition of 65°F dry-bulb, 62.7°F wet-bulb (sensible-only cooling is assumed due to the part-load operation).

With an SHR of 0.54, the resultant predicted indoor RH is 72%. This clearly fails the required design criteria (see Figure 2). In reality, the indoor RH will be higher than 72%, as the initial analysis assumed a drier starting point for the FCU cooling coil, which will result in a higher-enthalpy mixed-air condition and resultant higher indoor RH.

In this condition, the designer must modify the design or controls to meet the required maximum RH of 65%. The conventional, if inefficient, approach would be to subcool and reheat at the FCU to meet the RH target. The more energy-efficient approach would be to use a variable-speed (not multispeed/staged fan control) fan coil and maintain a constant supply air temperature off the fan coil of approximately 55°F. This would provide additional latent cooling without wasteful reheat energy consumption.

Short-term conditions

Section, Short-Term Conditions, addresses design situations that occur less frequently, but are helpful to give the designer flexibility in designing a right-sized ventilation system. Some space types, projects, or program areas have highly intermittent occupancies, which may not warrant the full ventilation flow rate on a constant basis that would otherwise be determined by the standard if occupancy were relatively constant.

The standard approaches averaging by looking at an averaging time that is equivalent to three time constants (approximately 95%) of the space without any averaging reductions. The formula is given as Equation (IP):


v = the volume of the ventilation zone, in ft3.

Vbz = the breathing zone outside airflow rate, in cubic feet per minute, calculated per equation at the design, steady-state, population.

T = averaging time (minutes).

Once the time period is calculated, the variable occupancy can be averaged over that period. In addition to allowing the ventilation rate to be based on an average occupancy over that time period, the standard also allows the designer to vary the supply of ventilation air to the space over time, provided the time-averaged ventilation rate is met.

For example, the ventilation supply could be varied by cycling a VAV box damper between closed and open positions or cycling an outside-air damper for a system. Averaging the ventilation supply can be helpful in meeting low-ventilation requirements in equipment with insufficient turndown.

Variable-occupancy example

Setting: A 10,000-sq-ft concert hall lobby with 43.5-ft ceilings. The lobby is expected to hold a maximum of 1,500 guests ahead of the concert, with guests expected to arrive over the course of 1 hour, all mingling for 15 minutes before the main hall doors are opened and then entering the hall over the course of another 15 minutes. The lobby remains empty for 2 hours during the concert. The guests then exit over 15 minutes, and the venue is then empty for the remainder of the night (staff ignored for simplicity). This occupancy profile is shown in Figure 3.

The first step is to determine the time period over which the averaging must occur. The space volume is 435,000 ft3. The breathing-zone ventilation rate, assuming a zone-ventilation effectiveness of 1.0, is calculated based on the peak occupancy of 1,500 people and an area of 10,000 sq ft = 1,500 x 2.5 + 10,000 x 0.06 = 4,350 cfm. Therefore:

The average occupancy over the course of the 5-hour time-averaging period is, therefore, 337.5 occupants. This revised average occupancy can be used to calculate the design ventilation rate.

This approach reduces the design breathing-zone ventilation rate from 4,350 cfm to 1,444 cfm. This can result in significant operational and capital cost savings if applied on known variable and intermittently occupied spaces, particularly those with high air volumes.

However, due to the large reductions, these assumptions should be thoroughly reviewed with any building owner to ensure they meet the current and future needs and flexibility of the space. Alternatively, the system could maintain the capacity and capability of the full ventilation rate and simply use controls to reduce the ventilation rate, leaving the flexibility for the owner to later change the basis for the design ventilation approach.

Secondary recirculation systems

For systems where the supply air consists partially of zone-level return air, the zone-level ventilation effectiveness needs to take the return system into account, as found in Normative Appendix A: A1.2.2 Secondary Recirculation Systems. The most common example of this situation is a fan-powered VAV unit. The zone-level ventilation effectiveness is determined by the equation:


  • Evz = zone-ventilation efficiency.
  • Fa = supply air fraction.
  • Xs = average outdoor-air fraction.
  • Fb = mixed-air fraction.
  • Zpz = primary outdoor-air fraction.
  • Ep = primary air fraction.
  • Fc = outdoor-air fraction.
  • Fa = supply air fraction.

One of the key variables for the designer to consider is Er—the one secondary recirculation fraction. The ASHRAE 62.1 User Manual provides guidance on selection of the zone’s secondary recirculation fraction. For a terminal unit with a ducted return to the zone served, Er = 0. For a design where the system-level return is used, such as a dual-fan or dual-duct, Er = 1. The purpose of this calculation is to appropriately credit zones for transfer air from other overventilated zones within the system.

Fan-powered VAV example

For this example, we shall consider a series fan-powered VAV terminal supplying 2,000 cfm of air to an open-office zone of 1,600 sq ft in heating mode using plenum return. At design conditions, the leaving air temperature of the terminal unit is 90°F with a design space temperature of 70°F. The heating primary airflow is set at 200 cfm by the designer. The air handling unit (AHU) at design heating condition has an outside-air fraction of 50%. Note that equations should be referenced from the standard-not all are repeated here for brevity.

Er is assumed to be 0.5 based on guidance from the User Manual, given the location of the terminal unit’s return.

Vbz = 136 cfm using default values from Table

Xs = 0.50.

Ep = 200/2,000 = 0.1.

Ez = 0.8 per Table

Voz = 136/0.8 = 170 cfm.

Zpz = 170/200 = 0.85.

Fa = 0.1 + (1 – 0.1) x 0.5 = 0.55.

Fb = Ep = 0.1.

Fc = 1 – (1 – 0.8) x (1 – 0.5) x (1 – 0.1) = 0.91.

