Ashrae Cooling And Heating Load Calculation Manual Grp 15801

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The cooling load temperature difference / cooling load factor (CLTD/CLF) method has been a popular method for performing cooling load calculations since the publication of ASHRAE GRP-158, the Cooling and Heating Load Calculation Manual (ASHRAE 1979). Originally developed as a hand calculation technique, it was constrained to use.

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  • The Trane Company believes that it is incumbent on manufacturers to serve the industry by regularly disseminating information gathered through laboratory research, testing programs, and field experience.
    The Trane Air Conditioning Clinic series is one means of knowledge sharing. It is intended to acquaint a nontechnical audience with various fundamental aspects of heating, ventilating, and air conditioning. We have taken special care to make the clinic as uncommercial and straightforward as possible. Illustrations of Trane products only appear in cases where they help convey the message contained in the accompanying text.
    This particular clinic introduces the reader to cooling and heating load estimation. It is intended to introduce the concepts of estimating building cooling and heating loads and is limited to introducing the components that make up the load on a building, the variables that affect each of these components, and simple methods used to estimate these load components. It is not intended to teach all the details or latest computerized techniques of how to calculate these loads.
    If you are interested in learning more about the specific techniques used for cooling and heating load estimating, this booklet includes several references in the back.
  • The three basic principles of heat transfer discussed in this clinic are:
    1) Heat energy cannot be destroyed; it can only be transferred to another substance.
    To produce cooling, heat must be removed from a substance by transferring the heat to another substance. This is commonly referred to as the principle of 'conservation of energy.' Ice cubes are typically placed in a beverage to cool it before being served. As heat is transferred from the beverage to the ice, the temperature of the beverage is lowered. The heat removed from the beverage is not destroyed, but instead is absorbed by the ice, melting the ice from a solid to a liquid.
    2) Heat energy naturally flows from a higher-temperature substance to a lower-temperature substance, in other words, from hot to cold.
    Heat cannot naturally flow from a cold substance to a hot substance. Consider the example of the beverage and the ice cubes. Because the temperature of the beverage is higher than the temperature of the ice cubes, heat will always flow from the beverage to the ice cubes.
    3) Heat energy is transferred from one substance to another by one of three basic processes: conduction, convection, or radiation.
  • The device shown is a baseboard convector that is commonly used for heating a space. It can be used to demonstrate all three processes of transferring heat.
    Hot water flows through a tube inside the convector, warming the inside surface of the tube. Heat is transferred, by conduction, through the tube wall to the slightly cooler fins that are attached to the outside surface of the tube. Conduction is the process of transferring heat through a solid.
    The heat is then transferred to the cool air that comes into contact with the fins. As the air is warmed and becomes less dense, it rises, carrying the heat away from the fins and out of the convector. This air movement is known as a convection current. Convection is the process of transferring heat as the result of the movement of a fluid. Convection often occurs as the result of the natural movement of air caused by temperature (density) differences.
    Additionally, heat is radiated from the warm cabinet of the convector and warms cooler objects within the space. Radiation is the process of transferring heat by means of electromagnetic waves, emitted due to the temperature difference between two objects. An interesting thing about radiated heat is that it does not heat the air between the source and the object it contacts; it only heats the object itself.
  • In the I–P system of units, the unit for measuring the quantity of heat is the British Thermal Unit (Btu). The Btu is defined as the quantity of heat energy required to raise the temperature of 1 lb of water 1°F.
    Similarly, in the Systeme International (SI) system, heat quantity can be expressed using the unit kiloJoule (kJ). A kcal is defined as the amount of heat energy required to raise the temperature of 1 kg of water 1°C. One kcal is equal to 4.19 kJ.
    In heating and cooling applications, however, emphasis is placed on the rate of heat transfer, that is, the quantity of heat that flows from one substance to another within a given period of time. This rate of heat flow is commonly expressed in terms of Btu/hr—the quantity of heat, in Btu, that flows from one substance to another during a period of 1 hour.
    Similarly, in the SI metric system of units, the rate of heat flow is expressed in terms of kilowatts (kW). One kW is equivalent to 1 kJ/sec. One kilowatts describes the quantity of heat, in kJ, that flows from one substance to another during a period of 1 second. Finally, the rate of heat flow may often be expressed in terms of watts (W). One kW is equivalent to 1000 W.
  • The process of comfort heating and air conditioning is simply a transfer of energy from one substance to another. This energy can be classified as either sensible or latent heat energy.
    Sensible heat is heat energy that, when added to or removed from a substance, results in a measurable change in dry-bulb temperature.
    Changes in the latent heat content of a substance are associated with the addition or removal of moisture. Latent heat can also be defined as the “hidden” heat energy that is absorbed or released when the phase of a substance is changed. For example, when water is converted to steam, or when steam is converted to water.
  • Heating and air conditioning systems use the principles of heat transfer to maintain comfortable indoor conditions for people. Human or Thermal comfort refers to the range of temperature, humidity and air movement conditions that most people feel comfortable most of the time. According to ASHRAE, the comfort temperature lies between 78F ( maximum in summer to 68F (minimum in winter). The relative humidity lies between 30%-60%.
  • The term “comfort” is often used to define a broader set of conditions than just temperature and humidity. Air movement, adequate fresh air, cleanliness of the air, noise levels in the space, adequate lighting, and proper furniture and work surfaces, are just a few of the other variables that contribute to making a space comfortable for its occupants. This clinic, however, will focus only on the aspects of thermal comfort.
    Thermal comfort depends on creating an environment of dry-bulb temperature, humidity, and air motion, that is appropriate for the activity level of the people in the space. This environment allows the body’s rate of heat generation to balance with the body’s rate of heat loss.
  • Research studies have been conducted to show that, with a specific amount of air movement, thermal comfort can be produced with certain combinations of dry-bulb temperature and relative humidity. When plotted on a psychrometric chart, these combinations form a range of conditions for delivering acceptable thermal comfort to 80% of the people in a space. This “comfort zone” and the associated assumptions are defined by ASHRAE Standard 55, Thermal Environmental Conditions for Human Occupancy.
    Determining the desired condition of the space is the first step in estimating the cooling and heating loads for the space. In this clinic, we will choose 78ºF [25.6ºC] dry-bulb temperature and 50% relative humidity (A) as the desired indoor condition during the cooling season.
  • The selection of heating, ventilating, and air conditioning (HVAC) system components and equipment should always be based on an accurate determination of the building heating and cooling loads.
    During this period we will estimate the cooling loads for a single space in a single-story office building. In Period Four we will estimate the heating loads for this same space. As stated in the preface, this clinic is intended to introduce the concepts of estimating building cooling and heating loads and is not intended to cover all of the details.
    The Cooling Load Temperature Difference/Solar Cooling Load/Cooling Load Factor (CLTD/SCL/CLF) load estimation method *, used throughout Period Two, is a simplified hand calculation procedure developed long ago by ASHRAE. Because of its simplicity, it is the most common method used for basic instruction on estimating cooling loads.
    * Reference: 1997 ASHRAE Handbook–Fundamentals, Chapter 28, Table 29
  • The space cooling load is the rate at which heat must be removed from a space in order to maintain the desired conditions in the space, generally a dry-bulb temperature and relative humidity. The cooling load for a space can be made up of many components, including:
    Conduction heat gain from outdoors through the roof, exterior walls, skylights, and windows. (This includes the effects of the sun shining on these exterior surfaces.)
    Solar radiation heat gain through skylights and windows.
    Conduction heat gain from adjoining spaces through the ceiling, interior partition walls, and floor.
    Internal heat gains due to people, lights, appliances, and equipment in the space.
    Heat gain due to hot, humid air infiltrating into the space from outdoors through doors, windows, and small cracks in the building envelope.
    In addition, the cooling coil in the building HVAC system has to handle other components of the total building cooling load, including:
    Heat gain due to outdoor air deliberately brought into the building for ventilation purposes.
    Heat generated by the fans in the system and possibly other heat gains in the system.
    Throughout this period, we will assume that the space has no plenum (the space between the ceiling and roof). Therefore, all of the heat gain due to the roof and lighting affects the space directly.
  • These load components contribute sensible and/or latent heat to the space. Conduction through the roof, exterior walls, windows, skylights, ceiling, interior walls, and floor, as well as the solar radiation through the windows and skylights, all contribute only sensible heat to the space.
    The people inside the space contribute both sensible and latent heat. Lighting contributes only sensible heat to the space, while equipment in the space may contribute only sensible heat (as is the case for a computer) or both sensible and latent heat (as is the case for a coffee maker). Infiltration generally contributes both sensible and latent heat to the space.
    The cooling coil has to handle the additional components of ventilation and system heat gains. Ventilation contributes both sensible and latent heat to the coil load. Other heat gains that occur in the HVAC system (from the fan, for example) generally contribute only sensible heat.
  • One of the more difficult aspects of estimating the maximum cooling load for a space is determining the time at which this maximum load will occur. This is because the individual components that make up the space cooling load often peak at different times of the day, or even different months of the year.
    For example, the heat gain through the roof will be highest in the late afternoon, when it is warm outside and the sun has been shining on it all day. Conversely, the heat gain due to the sun shining through an east-facing window will be highest in the early morning when the sun is rising in the east and shining directly into the window.
    Determining the time that the maximum total space cooling load occurs will be discussed later in this clinic.
  • Room 101 is the space that we will use as an example throughout this clinic. The windows face west and the solar heat gain through these windows will peak in the late afternoon when the sun is setting and shining directly into the windows. Because of this, we will assume that the maximum cooling load for our example space occurs at 4 p.m.
    For this example, the following criteria will be used as a basis for estimating the space cooling and heating loads.
    Open-plan office space located in a single-story office building in St. Louis, Missouri.
    Floor area = 45 ft x 60 ft [13.7 m x 18.3 m].
    Floor-to-ceiling height = 12 ft [3.7 m] (no plenum between the space and roof).
    Desired indoor conditions = 78ºF [25.6ºC] dry-bulb temperature, 50% relative humidity during cooling season; 72ºF [22.2ºC] dry-bulb temperature during heating season.
    West-facing wall, 12 ft high x 45 ft long [3.7 m x 13.7 m], constructed of 8 in. [203.2 mm] lightweight concrete block with aluminum siding on the outside, 3.5 in. [88.9 mm] of insulation, and ½ in. [12.7 mm] gypsum board on the inside.
    Eight clear, double-pane (¼ in. [6.4 mm]) windows mounted in aluminum frames. Each window is 4 ft wide x 5ft high [1.2 m x 1.5 m].
    Flat, 45 ft x 60 ft [13.7 m x 18.3 m] roof constructed of 4 in. [100 mm] concrete with 3.5 in. [90 mm] insulation and steel decking.
    Space is occupied from 8:00 a.m. until 5:00 p.m. by 18 people doing moderately active work.
    Fluorescent lighting in space = 2 W/ft2 [21.5 W/m2].
    Computers and office equipment in space = 0.5 W/ft2 [5.4 W/m2], plus one coffee maker.
    In order to simplify this example, we will assume that, with the exception of the west-facing exterior wall, room 101 is surrounded by spaces that are air conditioned to the same temperature as this space.
  • Outdoor Design Conditions
    In Period One, we discussed the indoor conditions required for thermal comfort. The next step toward estimating the cooling load of a space is to determine the highest, frequently-occurring outdoor air temperature. In the summer, for example, when the temperature outside is high, heat transfers from outdoors to indoors, thus contributing to the heat gain of the space.
    Obviously, HVAC systems would be greatly oversized if cooling load calculations were based on the most extreme outdoor temperature ever recorded for the location. Instead, outdoor design temperatures are based on their frequency of occurrence. Design outdoor conditions for many locations can be found in the ASHRAE Handbook—Fundamentals.
    Figure 16, outdoor design conditions for St. Louis, Missouri, includes three columns of dry-bulb temperatures and corresponding wet-bulb temperatures. The first column heading, 0.4%, means that the dry-bulb temperature in St. Louis exceeds 95ºF [35ºC] for only 0.4% of all of the hours in an average year (or 35 hours). Also, 76ºF [25ºC] is the wet-bulb temperature that occurs most frequently when the dry-bulb temperature is 95ºF [35ºC]. The second column heading, 1%, means that the temperature exceeds 93ºF [34ºC] for only 1% of all of the hours in an average year (or 87.6 hours). When the dry-bulb temperature is 93ºF [34ºC], the wet-bulb temperature that occurs most frequently is 75ºF [24ºC]. For our example, we will use the more severe 95ºF [35ºC] dry bulb and 76ºF [25ºC] wet bulb for the outdoor design conditions.
    The tables * published by ASHRAE include more weather data that can be useful for sizing certain HVAC system components, but that discussion is outside the scope of this clinic.