Evz = (0.55 + 0.50 x 0.1 – 0.85 x 0.1 x 0.91)/0.55 = 0.95.

This example is simplified for illustration of the calculation. In reality, the determination of the AHU’s outside-air fraction will be an iterative process amongst all the zones-some single-duct VAV interior zones and some fan-powered perimeter zones. The ASHRAE 62MZ spreadsheet calculation tool or another equivalent or approved method (alternative methods are typically approved by local authorities having jurisdiction, or AHJ), such as some load-calculation tools, is recommended for properly calculating the complexities of multiple-zone systems.

Determining distance L

For most designers, Table 5.5.1 is generally suitable for determining the separation distance required for an outside-air intake from pollutant sources. However, Appendix B provides designers a performance-based alternative for determining the required separation distance. ASHRAE 62.1 uses the “stretched string” measurement approach for determining compliance with the separation distance. However, there is one caveat for noxious or dangerous air (Class 4 air)—if the vertical separation between the exhaust location and the outside-air intake is less than 65 ft, the L separation distance shall be measured only in the horizontal plane (i.e., there can be no credit for any vertical separation).

Three examples are shown to illustrate the performance-based analytical approach of Appendix B, with comparisons to Table 5.5.1. All approaches use equation B2-1 for the calculation, taking into consideration the air class of the exhaust, the discharge velocity, and the flow rate.


  • Q = volumetric flow rate of the exhaust, in cubic feet per minute.
  • DF = dilution factor, from Table B2-2.
  • U = exhaust-air discharge velocity, in feet per minute.

Plumbing vent example

Three plumbing vents are clustered together on a roof. What is the minimum required separation for the nearest AHU outside-air intake?

For this example, we consider the three vents to be acting as a single vent. The standard guides the designer to use an exhaust flow rate of 150 cfm for a plumbing vent. Therefore, the total flow rate, Q, is 450 cfm. Plumbing vents are considered Class 3 air, so the dilution factor is 15. A discharge velocity of 0 is assumed per Table B2-3 for plumbing vents.

From Table 5.5.1, the designer can see that, prescriptively, for a vent less than 3 ft above the level of the outside-air intake, the minimum separation distance is 10 ft. Therefore, Appendix B provides a potential advantage for siting of the outside-air intake closer to the plumbing vents, if required for spatial coordination. However, this will need to be reviewed by the local AHJ for acceptability; AHJ approval for deviations from historic and prescriptive requirements may prove difficult.

Boiler example

Four 6,000-MBH boilers are vented through a single flue with a powered exhaust located on the roof. The exhaust fan has a discharge velocity of 1,500 fpm. What is the minimum required separation for the nearest operable window?

Note that an operable window, when used for natural ventilation, must meet the same separation requirements as an AHU outside-air intake.

A boiler flue is considered Class 4 air as it conveys dangerous particles of combustion, therefore the dilution factor is 50 from Table B2-2. The default assumption from the standard of 250 cfm/1 million Btu/h of boiler capacity is used for the calculation (note that this should always be verified with the final equipment selection); therefore Q = 6,000 cfm. Because the exhaust is hot and discharged upward, 500 fpm is added to the discharge velocity, thus U = 2,000.

In this case, Table 5.5.1 requires a separation of 15 ft; therefore, the design benefits of this analytical approach are marginal for the designer. However, if fewer boilers were combined with a similar discharge velocity, it would likely be advantageous if closer outside-air proximity was required. Also note that for gas-burning appliances, NFPA 54: National Fuel Gas Code must also be followed regardless of allowances within ASHRAE 62.1.

Grease-exhaust example

A neighboring building has a 10,000-cfm grease-exhaust fan discharging toward your building’s roof. The exhaust velocity is 1,500 fpm. Your building desires to locate a shallow balcony with a door leading into a lounge at a level 70 ft above the discharge, and the door is 20 ft horizontally away from the grease-fan discharge. Is this situation compliant with the standard?

Grease exhaust is considered Class 4 air per Table 5.16.1. Because the discharge is directed toward the building opening, U is given a negative value per Table B2-3.

Section B1.3 of the standard allows a separation of L/2 to be used for openings that are not used for mechanical or natural ventilation. However, that allowance does not apply to Class 4 exhaust-air streams-the separation must equal L. Because the balcony is more than 65 ft above the fan discharge, the “stretched string” measurement approach can be used. In this case, the distance between the discharge and the door is:

Using Appendix B, the calculation indicates that the door location is not acceptable, as the calculated L distance is greater than the actual “stretched string” distance.

However, the designer may also choose to use Table 5.5.1 instead-here, the separation is prescriptively listed as 30 ft, indicating the configuration meets the standard. While the designer is entitled to use Table 5.5.1 if lower, the designer should further evaluate the situation and try to remedy the condition. The Appendix B calculation clearly indicates that a potential odor/entrainment problem will exist and mitigation measures should be considered.

For example, if the grease-exhaust discharge can be redirected to be at least 45 deg away from the opening, U then equals 0, resulting in L = 63.6 ft, an acceptable condition per the standard.

Like many ASHRAE standards, Standard 62.1 has aspects that many designers either are not familiar with or do not typically use. This article sheds a bit of light on some of those aspects that will bring the designer a greater understanding of maintaining proper IAQ as well as increased design flexibility, which may be leveraged to save first cost, energy, operating costs, and provide greater design flexibility in architectural coordination.

As with most performance-based design approaches, the designer needs to carefully consider the pros and cons of deviating from accepted engineering norms or designing too tightly around singular scenarios that may limit future flexibility.

Peter Alspach is a principal and mechanical engineer at Arup. His expertise is in HVAC systems design, building physics analysis, and façade engineering. Alspach is a member of the Consulting-Specifying Engineer editorial advisory board and was a 2008 40 Under 40 award winner.