    * Reference: 1997 ASHRAE Handbook–Fundamentals, Chapter 26
  • Conduction through Surfaces
    Conduction is the process of transferring heat through a solid, such as a wall, roof, floor, ceiling, window, or skylight. Heat naturally flows by conduction from a higher temperature to a lower temperature. Generally, when estimating the maximum cooling load for a space, the temperature of the air outdoors is higher than the temperature of the air indoors.
    We will focus on the most common conduction heat gains to a space: through the roof, external walls, and windows.
  • Although often not applicable, a simplifying assumption when estimating the conduction heat gain through an exterior surface is to assume that the surface is completely shaded at all times. With this assumption, the amount of heat transferred through the surface is a direct result of the temperature difference between the space and outdoors. This assumption, however, does not include the additional heat transfer that occurs because of the sun shining on the surface. This will be discussed next.
    The amount of heat transferred through a shaded exterior surface depends on the area of the surface, the overall heat transfer coefficient of the surface, and the dry-bulb temperature difference from one side of the surface to the other. The equation used to predict the heat gain by conduction is:
    Q = U x A x T
    where,
    Q = heat gain by conduction, Btu/hr [W]
    U = overall heat-transfer coefficient of the surface, Btu/hr•ft2•F [W/m2•K]
    A = area of the surface, ft2 [m2]
    T = dry-bulb temperature difference across the surface, ºF [C]
    In the case of a shaded exterior surface, this temperature difference is the design outdoor dry-bulb temperature (To) minus the desired indoor dry-bulb temperature (Ti).
  • The overall heat transfer coefficient is also called the U-factor. The U-factor describes the rate at which heat will be transferred through the structure.
    Walls and roofs are typically made up of layers of several materials. The U-factor for a specific wall or roof is calculated by summing the thermal resistances (R-values) of each of these layers and then taking the inverse. The ASHRAE Handbook—Fundamentals tabulates * the thermal resistance of many common materials used in constructing walls, roofs, ceilings, and floors.
    The wall in our example space is comprised of:
    aluminum siding (R = 0.61 ft2•hr•ºF/Btu [0.11 m2•ºK/W])
    8 in. [200 mm] lightweight concrete block (R = 2.0 [0.35])
    3.5 in. [90 mm] of fiberglass insulation (R = 13.0 [2.29])
    ½ in. [12.7 mm] gypsum board (R = 0.45 [0.08])
    Additionally, there is a film of air on the outside surface of the wall (R = 0.25 [0.044], assuming air moving at 7.5 mph [12 km/hr] during the summer) and another film of air on the inside surface of the wall (R = 0.68 [0.12], assuming still air).
    * Reference: 1997 ASHRAE Handbook–Fundamentals, Chapter 24, Table 4
  • The U-factor of this wall is calculated by summing the thermal resistances of each of these layers and then taking the inverse.
    The U-factor of the roof in our example is calculated in a similar manner.
  • If the west-facing wall of our example space was completely shaded at all times, the conduction heat gain due to the wall would be 388 Btu/hr [133 W].
    Conduction heat gain through the west-facing wall (assume shaded at all times):
    U-factor = 0.06 Btu/hr•ft2•F [0.33 W/m2•K]
    Total area of wall + windows = 12 ft x 45 ft = 540 ft2 [3.7 m x 13.7 m = 50.7 m2]
    Area of windows = 8 windows x (4 ft x 5 ft) = 160 ft2 [8 x (1.2 m x 1.5 m) = 14.4 m2]
    Net area of wall = 540 – 160 = 380 ft2 [50.7 – 14.4 = 36.3 m2]
    T = outdoor temperature (95ºF [35ºC]) – indoor temperature (78ºF [25.6C])
    Q = U x A x T
    Q = 0.06 x 380 x (95 – 78) = 388 Btu/hr
    [Q = 0.33 x 36.3 x (35 – 25.6) = 113 W]
  • Most exterior surfaces of a building, however, are exposed to direct sunlight during some portion of the day. Solar heat energy is generated by the sun and radiated to earth. Radiant heat is similar to light, in that it travels in a straight line and can be reflected from a bright surface. Both light and radiant heat can pass through a transparent surface (such as glass), yet neither can pass directly through an opaque or non-transparent surface (such as a brick wall). When the sun’s rays strike an opaque surface, however, a certain amount of radiant heat energy is transferred to that surface, resulting in an increase in the surface temperature. The amount of heat transferred depends primarily on the color and smoothness of the surface, and the angle at which the sun’s rays strike the surface.
    When the sun’s rays strike the surface at a 90º angle, the maximum amount of radiant heat energy is transferred to that surface. When the same rays strike that same surface at a lesser angle, less radiant heat energy is transferred to the surface. The angle at which the sun’s rays strike a surface depends upon the latitude, the time of day, and the month of the year. Due to the rotation of the earth throughout the day, and the earth orbiting the sun throughout the year, the angle at which the sun’s rays strike a surface of a building is constantly changing. This varies the intensity of the solar radiation on an exterior surface of a building, resulting in a varying amount of solar heat transferred to the surface throughout the day and throughout the year.
    As mentioned previously, the assumption that the surface is completely shaded does not account for the additional heat gain that occurs when the sun shines on a surface. Solar heat, therefore, must be considered, as it constitutes an important part of the total cooling load of most buildings.
  • The walls and roof that make up a building’s envelope have the capacity to store heat energy. This property delays the heat transfer from outdoors to the space. The time required for heat to be transferred through a structure into the space is called the time lag.
    For example, the heat that is transferred through a sunlit wall into a space is the result of sunlight that fell on the outer surface of the wall earlier in the day. Curve A shows the magnitude of the solar effect on the exterior wall. Curve B shows the resulting heat that is transferred through the wall into the space. This delay in the heat gain to the space is the time lag. The magnitude of this time lag depends on the materials used to construct the particular wall or roof, and on their capacity to store heat.
  • A factor called the cooling load temperature difference (CLTD) is used to account for the added heat transfer due to the sun shining on exterior walls, roofs, and windows, and the capacity of the wall and roof to store heat. The CLTD is substituted for T in the equation to estimate heat transfer by conduction.
    Q = U x A x CLTD
  • This particular table * includes CLTD factors for a west-facing wall similar to the type used in our example building. It should be noted that the data in this table are based on the following assumptions:
    78ºF [25.6ºC] indoor air
    95ºF [35ºC] maximum outdoor air
    Average outdoor daily temperature range of 21ºF [11.7ºC]
    21st day of July
    40º north latitude
    Dark-colored surface
    Tables for various wall and roof types, as well as correction factors for applications that differ from these assumptions, can be found in the 1997 ASHRAE Handbook—Fundamentals and ASHRAE’s Cooling and Heating Load Calculation Principles manual.
    The wall in our example is classified as Wall Type 9. At 4 p.m. (Hour 17 in this table), the CLTD for a west-facing wall of this type is 22ºF [12C]. This means that, even though the actual dry-bulb temperature difference is only 17ºF (95ºF – 78ºF) [9.4ºC (35ºC – 25.6ºC)], the sun shining on the outer surface of this wall increases the “effective temperature difference” to 22ºF [12C].
    Notice that the CLTD increases later in the day, and then begins to decrease in the evening as the stored heat is finally transferred from the wall into the space.
    * This table can be found in the “Notes Page” view of Slide 96.
  • Using CLTD instead of T, we will determine the heat gain, by conduction, through the west-facing wall and the roof. The U-factors are the same as those calculated on Slide 20.
    Conduction heat gain through the west-facing sunlit wall:
    U-factor = 0.06 Btu/hr•ft2•F [0.33 W/m2•K]
    Net area of wall = 380 ft2 [36.3 m2]
    CLTDhour=17 = 22ºF [12C]
    Q = 0.06 x 380 x 22 = 502 Btu/hr
    [Q = 0.33 x 36.3 x 12 = 144 W]
    Table 3 [Table 4] * includes CLTD factors for several types of roofs. The roof in our example building is classified as Roof Type 2. The data in Table 3 [Table 4] are based on assumptions similar to the CLTD table for walls. At Hour 17, the CLTD for a flat roof of this type is 80ºF [44C].
    Conduction heat gain through the roof:
    U-factor = 0.057 Btu/hr•ft2•F [0.323 W/m2•K]
    Area of roof = 45 ft x 60 ft = 2,700 ft2 [13.7 m x 18.3 m = 250.7 m2]
    CLTDhour=17 = 80ºF [44C]
    Q = 0.057 x 2,700 x 80 = 12,312 Btu/hr
    [Q = 0.323 x 250.7 x 44 = 3,563 W]
    * These tables can be found in the “Notes Page” view of Slide 96.
  • Estimating the heat gain by conduction through a window is very similar to walls and roofs.
    The data on this slide * is an excerpt from the 1997 ASHRAE Handbook—Fundamentals and includes U-factors for common window assemblies.
    The windows in our example are double-pane windows with a ¼-inch [6.4 mm] air space between the panes. Assuming that the windows are fixed (not operable), with aluminum frames and a thermal break, the U-factor is 0.63 Btu/hr•ft2•F [3.56 W/m2•K].
    * Reference: 1997 ASHRAE Handbook–Fundamentals, Chapter 29, Table 5
  • Using this U-factor, we will determine the heat gain, by conduction, through the eight west-facing windows. Table 5 [Table 6] is an excerpt from the 1997 ASHRAE Handbook—Fundamentals and includes CLTD factors for glass. It is also based on similar assumptions as the CLTD tables for walls and roofs. At Hour 17, the CLTD for a glass window is 13ºF [7C].
    Conduction heat gain through the west-facing windows:
    U-factor = 0.63 Btu/hr•ft2•F [3.56 W/m2•K]
    Total area of glass = 8 windows x (4 ft x 5 ft) = 160 ft2 [8 x (1.2 m x 1.5 m) = 14.4 m2]
    CLTDhour=17 = 13ºF [7C]
    Q = 0.63 x 160 x 13 = 1,310 Btu/hr
    [Q = 3.56 x 14.4 x 7 = 359 W]
    * These tables can be found in the “Notes Page” view of Slide 96.
  • Solar Radiation through Glass
    Previously, we estimated the heat transferred through glass windows by the process of conduction. A large part of the solar heat energy that shines on a window or skylight is radiated through the glass and transmitted directly into the space. The amount of solar heat radiated through the glass depends primarily on the reflective characteristics of the glass and the angle at which the sun’s rays strike the surface of the glass.
    While glass windows of double- or triple-pane construction do an excellent job of reducing heat transfer by conduction, they do not appreciably reduce the amount of solar directly into a space. To limit the amount of solar radiation entering the space, heat-absorbing glass, reflective glass, or internal or external shading devices can be used.
  • Although the 1997 ASHRAE Handbook—Fundamentals contains new, more advanced methods of estimating solar heat gain through glass, they are beyond the scope of this clinic.
    The equation used to predict the solar heat gain through glass is:
    Q = A x SC x SCL
    where,
    Q = heat gain by solar radiation through glass, Btu/hr [W]
    A = total surface area of the glass, ft2 [m2]
    SC = shading coefficient of the window, dimensionless
    SCL = solar cooling load factor, Btu/hr•ft2 [W/m2]
  • The solar cooling load (SCL) factor is used to estimate the rate at which solar heat energy radiates directly into the space, heats up the surfaces and furnishings, and is later released to the space as a sensible heat gain. Similar to CLTD, the SCL factor is used to account for the capacity of the space to absorb and store heat.
    The value of SCL is based on several variables, including the direction that the window is facing, time of day, month, and latitude. These four variables define the angle at which the sun’s rays strike the surface of the window. The next two variables, the construction of the interior partition walls and the type of floor covering, help define the capacity of the space to store heat. This affects the time lag between the time that the solar radiation warms up the space furnishings and the time that the heat is released into the space. The last variable, whether or not internal shading devices are installed, affects the amount of solar heat energy passing through the glass.
    The 1997 ASHRAE Handbook—Fundamentals contains tables of SCL values for common space types, based on combinations of these variables. Table 7 [Table 8] * is an excerpt from the handbook and includes SCL factors for a space type similar to the one in our example. The space in our example is classified as Space Type A. The data in this table is based on the 21st day of July and 40º north latitude. At Hour 17, the SCL for the west-facing windows in our example space is 192 Btu/hr•ft2 [605 W/m2].
    * These tables can be found in the “Notes Page” view of Slide 96.
  • The shading coefficient (SC) is an expression used to define how much of the radiant solar energy, that strikes the outer surface of the window, is actually transmitted through the window and into the space. The shading coefficient for a particular window is determined by comparing its reflective properties to a standard reference window. The table * on this slide includes shading coefficients for common window systems. When the value for the shading coefficient decreases, more of the sun’s rays are reflected by the outer surface of the glass.
    The windows in our example space are constructed of two panes of ¼-inch [6.4 mm] clear glass with an air space between the panes. The glass is mounted in an aluminum frame and the windows are fixed (not operable). The SC for this type of window is 0.74.
    * Reference: 1997 ASHRAE Handbook–Fundamentals, Chapter 24, Table 4 (SHGC values were converted to SC by dividing SHGC by 0.87, per equation 38)
  • Now, determine the heat gain by solar radiation through the windows on the west-facing wall of our example space.
    Solar radiation heat gain through the windows on the west-facing wall:
    Total area of glass = 8 windows x (4 ft x 5 ft) = 160 ft2 [8 x (1.2 m x 1.5 m) = 14.4 m2]
    SC = 0.74
    SCLhour=17 = 192 Btu/hr•ft2 [605 W/m2]
    Q = 160 x 0.74 x 192 = 22,733 Btu/hr
    [Q = 14.4 x 0.74 x 605 = 6,447 W]
  • Internal Heat Gains
    The next component of the space cooling load is the heat that originates within the space. Typical sources of internal heat gain are people, lights, cooking processes, and other heat-generating equipment, such as motors, appliances, and office equipment.
    While all of these sources contribute sensible heat to the space, people, cooking processes, and some appliances (such as a coffee maker) also contribute latent heat to the space.
  • The necessity for comfort air conditioning stems from the fact that the metabolism of the human body normally generates more heat than it needs. This heat is transferred by convection and radiation to the environment surrounding the body. The average adult, seated and working, generates excess heat at the rate of approximately 450 Btu/hr [132 W]. About 60% of this heat is transferred to the surrounding environment by convection and radiation, and 40% is released by perspiration and respiration.
    As the level of physical activity increases, the body generates more heat in proportion to the energy expended. When engaged in heavy labor, as in a factory for example, the body generates 1,450 Btu/hr [425 W]. At this level of activity, the proportions reverse and about 40% of this heat is transferred by convection and radiation and 60% is released by perspiration and respiration.
  • Similar to the use of the CLTD for conduction heat gain and SCL for solar heat gain, the cooling load factor (CLF) is used to account for the capacity of the space to absorb and store heat. Some of the sensible heat generated by people is absorbed and stored by the walls, floor, ceiling, and furnishings of the space, and released at a later time. Similar to heat transfer by conduction through an external wall, the space can therefore experience a time lag between the time that the sensible heat is originally generated and the time that it actually contributes to the space cooling load. For heat gain from people, the value of CLF depends on 1) the construction of the interior partition walls in the space, 2) the type of floor covering, 3) the total number of hours that the space is occupied, and 4) the number of hours since the people entered the space.
    Figure37, * CLF Factors for People, is an excerpt from the 1997 ASHRAE Handbook—Fundamentals. It shows that one hour after people enter the space, 35% (1 – 0.65) of the sensible heat gain from the people is absorbed by the surfaces and furnishings in the space, and 65% is the actual cooling load in the space. Following the table to the right, however, you see that, as the people are in the space for a longer period of time, the surfaces and furnishings of the space can no longer absorb as much heat, and they release the heat that was absorbed earlier in the day. For example, if the people enter the space at 8 a.m. and remain for a total of 8 hours, at 2 p.m. (6 hours after entering) 91% of the sensible heat gain from the people is seen as a cooling load in the space. Only 9% is absorbed by the surfaces and furnishings of the space.
    If the space is not maintained at a constant temperature during the 24-hour period, however, the CLF is assumed to equal 1.0. Most air-conditioning systems designed for non-residential buildings either shut the system off at night or raise the temperature set point to reduce energy use. Thus, it is uncommon to use a CLF other than 1.0 for the cooling load due to people.
    * Reference: 1997 ASHRAE Handbook–Fundamentals, Chapter 28, Table 37
  • Determine the internal heat gain from people in our example space. Based on the table in Figure 36, people participating in moderately active office work generate 250 Btu/hr [75 W] sensible heat and 200 Btu/hr [55 W] latent heat.
    Internal heat gain from people:
    Number of people = 18
    Sensible heat gain/person = 250 Btu/hr [75 W]
    Latent heat gain/person = 200 Btu/hr [55 W]
    CLF = 1.0 (because the space temperature set point is increased at night)
    QS = 18 people x 250 Btu/hr per person x 1.0 = 4,500 Btu/hr
    [QS = 18 people x 75 W per person x 1.0 = 1,350 W]
    QL = 18 people x 200 Btu/hr per person = 3,600 Btu/hr
    [QL = 18 people x 55 W per person = 990 W]
  • Heat generated by lights in the space is a significant contribution to the cooling load. For example, a 120-watt light fixture generates 410 Btu/hr [120 W] of heat–approximately the same amount of heat gain from an average office worker.
    Additionally, when estimating the heat gain generated by fluorescent lights, approximately 20% is added to the lighting heat gain to account for the additional heat generated by the ballast.
    The equation used to estimate the heat gain from lighting is:
    Q = watts x 3.41 x ballast factor x CLF
    [Q = watts x ballast factor x CLF]
    where,
    Q = sensible heat gain from lighting, Btu/hr [W]
    Watts = total energy input to lights, W
    3.41 = conversion factor from W to Btu/hr (when using I–P units)
    Ballast factor = 1.2 for fluorescent lights, 1.0 for incandescent lights
    CLF = cooling load factor, dimensionless
    Similar to the sensible heat gain from people, a cooling load factor (CLF) can be used to account for the capacity of the space to absorb and store the heat generated by the lights. If the lights are left on 24 hours a day, or if the air-conditioning system is shut off or set back at night, the CLF is assumed to be equal to 1.0.
  • Next, determine the heat gain from lighting in our example space.
    Internal heat gain from lighting:
    Amount of lighting in space = 2 W/ft2 [21.5 W/m2]
    Floor area = 45 ft x 60 ft = 2,700 ft2 [13.7 m x 18.3 m = 250.7 m2]
    Total lighting energy = 2 W/ft2 x 2,700 ft2 = 5,400 W [21.5 W/m2 x 250.7 m2 = 5,400 W ]
    Ballast factor = 1.2 (fluorescent lights)
    CLF = 1.0 (because the space temperature set point is increased at night)
    Q = 5,400 x 3.41 x 1.2 x 1.0 = 22,097 Btu/hr
    [Q = 5,400 x 1.2 x 1.0 = 6,480 W]
  • Infiltration
    In a typical building, air leaks into or out of a space through doors, windows, and small cracks in the building envelope. Air leaking into a space is called infiltration. During the cooling season, when air leaks into a conditioned space from outdoors, it can contribute to both the sensible and latent heat gain in the space because the outdoor air is typically warmer and more humid than the indoor air.
  • Before estimating the heat gain from infiltration, we must first estimate the amount of air that is leaking into the space. There are three methods commonly used to estimate infiltration airflow.
    The air change method is the easiest, but may be the least accurate of these methods. It involves estimating the number of air changes per hour that can be expected in spaces of a certain construction quality. Using this method, the quantity of infiltration air is estimated using the equation:
    infiltration airflow = (volume of space x air change rate)  60
    [infiltration airflow = (volume of space x air change rate)  3,600]
    where,
    Infiltration airflow = quantity of air infiltrating into the space, cfm [m3/s]
    Volume of space = length x width x height of space, ft3 [m3]
    Air change rate = air changes per hour
    60 = conversion from hours to minutes
    3,600 = conversion from hours to seconds
    The crack method is a little more complex and is based upon the average quantity of air known to enter through cracks around windows and doors when the wind velocity is constant. The effective leakage-area method takes wind speed, shielding, and “stack effect” into account, and requires a very detailed calculation.
  • The table below includes estimates for infiltration using the air change method. Assuming that the space in our example is of average construction and kept at a positive pressure relative to the outdoors, we estimate 0.3 air changes/hr of infiltration.
    volume of space = 45 ft x 60 ft x 12 ft = 32,400 ft3
    [13.7 m x 18.3 m x 3.7 m = 927.6 m3]
  • The equation used to estimate the sensible heat gain from infiltration is:
    QS = 1.085 x airflow x T
    [QS = 1,210 x airflow x T]
    where,
    QS = sensible heat gain from infiltration, Btu/hr [W]
    1.085 [1,210] = product of density and specific heat, Btu•min/hr•ft3•ºF [J/m3•ºK]
    Airflow = quantity of air infiltrating the space, cfm [m3/s]
    T = design outdoor dry-bulb temperature minus the desired indoor dry-bulb temperature, ºF [ºC]
    The equation used to estimate the latent heat gain from infiltration is:
    QL = 0.7 x airflow x W
    [QL = 3,010 x airflow x W]
    where,
    QL = latent heat gain from infiltration, Btu/hr [W]
    0.7 [3,010] = latent heat factor, Btu•min•lb/hr•ft3•gr [J•kg/m3•g]
    Airflow = quantity of air infiltrating the space, cfm [m3/s]
    W = design outdoor humidity ratio minus the desired indoor humidity ratio, grains of water/lb of dry air [grams of water/kg of dry air]
    The psychrometric chart can be used to determine the humidity ratio for both outdoor and indoor conditions.
  • Heat gain from infiltration:
    Infiltration airflow = 162 cfm [0.077 m3/s]
    Outdoor conditions: 95ºF [35ºC] dry bulb and 76ºF [25ºC] wet bulb results in Wo = 105 grains of water/lb dry air [15 grams of water/kg dry air]
    Indoor conditions: 78ºF [25.6ºC] dry bulb and 50% relative humidity results in Wi = 70 grains of water/lb dry air [10 grams of water/kg dry air]
    QS = 1.085 x 162 x (95 – 78) = 2,988 Btu/hr
    [QS = 1,210 x 0.077 x (35 – 25.6) = 876 W]
    QL = 0.7 x 162 x (105 – 70) = 3,969 Btu/hr
    [QL = 3,010 x 0.077 x (15 – 10) = 1,159 W]
    Realize that 1.085 and 0.7 [1,210 and 3,010] are not constants, but are derived from properties of air at “standard” conditions (69°F [21°C] dry air at sea level). Air at other conditions and elevations will cause these factors to change.
    Density = 0.075 lb/ft3 [1.2 kg/m3]
    Specific heat = 0.24 Btu/lb•°F [1,004 J/kg•°K]
    Latent heat of water vapor = 1,076 Btu/lb [2,503 kJ/kg]
    0.075 x 0.24 x 60 min/hr = 1.085 [1.2 x 1,004 = 1,210]
  • This completes the estimation of the components of the cooling load for the space. The total space cooling load will be used during a simplified psychrometric analysis in Period Three to determine the quantity and temperature of air required to condition this space.
    As mentioned earlier, the space used in this example has no plenum, so all of the heat gain from the roof and lights affects the space directly. Applications in which this assumption does not apply will be discussed briefly in Period Five.
    In addition to these space cooling loads, there are other loads that affect the cooling coil in the building HVAC system. These include the load of the outdoor air, deliberately brought into the building for ventilation purposes, and heat generated by the fans in the system. These loads are added to the space load to determine the total cooling load for the building. Estimating these additional components is is necessary to properly size the cooling coil for the system.
  • Ventilation
    Outdoor air is often used to dilute or remove contaminants from the indoor air. The intentional introduction of outdoor air into a space, through the use of the building’s HVAC system, is called ventilation. This outdoor air must often be cooled and dehumidified before it can be delivered to the space, creating an additional load on the air-conditioning equipment.
    You should never depend on infiltration to satisfy the ventilation requirement of a space. On days when the outdoor air is not moving (due to wind), the amount of infiltration can drop to zero. Instead, it is common to introduce outdoor air through the HVAC system, not only to meet the ventilation needs, but also to maintain a positive pressure (relative to the outdoors) within the building. This positive pressure reduces, or may even eliminate, the infiltration of unconditioned air from outdoors. To pressurize the building, the amount of outdoor air brought in for ventilation must be greater than the amount of air exhausted through central and local exhaust fans.
  • The sensible and latent loads from ventilation are calculated using the same equations as for infiltration:
    QS = 1.085 x airflow x T
    [QS = 1,210 x airflow x T]
    QL = 0.7 x airflow x W
    [QL = 3,010 x airflow x W]
    Cooling load due to the conditioning of ventilation air:
    Ventilation airflow = 360 cfm [0.18 m3/s]
    Outdoor conditions: To = 95ºF [35ºC], Wo = 105 grains of water/lb dry air [15 grams of water/kg dry air]
    Indoor conditions: Ti = 78ºF [25.6ºC], Wi = 70 grains of water/lb dry air [10 grams of water/kg dry air]
    QS = 1.085 x 360 x (95 – 78) = 6,640 Btu/hr
    [QS = 1,210 x 0.18 x (35 – 25.6) = 2047 W]
    QL = 0.7 x 360 x (105 – 70) = 8,820 Btu/hr
    [QL = 3,010 x 0.18 x (15 – 10) = 2,709 W]
  • System Heat Gains
    There may be others sources of heat gain within the HVAC system. One example is the heat generated by fans. When the supply fan, driven by an electric motor, is located in the conditioned airstream, it adds heat to the air. Heat gain from a fan is associated with three energy conversion losses.
    Fan motor heat is due to the energy lost in the conversion of electrical energy (energy input to the motor) to mechanical energy (rotation of the motor shaft). It is dissipated as heat from the motor and is represented by the inefficiency of the motor.
    fan motor heat gain = power input to motor  (1 – motor efficiency)
    If the fan motor is also located within the conditioned airstream, such as inside the cabinet of an air handler (as shown in Figure 51), it is considered an instantaneous heat gain to the airstream. If it is located outside the conditioned airstream, it is considered a heat gain to the space where it is located.
    Fan blade heat gain is due to the energy lost in the conversion of mechanical energy to kinetic energy (moving of the air). It is dissipated as heat from the fan blades, it is considered an instantaneous heat gain to the airstream, and it is represented by the inefficiency of the fan.
    fan blade heat gain = power input to fan  (1 – fan efficiency)
    Finally, the remaining (useful) energy input to the fan, the energy used to pressurize the supply duct system, is eventually converted to heat as the air travels through the ductwork. For simplicity, most designers assume that this heat gain occurs at a single point in the system, typically at the location of the fan.
    duct friction heat gain = power input to fan  fan efficiency
    Reference: 1997 ASHRAE Handbook–Fundamentals, Chapter 28, page 28.16
  • It is important to know where the fan heat gain occurs with respect to the cooling coil. If the fan is located upstream and blows air through the cooling coil, the fan heat causes an increase in the temperature of the air entering the coil.
    If, however, the fan is located downstream and draws air through the cooling coil, the fan heat causes an increase in the temperature of the air supplied to the space.
  • Another source of heat gain in the system may be heat that is transferred to the conditioned air through the walls of the supply and return ductwork. For example, if the supply ductwork is routed through an unconditioned space, such as a ceiling plenum or an attic, heat can be transferred from the air surrounding the duct to the supply air.
    Supply ductwork is generally insulated to prevent this heat gain and the associated increase in temperature of the supply air. An increased supply air temperature requires a greater amount of supply air to maintain the desired space conditions, resulting in more fan energy use. Insulation also reduces the risk of condensation on the cool, outer surfaces of the duct.
    Return ductwork, on the other hand, is generally not insulated unless it passes through a very warm space. Any heat picked up by the return air is generally heat that would have eventually entered the space as a cooling load. Therefore, the cooling load caused by this heat gain to the return air is not wasted.
    For the example used in this clinic, we will assume that fan heat gains and other system heat gains are negligible.
  • In summary, the total cooling load for our example space is made up of the following components:
    Conduction heat gain from outdoors through the roof and west-facing exterior wall and windows
    Solar radiation heat gain through the west-facing windows
    Internal heat gains from people, lights, office equipment, and a coffee maker in the space
    Heat gain due to hot, humid air infiltrating into the space from outdoors
    In addition, the cooling coil in the building HVAC system has to cool the outdoor air that is deliberately brought into the building for ventilation purposes.
    We will use these results in the next period to conduct a simplified psychrometric analysis of our example space.
  • In this period we will use the results from our example cooling load calculations to 1) determine the sensible heat ratio for the space, 2) perform a simplified psychrometric analysis to calculate the supply airflow and temperature required to properly condition the space, and 3) calculate the required capacity of the cooling coil.
    Then we will analyze a system that serves multiple spaces to determine the design airflow for the supply fan and the total building cooling load.
  • Single-Space Analysis
    The first step in our psychrometric analysis is to determine which components of the cooling load are space loads and which only affect the coil load. This is important because, although all heat gains that occur inside the building contribute to the total load on the cooling coil, only those heat gains that occur within the space need to be offset by the cool air supplied to the space. Notice that all space loads are also coil loads, but all coil loads are not necessarily also space loads.
    For example, in most buildings, ventilation air is conditioned prior to being delivered to the space. Therefore, the ventilation load adds to the total cooling coil load, but does not add to the space cooling load. Additionally, heat gains that occur within the HVAC system, such as fan heat and duct heat gain, are considered coil loads, but not space loads.
  • This table lists only the space cooling loads calculated for the example used in Period Two. The total sensible cooling load for this space is 74,626 Btu/hr [21,623 W] and the total latent cooling load for this space is 9,109 Btu/hr [2,599 W].
  • The proportions of sensible and latent heat must be known in order to determine the proper condition of the air being supplied to cool the space. This sensible heat ratio (SHR) is the ratio of sensible heat gain to total (sensible plus latent) heat gain, and is defined as follows:
    The SHR for our example space is 0.89. That is, 89% of the cooling load for this space is sensible and 11% is latent.
  • After the SHR has been determined for the space, a simple psychrometric analysis can be performed to determine the quantity of air that must be supplied to condition this space, and the proper temperature of that air. The next few illustrations will use some of the basic concepts presented in the Psychrometry Air Conditioning Clinic.
    This analysis assumes that the example space is served by its own dedicated air-conditioning system, consisting of a cooling coil and supply fan.
  • The quantity of air required to offset the space sensible heat gain is determined using the following formula:
    where,
    Sensible heat gain = sensible heat gain in the space, Btu/hr [W]
    1.085 [1,210] = product of density and specific heat, Btu•min/hr•ft3•ºF [J/m3•ºK]
    Supply airflow = quantity of air supplied to the space, cfm [m3/s]
    Room DB = desired space dry-bulb temperature, ºF [ºC]
    Supply DB = supply air dry-bulb temperature, ºF [ºC]
    Remember that 1.085 [1,210] is not a constant—it is derived from the density and specific heat of the air at actual conditions.
  • The next step is to either assume the supply air dry-bulb temperature and calculate the supply airflow, or assume the supply airflow and calculate the supply air temperature.
    For our example, we will assume that the supply air dry-bulb temperature is 55ºF [12.8ºC]. Based on this assumption, the quantity of air required to offset the sensible heat gain in this space is 2,990 cfm [1.40 m3/s].
    The rest of the analysis will fine-tune this assumption.
  • Next, we need to calculate the condition of the air entering the cooling coil. This air is a mixture of return air (RA) from the space and outdoor air (OA). In Period Two, we determined that 360 cfm [0.18 m3/s] of outdoor air is required to properly ventilate this space.
    The percentage of the total supply airflow that is made up of outdoor air is determined as follows:
    This indicates that 12% of the air entering the cooling coil is outdoor air and 88% is air being recirculated from the space.
  • Assuming that the air being recirculated from the space is the same condition as the space, we can determine the condition of the air entering the cooling coil. First, the dry-bulb temperature of this air mixture is determined as follows:
    Outdoor air conditions: 95ºF [35ºC] dry bulb, 76ºF [24.4ºC] wet bulb
    Recirculated air conditions: 78ºF [25.6ºC] dry bulb, 50% relative humidity
    (95ºF  0.12) + (78ºF  0.88) = 80ºF
    [(35ºC  0.12) + (25.6ºC  0.88) = 26.7ºC]
    Using the psychrometric chart, the condition of this air mixture (C) must fall on a line connecting the condition of the recirculated air (A) and the condition of the outdoor air (B). The wet-bulb temperature that marks the intersection of the connecting line and the 80°F [26.7°C] dry-bulb temperature mark is approximately 66.5°F [19.2°C]. Because the recirculated air constitutes a larger percentage (88%) of the mixture, the mixed-air condition (C) is much closer to the recirculated air condition (A) than the outdoor design condition (B).
  • The next step is to determine the supply-air condition (dry-bulb and wet-bulb temperatures) necessary to absorb the sensible and latent heat in the space.
    A sensible-heat-ratio line is drawn by connecting the 0.89 value on the SHR scale with the index point. Since the index point is the same as the desired space condition for this example (A), this line is extended until it intersects the saturation curve. If the desired space condition was different, a line would be drawn parallel to the 0.89 SHR line through the space condition.
    Using the curvature of the nearest two coil curves as a guide, draw a curve from the mixed-air condition (C) until it intersects the SHR line. This point of intersection (D) represents the supply-air condition that will offset the space sensible and latent heat gains in the correct proportions required to maintain the desired space condition. Here, this supply-air condition is 59°F dry bulb, 57.4°F wet bulb [15°C dry bulb, 14.1°C wet bulb].
  • Because this resulting supply air temperature is different than the 55ºF [12.8ºC] that we assumed, we need to recalculate the supply airflow.
    To complete another iteration of this analysis, we would 1) calculate the new percentage of outdoor air (10%), 2) determine the new entering coil conditions (79.7ºF dry bulb, 66.2ºF wet bulb [26.5ºC dry bulb, 19ºC wet bulb]), and 3) draw a curve from this new mixed-air condition until it intersects the 0.89 SHR line. This second iteration will likely result in the same supply air condition as the first iteration.
    This simple psychrometric analysis indicates that our example space requires 3,620 cfm of air at 59°F dry bulb, 57.4°F wet bulb [1.69 m3/s of air at 15°C dry bulb, 14.1°C wet bulb] to offset the sensible and latent heat gains for the space in the correct proportions.
    For more information on the process of analyzing an HVAC system on the psychrometric chart, refer to the Psychrometry Air Conditioning Clinic.
  • Finally, the cooling coil in the air-conditioning system that serves this space must be capable of handling the sensible and latent loads for the space, plus any additional loads that affect the coil only. In our example, outdoor air for ventilation is the only additional load affecting the cooling coil. Based on the calculations performed in Period Two, the total load on the cooling coil is 99,195 Btu/hr (8.3 refrigeration tons) [28,978 W].
  • Multiple-Space Analysis
    So far in this clinic, we have calculated the cooling load for a single space and learned how to size the fan and cooling coil for a system serving that space. Referring back to the example office space in Figure 15, now we will consider an example where rooms 101 (west-facing wall) and 102 (east-facing wall) are served by the same air-conditioning system.
    As shown in the previous section, if each space is conditioned by a separate system, then the fan and coil would be sized to handle the maximum load for the particular space. If, however, a single system is used to condition several spaces in a building, the method used to size the fan and coil depends on whether the system is a constant-volume (CV) or variable-air-volume (VAV) system. The example system shown here has two spaces. If the supply fan delivers a constant volume of air, the fan must be sized by summing the peak sensible loads for each of the spaces it serves. If, however, it is a VAV system and the fan delivers a varying amount of air to the system, the fan is sized based on the one-time, worst-case airflow requirement of all of the spaces it serves.
    The next example will help explain this important difference.
  • Room 101 is the same space that we have used in the two previous periods of this clinic. In Period Two, we calculated the components of the space cooling load at 4 p.m. This table shows the components of the space sensible load at both 8 a.m. and 4 p.m. (The cooling loads at 8 a.m. were calculated using the same methods introduced in Period Two.)
    Because room 101 has several west-facing windows, the peak (highest) space sensible load occurs in the late afternoon when the sun is shining directly through the windows.
  • Room 102 is exactly the same as room 101, except that the exterior wall faces east instead of west. Because of this difference, these two rooms will not only have different cooling loads, but the peak space cooling load for room 102 will occur at a different time of day than for room 101.
    Because room 102 has several east-facing windows, the peak space cooling load occurs in the morning when the rising sun shines directly through the windows.
  • As seen in this example, the peak space loads do not necessarily occur at the same time for all spaces served by the same system. As discussed earlier in this period, the supply airflow for a space is calculated based on the maximum sensible load for that space—74,626 Btu/hr [21,623 W] for room 101 and 62,414 Btu/hr [18,087 W] for room 102. If the supply fan in this example delivers a constant volume of air at all times, it must be sized by summing the peak sensible loads of both spaces—137,040 Btu/hr [39,710 W]. This is called the sum-of-peaks load.
    Although rooms 101 and 102 peak at different times of the day, there will be a single instance in time when the sum of these two space loads is highest. This is called the block load. If these two spaces are served by a single VAV system, in which the supply fan delivers a varying amount of air to the system, the fan only needs to be sized for the time when the sum of the space sensible loads is the highest—129,939 Btu/hr [37,781 W].
  • Assume that the supply air dry bulb is the 59°F [15°C] that was calculated during the psychrometric analysis. The sum-of-peaks and block airflows for sizing the supply fan in these two cases can then be calculated as follows:
    This is the reason that VAV systems can use smaller supply fans than constant-volume systems.
  • Similar to the single-space example, the cooling coil in the HVAC system serving these two spaces must be capable of handling the space sensible and latent loads, plus any additional loads that only affect the coil. Again, in our example, outdoor air for ventilation is the only additional load affecting the coil.
    In general, a cooling coil in a system serving multiple spaces is sized based on the block cooling load. In our example, this block load occurs at 4 p.m., the time when the sum of the space loads for rooms 101 and 102, plus the ventilation load, is the highest. Based on the calculations performed in Period Two and earlier in this period, the block load for sizing the cooling coil in this multiple-space system is 179,077 Btu/hr (14.9 refrigeration tons) [52,491 W].
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